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Distinguished Guests,

On behalf of the Organizing Committee and the Scientific Committee, I would like to welcome you to Istanbul for the XII. International HVAC+R Technology Symposium. This three day Symposium coupled with Boat Cruise and Gala Dinner, will provide an efficient and at the same time, a productive knowledge exchange platform for scientists and engineers from all over the world.

Besides the opening lecture with Bjarne W. Olesen from Technical University of Denmark; highlighting the symposium, there are eleven Keynote Lectures given by Wei Sun, Engsysco Inc., Marija S. Todorovic University of Belgrade, Eckhard A. Groll Purdue University, Michael M. Ohadi University of Maryland, Nejat Babür Anel Group, Formerly; STV Inc., Ch2M HILL, US Department of Energy, S.A. Sherif University of Florida, William M. Worek Texas A&M University, Branislav B. Todorovic University of Belgrade.

There will also be three panel discussions, namely, “Turkish HVAC+R Industry BIM (Building Information Modeling) Strategy”, “Widen Your World For Efficient Solutions: The Importance of Thermodynamics” and “Design and Planning in Achieving Successful Buildings”.

The Scientific Committee of the Symposium consists of twenty-eight members from different well known universities in Turkey.

After a very intense evaluation and collection processes fifty-seven papers were selected to be presented orally in fifteen sessions. Abstracts of all presented papers are included in the Programme/Abstract Book and all the full papers can be observed in the Proceedings section of the USB provided for the participants.

I would like to take this opportunity to thank Turkish Society of HVAC and Sanitary Engineers (TTMD) for organizing HVAC symposiums in a sustainable manner since 1994, Keynote Speakers for their valuable contributions, Authors for submitting their work, the Scientific Committee members for taking time and reviewing papers, secretariat Etix Inc. and Organizing Committee members for making XII. International HVAC+R Technology Symposium a very highly successful conference, both technically and socially.

I wish all participants a pleasant week during the symposium and for our international guests; I wish you a favorable stay in Istanbul and in Turkey. I sincerely hope this will be a very productive and effective symposium for all of our participants.

Prof. Dr. A. Nilüfer Egrican

President of Executive, Organizing and Scientific Board

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Distinguished Guests,

It is our great pleasure to invite you to, organized biannually, XII. International HVAC+R and Sanitary Technology Symposium the leading international scientific and engineering Symposium in the field of HVAC&R (Heating, Ventilating, Air Conditioning and Refrigeration).

Heating and airconditioning are at the center of energy consumption in our world today. As the living standards improve and the demand for comfort increases, airconditioning is becoming indispensable. The need to find a sustainable solution for growing energy use has become ever more important. Achieving sustainable energy consumption is one of the major goals of developed societies around the world. Likewise, research on buildings for tomorrow has become increasingly popular and highly supported by many governments.

The XII. International HVAC+R and Sanitary Technology Symposium will address “Widen Your World for the Efficient Solutions” as its main theme. The Symposium has received 84 abstracts from 11 countries.

In addition to inviting world renowned speakers at the Symposium, we organized 3 panels titled “Turkish HVAC+R Industry BIM (Building Information Modeling) Strategy”, “Widen Your World For Efficient Solutions: The Importance of Thermodynamics” and “Design and Planning in Achieving Successful Buildings” where experts from different countries will share research expertise to broaden the scope of discussions.

On behalf of the Organizing Committee, I warmly invite you to join in Istanbul for the XII. International HVAC+R and Sanitary Technology Symposium, the legendary capital of the Byzantine, Roman, and Ottoman Empires and the only city to connect two continents, Istanbul is among the world’s largest cosmopolitan centers. A beautiful city with a mild Mediterranean climate, Istanbul will provide the ideal backdrop for the Symposium attendees.

Sarven Çilingiroğlu,

President of TTMD and Symposium Chair

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Symposium Chair Sarven Çilingiroğlu

Executive Board and Organizing Committee

Prof. Dr. Nilüfer Eğrican (President)

* Are arranged in alphabetical order by name.

Bahri Türkmen Doç. Dr. Derya Burcu Özkan

Nermin Köroğlu Isın Doç. Dr. Özden Ağra Sarven Çilingiroğlu

Seçil Kızanlık İskender

Scientific Board Prof. Dr. Nilüfer Eğrican (President)

* Are arranged in alphabetical order by name.

Prof. Dr. Abdurahman Kılıç

Prof. Dr. Ahmet Arısoy Prof. Dr. Ali Güngör

Prof. Dr. Arif Hepbaşlı Prof. Dr. Birol Kılkış

Prof. Branislav B. Todorovic Prof.Dr. Bülent Yeşilata

Doç. Dr. Derya Burcu Özkan Prof. Dr. Gönül Utkutuğ Prof. Dr. Gülden Gökçen Prof. Dr. Halime Paksoy

Prof. Dr. Hasan Heperkan Doç. Dr. Hatice Sözer

Prof. Dr. İlhan Tekin Öztürk Prof. Marija S. Todorovic Prof. Michael M. Ohadi

Dr. Murat Çakan Doç. Dr. Murat Kadri Aktaş

Nejat Babür Doç. Dr. Özden Ağra Prof. Dr. Sadık Kakaç

Dr. Tahir Yavuz Prof. Dr. Vahan Kalenderoğlu

Prof. Dr. Yunus Ali Çengel Prof. Dr. Zerrin Yılmaz

Wei Sun

Advisory Board Celal Okutan

Kevork Çilingiroğlu Akdeniz Hiçsönmez

B.Erdinç Boz Numan Şahin Serdar Gürel

Ömer Kantaroğlu Engin Kenber

Hüseyin Erdem Abdullah Bilgin

Cafer Ünlü Gürkan Arı

Bahri Türkmen

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PANEL – I

Turkish HVAC+R Industry BIM (Building Information Modeling) Strategy Moderator : Ozan Atasoy - TTMD Member ISKID Secretary General Panelists : Daniel Kazado - BIM Consultant ProCS Professional Construction Solutions Taner Yönet - ISKID Vice Chairman Önder Boyalıklı - MTMD Board of Directors Member Mehmet Oskay - Moskay Engineering & Contracting Ltd. Co. Berke Çelikel - ARUP Turkey

Utku Başyazıcı - DOXA87 Melike Altınışık - Melike Altinisik Architects

Abstract

Building Information Modeling (BIM) is an intelligent 3D model-based process that equips architecture, engineering, and construction professionals with the insight and tools to more efficiently plan, design, construct, and manage buildings and infrastructure.

BIM is a fast developing application in especially US, UK and Western European countries. It is getting more common in especially international projects and started to be requested in some big projects in Turkey. It is expected that in near future we will see BIM application more common.

In this panel BIM strategy of Turkish Hvac+R industry will be discussed to find out the influence of BIM to the industry. Panelists are selected from the building design, construction and material suppliers to share their opinion about influence on their business and what they will need to manage this transformation to BIM.

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PANEL - II

WIDEN YOUR WORLD FOR EFFICIENT SOLUTIONS: The Importance of Thermodynamics

Moderator : Prof.Dr.Sadık Kakaç - Emeritus, University of Miami, USA, TOBB University of Economics and Technology Panelists : Prof.Dr.Yunus Ali Çengel - Adnan Menderes University (Also University of Nevada, Reno, NV, USA) Prof.Dr.Birol Kılkış - Baskent University Dr.Cemil İnan - Product Director of Arçelik (Global Operations of Refrigerator and Compressor)

Abstract

In this panel, it will be showen how to today’s science of thermodynamics evolved starting from the Sadi Carnot’s heat engine studies in which heat was seen as a wieghtless fluids called “caloric”. Thermodynamics is a basic science that has been essential part of Engineering curricula. Therefore some techniques in teaching that will make the thermodynamics experience of students a more pleasent and fruitful one relating the subject matter to real applications and experience are explained; in addition, the role of thermodynamics in today’s indusrial Research agenda with some examples and cases from Industrial R&D are also given.

In this panel, exergy dimension to zero-enegy building concept and its importance is also discussed; net zero-exergy, near –zero and low-exergy building concepts are defined, and differences from the laws of science of thermodynamics are indicated.

As a practical application to the analysis of heat exchanger design, an experimental set up has been designed and built to test the performance of gasketed-plate heat exchangers. The analysis is the application of the first law of thermodynamics for an open system under the steady-state conditions.

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PANEL - III

Design and Planning in Achieving Successful Buildings

Moderator : Can Çinici - Cinici Architects Panelists : Mehmet Okutan - Okutan Engineering

Burak Ünder - Under Architects Serdar Binzet - Renaissance Technical Construction

Industry and Commerce Mehmet Karadurak - Aykar Engineering

Abstract

Mainstream building practice presupposes technical consultancies such as civil, mechanical, infrastructural and electrical engineering besides architecture as separate fields of project making, keping in mind that these are all the different aspects of the same building activity. During the last two decades, green concerns have given rise to sustainability consultancies also besides others The panel focuses on the interdisiplinary nature of the building activity, most common traps the current practice faces as well as future possibilites in developing a more enhanced physical environment.

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[ABSTRACT:001] p.14 HIGH EFFICIENCY ELECTRONIC SUPER HEAT CONTROLS IN COOLING PLANTS AND HEAT PUMPS Mehmet Ümit Sunar, Uwe Knief, Matthias Erlewein [ABSTRACT:004] p.18 AN ECONOMIC TOOL FOR COGENERATION SYSTEMS Birol Kılkış [ABSTRACT:008] p.20 COMPARISON OF 2D/3D NUMERICAL ANALYSIS OF THE AIR SIDE HEAT TRANSFER AND LAMINAR FLOW IN MULTILOUVERED FIN AND FLAT TUBE HEAT EXCHANGER Nihal Uğurlubilek, Latife Berrin Erbay [ABSTRACT:009] p.28 NEXT GENERATION REFRIGERANT R-32 AND THE LATEST DEVELOPMENTS ABOUT THE DISSEMINATION OF R-32 AROUND THE WORLD Andaç Yakut [ABSTRACT:011] p.33 SCOPE AND COMPARISON OF NORMS REGULATING THE REQUIREMENTS FOR HYGIENIC AIR HANDLING UNIT DESIGN Orkun Yılmaz, Hamza Sonkur [ABSTRACT:012] p.44 URBAN TRANSFORMATION AND QUALITY OF LIFE IN BURSA DOĞANBEY Miray Gür, Neslihan Dostoğlu [ABSTRACT:013] p.50 ENERGY SAVING APPLICATIONS IN FLUE GAS INSTALLATIONS Muammer Akgün [ABSTRACT:015] p.55 DETERMINING THE OPTIMUM FINANCIAL AND ENERGY SAVING SOLUTIONS FOR BUILDINGS IN DIFFERENT CLIMATES Aslıhan Şenel Solmaz [ABSTRACT:017] p.62 ENERGY EFFICIENT CLEANROOM DESIGN Nejat Babur [ABSTRACT:018] p.70 THERMODYNAMIC ANALYSIS OF GORUND SOURCE HEAT PUMP Fatih Yılmaz, Mustafa Tolgaz Balta, Reşat Selbaş [ABSTRACT:021] p.76 THE COMPARISON OF DIFFERENT AIR-CONDITIONING SYSTEMS IN AN INDUSTRIAL COMPANY İlhami Horuz, Nazım Kurtulmuş [ABSTRACT:022] p.82 SAP2000 ANALYSIS OF STEEL SUPPORTS USED FOR SUSPENDED MEP SERVICES Alp Yucerman, Merve Turkoglu, Anil Cagatay

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[ABSTRACT:023] p.88 ANALYSIS OF THE CHIMNEY RELATED CARBON MONOXIDE POISONING Muammer Akgün [ABSTRACT:024] p.95 CUI – CORROSION UNDER INSULATION Fulya Hamidiye [ABSTRACT:026] p.98 EXPERIMENTAL INVESTIGATION OF OPTIMUM RADIANT TO CONVECTIVE HEAT TRANSFER SPLIT RATIO FOR HYBRID HVAC SYSTEMS Mustafa Fatih Evren, Abuzer Kadir Özsunar, Birol Kılkış [ABSTRACT:028] p.105 AN EXERGY BASED ANALYSIS OF END USE FOR ELECTRICITY AND NATURAL GAS MARKET Nezih Enes Evren [ABSTRACT:030] p.112 STUDY OF DIFFERENT SCENARIOS OF BUILDING EXTERIOR INSULATION USING EQUEST Abdellah Zerroug, Egils Dzelzitis [ABSTRACT:032] p.118 INNOVATIVE ANTIBACTERIAL NANO ADDITIVE SOUND AND NOISE INSULATION USAGE AT HOSPTIAL'S AIR CONDITIONING DUCTS' INSIDE INSULATION Cudi Volkan Dikmen [ABSTRACT:033] p.124 AN INVESTIGATION OF DOUBLE EFFECT ABSORPTION REFRIGERATION SYSTEM BASED ON DOUBLE STAGE HEAT SOURCE Kenan Saka, İbrahim Halil Yılmaz, Ömer Kaynaklı [ABSTRACT:034] p.131 HEAT STRESS, THERMAL EXPOSURE AND COMFORT, AMBIENT NOISE AND LEVEL OF ILLUMINATION ON WORKPLACES; ANALYSIS OF MEASUREMENT TECHNIQUES WITHIN THE SCOPE OF WORKPLACE SAFETY Şükrü Onur Yiğit [ABSTRACT:043] p.143 ENERGY EFFICIENT RETROFITTING SCENARIOS COMPARISON IN TERMS OF BUILDING FUNCTION Kemal Ferit Çetintaş, Zerrin Yılmaz [ABSTRACT:046] p.150 ENERGY MANAGEMENT THROUGH DYNAMIC MONITORING OF AN EDUCATIONAL BUILDING AT THE UNIVERSITY OF VALLADOLID Tejero González Ana, Francisco Javier Rey Martínez, Eloy Velasco Gómez,Luis Manuel Navas Gracia, González Sergio Lorenzo, Andrés Chicote Manuel, Javier María Rey Hernández, Samuel De La Fuente [ABSTRACT:047] p.156 INVESTIGATION OF RESISTANT WIRES’ RUPTURE PHENOMENON USED IN FUEL-OIL BURNER PRE-HEATERS BY COMPARISON OF ANALYTICAL AND NUMERICAL RESULTS Barış Elbüken

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[ABSTRACT:054] p.168 INVESTIGATION OF THE EFFECT OF AIRFLOW MALDISTRIBUTION ON EVAPORATOR THERMAL PERFORMANCE Ergin Bayrak, Alper Şevki Konukman [ABSTRACT:055] p.176 GREEN BUILDINGS IN INSULATION MATERIAL PRODUCTION Ayhan Gökbağ [ABSTRACT:056] p.181 HEAT TRANSFER MEASUREMENTS FOR NON-DARCY FLOW IN 10-PPI METAL FOAM Özer Bağcı, Nihad Dukhan, Altay Arbak [ABSTRACT:057] p.187 IMPORTANCE OF CLIMATE FOR DESIGNING OF A ZERO ENERGY BUILDING Ahmet Arısoy, Burcu Sağlam, Burhan Yörük [ABSTRACT:059] p.196 INVESTIGATION OF THERMAL COMFORT IN A CLASSROOM IMPROVED WITH HEAT RECOVERY VENTILATION SYSTEM İbrahim Atmaca, Orhan Ekran, Sait C. Sofuoğlu, Z. Haktan Karadeniz, Macit Toksoy [ABSTRACT:064] p.202 MULTIPLE CRITERIA DECISION FOR INTEGRATED BUILDINGS Metin Selcuk Ercan [ABSTRACT:066] p.210 FIRST LAW ANALYSIS OF SINGLE STAGE ABSORPTION HEATING SYSTEM USING LIBR-H2O, H2O-NH3, NASCN-NH3 AND LINO3-NH3 Veysel Ergül, Yunus Emre Talu, Hakan Demir, Şevket Özgür Atayılmaz [ABSTRACT:067] p.218 STATE OF THE ART OF HVAC TECHNOLOGY IN EUROPE AND AMERICA Ongun Berk Kazanci, Bjarne W. Olesen [ABSTRACT:071] p.223 ENERGY EFFICENCY APPROACHES IN DESIGN OF METRO STATIONS İbrahim Ethem Özbakır, Levent Tosun [ABSTRACT:072] p.229 THE GENERATION OF TYPICAL METEOROLOGICAL YEAR AND CLIMATIC DATABASE OF TURKEY FOR THE ENERGY ANALYSIS OF BUILDINGS Serpil Yılmaz, Mustafa Yılmaz, İsmail Ekmekçi [ABSTRACT:073] p.235 AN EXERGY-BASED AUTOMATION SYSTEM IN ESER LEED PLATINUM BUILDING Birol Kılkış, Ayşe Gülbeden [ABSTRACT:074] p.243 ENERGY MANAGEMENT IN HOSPITALS Mustafa Zeki Yılmazoğlu

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[ABSTRACT:075] p.248 ENERGY AND EXERGY ANALYSIS OF WATER AND AIR COOLED PVT SYSTEMS WITH HEAT PIPE TECHNOLOGY Birol Kılkış, Tuğberk Ozar [ABSTRACT:076] p.264 POTENTIAL FOR CO2 EMISSION REDUCTION AT SETTLEMENT SCALE THROUGH COST OPTIMAL AND NEARLY ZERO ENERGY BUILDINGS Ece Kalaycıoğlu, Touraj Ashrafian, Nazanin Moazzen, Ayşe Zerrin Yılmaz [ABSTRACT:077] p.270 COMPARISON OF VACUUM COOLING AND CONVENTIONAL COOLING OF PARSLEY, COOKED POTATO AND LEEK Hande Mutlu Öztürk, Günnur Koçar, Harun Kemal Öztürk [ABSTRACT:080] p.297 ENERGY EFFICIENT ROOM AUTOMATION PRACTICES BASED ON EN15232 AND VDI3813 Cristóbal Fernández, Mustafa Değirmenci [ABSTRACT:081] p.303 THE ISO AND CEN STANDARS ON ENERGY PERFORMANCE OF BUILDINGS ASSESMENT PROCEDURES, ALLOWING MAXIMAL FLEXIBILITY AND TRANSPARENCY, SUPPORTING SMART BUILDING CONCEPTS Jaap Hogeling [ABSTRACT:087] p.318 ENERGY AND COST ANAYSIS OF THE ALTERNATIVE HVAC SYSTEM APPLICATIONS FOR GREEN BUILDING APPROACH Ünal Sınar, Hale Tuğçin Kırant, Ebru Mançuhan, Barış Yılmaz, Mustafa Kemal Sevindir [ABSTRACT:091] p.326 THE NUMERICAL ANALYSIS OF THERMAK CONDUCTIVITY OF AUTOCLAVED AERATED CONCRETE USED IN BUILDINGS Hüsamettin Tan, Battal Doğan [ABSTRACT:095] p.334 EFFECT OF ICE THERMAL ENERGY STORAGE SYSTEM ON COOLING COST IN A SHOPPING CENTER Doğan Erdemir, Necdet Altuntop [ABSTRACT:096] p.341 PREDICTION OF SOIL TEMPERATURES FOR UNDERGROUND HEAT EXCHANGER APPLICATIONS IN IZMIR, TURKEY Deniz Yener, Önder Özgener, Leyla Özgener [ABSTRACT:097] p.344 EFFECT OF THE HFC AND HC REFRIGERANTS AS SECONDARY WORKING FLUID ON PERFORMANCE OF BINARY GEOTHERMAL POWER PLANT Leyla Özgener, Anıl Başaran [ABSTRACT:098] p.351 ANALYZING AUTOMATION SYSTEMS WHICH IS USED FOR HEATING-COOLING THE BULDINGS, IN TERMS OF COMFORT AND ENERGY EFFICIENCY Semiha Öztuna, Seyhan Özkan

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[ABSTRACT:099] p.357 ESTIMATING OF THE EFFECT OF COOL ROOFS ON ENERGY SAVINGS IN VARIOUS CLIMATE ZONES Murat Bulut, Gul Eroglu Bulut, Nedim Sozbir [ABSTRACT:100] p.363 EVALUATING THE LCA OF A HOUSE WITH THREE DIFFERENT WALLS Pooya Pakmehr, Mustafa Erkan Karagüler [ABSTRACT:101] p.370 THERMODYNAMIC ANALYSIS OF CENTRAL HEATING SYSTEMS AND INVESTIGATION EFFECTS OF EXERGY EFFICIENCY OF TEMPERATURE CHANGING IN AIR BURNING Zafer Utlu, Mesut Yenıgün [ABSTRACT:102] p.378 EXPERIMENTAL INVESTIGATION OF THE EFFECTS OF USING EVAPORATIVE CONDENSER IN COOLING SYSTEMS Abdulaziz Yıldız, Ali Etem Gürel, Emrah Deniz [ABSTRACT:103] p.384 SELECTION OF GRILLERS AND DIFFUSERS FOR COMFORT APPLICATIONS Hüseyin Bulgurcu, Bekir Cansevdi [ABSTRACT:106] p.395 THE OBSERVATION OF TARGETS ACHIEVED DURING DISTRICT HEATING DEVELOPMENT IN RIGA CITY Egils Dzelzitis [ABSTRACT:107] p.401 NEXT GENERATION BUILDING AUTOMATION TECHNOLOGIES Jagdish Naik

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[Abstract:0001][Comfort Cooling] ADAPTIVE ELECTRONIC SUPER HEAT CONTROL FOR COOLING AND

HEAT PUMP SYSTEMS

Dipl.-Ing (FH) Uwe Knief Dipl.-Ing. (BA) Matthias Erlewein

Honeywell GmbH, Cooling Solutions Hardhofweg, D-74821 Mosbach

Mechanical Engineer (M.Sc.) Umit Sunar

Honeywell Technology A.S. Içerenkoy - Istanbul

Abstract Refrigeration and heat pump systems are used increasingly in advanced fields of applications. To optimize the performance for such systems, frequency controlled compressors, pumps and fans are used. This leads to a much wider operating range of the expansion valves. Thermostatic Expansion Valves are limited in their operating range and are not able to fulfill the extended operating ranges in some applications. In these applications Electronic Expansion Valves with extended operating ranges are used. However, the efficient operation of an Electronic Expansion Valve also depends on the electronic superheat controllers used. A new self adapting superheat controller automatically identifies the stability limits of the evaporator and adapts to the learned characteristics. This leads to the result that the minimum necessary superheat for a stable operation of the cooling system is used under all conditions. The characteristics and limits of Thermostatic Expansion Valves are discussed and the overall performance of the combination of Superheat Controllers together with Electronic Expansion Valves is compared. Key words: Super Heat, Expansion Valve 1. Introduction Heat pump systems and cooling- and refrigeration systems increasingly operate under wide application ranges. Frequency controlled compressors, pumps and fans extend the operating range of the expansion devices. Thermostatic Expansion Valves (TXV) are limited to a capacity ratio of about 1 to 4 and are not able to follow the requirements of the system under some operating conditions. Electronic Expansion Valves (EEV), with a modulation ratio of about 1 to 10, helps to solve this issue. However the resulting performance is greatly dependent upon the Electronic Superheat controller (SHC) used to drive the EEV. Optimized energy efficient operation of heat pump, cooling and refrigeration systems will become more important in future. The drivers are costs and legislations in different countries. 2. Problem How can we design more efficient cooling and heat pump systems? Theoretically it is quite simple as we see in picture 1. In the cooling system heat is transferred from a low temperature level to a higher temperature level. The cooling system uses the low end of the temperature and the heat pump uses the high end of the temperature. In both cases the main conditions are given by the temperature of the source tN- and by the temperature of the sink tN+. The main target for an optimized design of a cooling system is to minimize the temperature difference ΔtKM between condensing temperature tc and evaporating temperature to. This minimizes the pressure difference between condensing pressure pc and evaporating pressure po which is the operating pressure difference of the compressor. The lower the pressure difference for the compressor the better the performance of the system.

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Picture 1: Temperature differences in a cooling system [Küpper 10] A reduction of the condensing temperature by 1 K increases the COP by 2 to 3 %. Whereas 1 K higher evaporating temperature increases the COP by 3 to 4 %. 3. Superheat in Evaporators Picture 2 shows the temperatures inside an evaporator working in counter current flow operation. For heat transfer a temperature difference between the source, for example water or air, and the evaporating temperature is needed. This is particularly relevant in the area where the superheated refrigerant to2h (Suction line temperature) leaves the evaporator. It is obvious that with given source temperatures (tQein–tQaus) the performance of the system (compressor) increases with increased evaporating temperature. To control the superheat dtoh at a low level the evaporating temperature to can be on a higher level. Now it must be guaranteed that no liquid refrigerant, which could damage the compressor, leaves the evaporator. But each evaporator also has a given Minimum Stable Superheat MSS characteristic. We can say that with an increase of the cooling capacity the necessary superheat will also increase. If the superheat is too low, the system becomes instable and the total performance decreases. If the superheat is controlled on a too high stable level the evaporating temperature will decrease and again the performance of the system decreases. The overall target for a SHC together with an EEV is now to control the superheat under all conditions on a stable level that is as low as possible. This leads to the best performance of a cooling or heat pump system.

Picture 2: Temperature profiles in an evaporator [Osthues 03] 4. Thermostatic Expansion Valves TXVs have been used successfully for many years to control the superheat in evaporators. TXVs (with external equalization) measure the saturated pressure and the temperature of the superheated refrigerant at the outlet of the evaporator and control the superheat by changing the mass flow of refrigerant into the evaporator. TXVs have a linear characteristic; this means that the mass flow is proportional to the superheat. In addition TXVs need a static superheat setting to protect the compressor under all operating conditions.

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Picture 3 shows an ideal superheat setting of the TXV. The characteristic of the TXV is tangential to the MSS characteristic of the evaporator. However, in real systems TXVs need a higher superheat setting than shown in the diagram. Manufacturers optimize the characteristic of these valves for specific appliance applications, but this cannot avoid that, under extreme conditions, the difference between the valve characteristic and the MSS increases. This additional superheat leads to loss of efficiency of the system. TXVs are driven mechanically by pressure changes in the thermo system. The modulation ratio is limited to about 1:4. The main reasons are the limited pressure differences in the thermo system and the limited stroke of the valve. For some applications this limited modulation ratio leads to the use of multiple TXVs. A typical example is an air/water heat pump designed for operation in both heating and in cooling mode.

Picture 3: MSS and characteristic of a TXV [Küpper 10] 5. Electronic Superheat Control Electronic Expansion Valves (EEV) can reach a modulation ratio of 1:10 or more. The reason is that EEVs are driven by stepper motors or by solenoids. This makes the opening independent of the refrigerant used and it´s pressure level. EEVs are driven by electronic superheat controllers. Usually, as for TXVs, the suction line pressure and the temperature of the superheated refrigerant at the evaporator outlet are used. The saturated temperature at the dew point is calculated from the pressure. The superheat dtoh is now the difference between the suction line temperature and the calculated evaporating temperature to. For evaporators with limited pressure losses on the refrigerant side the temperature at the evaporator inlet can be used to represent the evaporating temperature. A non adaptive SHC uses a fixed set point for the superheat. The controller tries to keep the set value. If, at a given load, the set point is close to the MSS value of the evaporator then the control behavior is fine. If the set point however is higher than the MSS for the specific operating capacity then the superheat is too high and the system does not run under optimal conditions. If the set point is too low for the specific operating capacity then the system becomes instable and starts to oscillate. As a result liquid refrigerant can enter the compressor. In single evaporator systems the evaporating pressure starts to oscillate and the efficiency of the system decreases. To avoid this instability and risk of compressor damage the superheat is often set higher than needed. The superheat is usually set high enough to ensure that the appliance stays stable at the maximum cooling capacity. With this high superheat setting the performance of the system, with SHC control, can be lower than that of a correctly set-up thermostatic expansion valve. A simple adaptive superheat controller uses a set point for a fixed superheat at maximum load on the evaporator. This kind of adaptive superheat controller now tries to adapt to the MSS under each evaporator working condition. After working with a stable superheat for a certain time the controller reduces the superheat in small steps. As long as the suction line temperature is stable the controller tries to run with lower superheat. When the suction line temperature decreases and the superheat becomes

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instable the controller increases the superheat again. With this kind of super heat control it is guaranteed that no or nearly no liquid refrigerant will reach the compressor. On the other side the response time of standard evaporators is typically 2 minutes and it takes 30 minutes or more to reach the optimized setting for the given conditions. This means that after each restart the evaporator will operate with a superheat which is too high for a long time. This means that this kind of adaptive superheat controller is only useful for systems with long compressor running times. Current heat pumps, for example, often have up to 5-6 starts per hour under part load conditions. Under these conditions such controllers will not reach optimized operating conditions. 6. Knowledge based adaptive superheat controller A new knowledge based adaptive superheat controller is able to improve the performance of cooling refrigeration and heat pump systems. Based on a thermodynamic model for the controlled system (evaporator), the controller uses the evaporating pressure, suction line temperature and the degree of opening of the EEV to estimate cooling capacity. With this information the controller knows in which range of capacity the system is operating. As described in section 5 the controller adapts to the minimum possible superheat. The controller then writes the learned MSS value into a table. This is done for the different operating conditions (cooling capacities) of the cooling system. If the cooling capacity changes, or even when the system restarts after switching off the compressor, the controller immediately looks-up the learned MSS value from the table and operates with the learned optimized superheat. With this behavior the controller works under optimized conditions even when the system has short running times of the compressor. Picture 4 shows the superheat behavior for a TXV and an EEV with and without an adaptive superheat controller. The knowledge based adaptive controller tries to use the optimized superheat under all operating conditions. For reversible systems with hot gas-defrost or heating/cooling operation the controller is able to “learn” two different optimized MSS values. One for the “heating-” and another one for the “cooling-” conditions.

Picture 4: MSS and super heat [Küpper 10] 7. Conclusions and Outlook The self adapting knowledge based superheat controller allows continuous adaption of the superheat dtoh to the minimum possible values. This ensures a high COP of the cooling system. For the COP, for heat pump systems with a wide range of working conditions, can improve typically up to 15%, in some cases even higher. Future cooperation with OEM customers with focus on an optimized algorithm for defrost control leads to even more energy efficient systems. 8. References [Küpper 10] Küpper, H.-D.: Effiziente elektronische Überhitzungsregelung für Kälte- und Wärmepumpenregelung. DKV-

Jahrestagung Magdeburg, 2010 [Osthues 03] Osthues, J.; Stellan, B.: Erhöhung der Leistungszahl durch Plattenverdampfer mit Kältemittelverteiler und

abgestimmten Expansionsventil. Die Kälte & Klimatechnik (08/2003), S. 30-33 [Küpper 11] Küpper, H.-D.: Energy Efficient Electronic Superheat Controls for Heat Pumps.

European Heat Pump Summit Nuremberg, 2011 [Winter 12] Winter, J.; Handschuh, R.; Frei, G., Roth, P.: Die MSS-Theorie: Wie Verdampfer und Expansionsventil

zusammenwirken. Die Kälte Luft Klimatechnik (11/2012), S. 21-27

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[Abstract:0004][Energy Efficient Buildings] AN ECONOMIC ANALYSIS TOOL FOR TRIGENERATION SYSTEMS IN NET-ZERO

EXERGY BUILDINGS (NZEXB)

Birol Kılkış Fellow ASHRAE, Baskent University, Ankara, Turkey

Corresponding email: bkilkis@baskent.edu.tr SUMMARY This paper deals with a demand-simplified economic analysis tool for trigeneration systems in NZEXB buildings based on hourly predicted building loads. First, whether the concept of net-zero energy building or net- zero exergy building should be used is discussed in reference to district energy systems with new definitions. It has been quantified and shown that NZEXB concept should be used when both heat at different temperatures and power are exchanged among buildings and the district energy system. Next, a new fast-track hourly load prediction algorithm is introduced. This algorithm has two components. The first component is the prediction of hourly thermal loads, including loads for space heating and space cooling (including latent loads). The second component is the hourly prediction of electric power demand. Emphasis is given to commercial buildings, in particular to hospital buildings. Optimum sizing of tri-generation systems are discussed, including CO2 emissions. A case study is presented for the retrofit project of Turgut Özal Health Center at İnönü University in Malatya. INTRODUCTION Figure 1 shows typical hourly variation of thermal and power loads of a commercial building. According to this figure, peaks of different loads do not coincide and every type of load exhibits a different pattern. Consequently not only the magnitudes but different load proportions change almost instantly. Changes in load proportions are particularly important for cogeneration systems, because they supply heat and power almost at a constant ratio. This makes it necessary to bundle more than one energy conversion system and energy source, preferably utilizing different sustainable systems. It is evident that at the absence of detailed hourly load data it is quite impossible to size, design, and analyze the performance of the system bundle by using total or averaged load figures over a day, month, or a year [1]. In addition renewable energy resources are intermittent and most of the time they do not match with the loads. This makes it essential to extensively use TES (Thermal energy storage) systems and –if feasible to employ- electrical storage systems, besides grid exchange. One also needs to know the availability of renewable energy systems on an hour-by-hour basis. Keywords: Cogeneration, trigeneration, exergy-based economic analysis, hourly load prediction In addition, controls, systems and equipment respond to demands for cooling and heating in a building. Therefore, the factors that affect part-load performance of the equipment must be considered to yield accurate equipment energy use data [2]. In Figure 1, not only the magnitudes and proportions of loads change but at the same time demand temperatures (not shown) of thermal equipment and systems change with respect to the load magnitude. For example, while the outdoor temperature (in heating) decreases comfort heating loads increase and the heating equipment and systems respond by increased demand temperature [2, 3]. This fact leads us to the conclusion that matching of different loads with different supplies must be based on both magnitude (quantity) and exergy (quality).

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Figure 1. Typical Commercial Building Loads [2, 3, 4-a, 4-b]. LITERATURE SURVEY The most recent trend is 4G (Fourth Generation) district energy systems where low temperature heat (for DHW, LowEx heating), high temperature heat (for conventional space heating, process heat), cold (space cooling), power generated with conventional and renewable energy systems and resources to exchange them all with different exergy with buildings. In spite of this complexity U. S. Department of Energy (DOE) has recently defined Zero-Energy Building (ZEB) in the following form [5, 6]: Zero Energy Building is “an energy-efficient building where, on a source energy basis, the actual annual delivered energy is less than or equal to the on-site renewable exported energy.” This definition also applies to campuses, portfolios, and communities." The first obvious flaw of this definition is the fact that energy supplied-received balance is simply based on an annual single cumulative basis of all kinds of energy with different exergy values without regarding exergy balance of different resources and demands on an instantly basis like hourly basis, which result in varying exergy destructions and associated CO2 emissions that are hidden in the Second Law. Furthermore, an on-site cogeneration system running on natural gas may be regarded to export "renewable" or "waste" energy and exergy in the amount of its actual fuel savings. According to the above DOE definition this is not possible. This issue was first addressed by Kilkis, Ş. [7] by developing new definitions for net or near building in terms of exergy (NZEXB, nZEXB). In fact, according to Marszal and Heiselberg, the definiton of net or near zero building is quite complicated [8]. One complication is the difficulty of load matching [9]. Total energy demand in the building is a sum of thermal and electricity demand; however, many studies focus only on one demand neglecting the other. This issue is raised by Able [10]: “Many low-energy building projects seem to have been based on the idea 'decrease heat supply at any cost'. In some cases, this has resulted in 'zero-energy buildings' which, it is true, do not need any heat supply but do, instead, indirectly need electricity, e.g., to operate the heat pump included in the system.” In the 1970’s and 80’s, when large part of energy use in the buildings was mostly due to the heating (space heating and domestic hot water) in publications the zero energy buildings were actually zero-heating buildings, since only heating demand was accounted into a zero balance. Esbensen, et al. [11] describe an experimental ZEB house in Denmark and point out: “With energy conservation arrangements, such as high-insulated constructions, heat-recovery equipments and a solar heating system, the Zero Energy House is dimensioned to be self-sufficient in space heating and hot-water supply during normal climatic conditions in Denmark. Energy supply for the electric installations in the house is taken from the municipal mains.” Saitoh, [12] and Saitoh, et al. [13] in their studies present a Natural Energy Autonomous House in Japan. According to authors: “… a multi-purpose natural energy autonomous house will meet almost all the energy demands for space heating and cooling as well as supply of hot water for standard Japanese house in 10-15 years. For this purpose, solar energy, the natural underground coldness and sky radiation cooling are utilized.”

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[Abstract:0008][Modeling and Software] COMPARISON OF 2D/3D NUMERICAL ANALYSIS OF THE AIR SIDE HEAT TRANSFER AND

LAMINAR FLOW IN MULTILOUVERED FIN AND FLAT TUBE HEAT EXCHANGER

Nihal Ugurlubilek1, L Berrin Erbay1 1Eskisehir Osmangazi University, Turkey

Corresponding email: nihalugrl@gmail.com, lberbay@gmail.com

SUMMARY

In this study, an analysis for 2D / 3D air side heat transfer and laminar flow in multilouvered fin heat exchanger has been presented. The louver angles (Lα), the fin height (Fh) and the flow depth (Fd) are assumed Lα = 20o, 24o, 28o, 30o and 32o Fh= 8 mm and Fd =20 mm, respectively. The inlet air and the wall temperatures are taken 298 K and 314 K, respectively. Air velocity of 3.75 m/s is used for laminar flow. Colburn number (j) and friction factor (f) versus louver angles are presented. It has been found that the maximum volume goodness factor value (η) occurs for Lα= 20o in the problem under consideration. 2D/3D results are compared and observed that 2D heat transfer results are bigger than that of 3D up to about 120% as the friction results are closer for all louver angles. INTRODUCTION In the air cooled compact heat exchangers, the air side thermal resistance constitutes 80% or more of the total thermal resistance [1]. Thus, any enhancement on the air side greatly increases the effectiveness of the heat exchanger. One of the most important applications that achieve this enhancement is the application of extended surfaces. Extended surfaces with multilouvered fins are one sort of passive applications. They increase the thermo-hydraulic performance of the heat exchanger by increasing the outher surface area of the heat exchanger. The weight and dimension of the heat exchanger decrease with the increasing of thermo-hydraulic performance. Additionally, the quantity of coolant fluid in the heat exchanger decreases as a result of decreasing of heat exchanger volume which in turn decreases the pumping power. The heat transfer and fluid flow analysis on the fin and tube heat exchanger have been studied by several researchers [1,2,3,4,5,6,,7,8,9,10] Numerous experimental studies also found in the open literature have been performed on the multilouvered fin heat exchanger. Malapure et al. [1] numerically investigated fluid flow and heat transfer characteristics over multilouvered fins and flat tube compact heat exchangers. They performed simulations for 15 different geometries with varying louver pitch, louver angle, fin pitch and tube pitch and for different Reynolds number of 60-1800 reported that the Stanton numbers and friction factors were in good agreement with the experimental data presented by Achaichia and Cowell [11] except at low Reynolds number; both Stanton number and friction factor decreased with the increasing fin pitch; the maximum heat transfer coefficient was obtained at the louver angle of 28-29° for Re=1000. Kim and Bullard [12] performed an experimental study on the air-side heat transfer and pressure drop characteristics for multilouvered fin and flat tube heat exchangers.They found that heat transfer coefficient decreased with flow depth and increased with air velocity; the effect of louver angle on heat transfer was different for flow depth, fin pitch and Reynolds number; pressure drops increased with louver angle, flow depth and decreased with fin pitch. Dong et al. [13] performed experimental studies on the air side heat transfer and pressure drop characteristics for 20 types of multilouvered fin and flat tube heat exchangers. They reported that heat transfer coefficients decreased with fin length and the fin pitch increasing and fin height decreasing; besides the pressure drop decreases with fin length decreasing and the fin pitch and fin height increasing at the same frontal velocity. Atkinson et al. [3] investigated two- and three-dimensional numerical models of flow and heat transfer over multilouvered fin arrays in compact heat exchangers. They compared Stanton number and friction factor values with experimental data by Achaichia and Cowell [11]. Aoki et al. [14] performed an experimental study on the heat transfer coefficient at different multilouvered arrangement and reported that the heat transfer coefficient decreases with the increasing the fin pitch at low air velocity and increased up to 28°-30° of louver angle, then decreased after this values. Perrotin and Clodic [9] investigated the heat transfer and pressure drop characteristics of compact multilouvered heat exchangers by numerically both 2D and 3D. They found that 2D results considering constant fin temperature overpredicted the heat transfer coefficient up to 80% whereas 3D results considering the effects of the fin thermal conductivity and the tube geometry was about13% compared to the experimental data; 2D model was sufficient for the calculations of pressure drop. As it is seen in studies mentioned above, the louver pitch, the fin pitch, the louver angle and the flow depth have significant effect on the thermo-hydraulic performance of the multilouvered fin heat exchanger and there is still a requirement on numerical investigations considering both 2 and 3 dimensional cases. Therefore; in this study, air side heat transfer and pressure drop characteristics were numerically investigated over multilouvered fin and flat tube heat exchanger. For 5 types of geometry with different louver angles (Lα) 20°, 22o,24°, 26o, 28°, 30°and 32° were considered at the same Reynols number of 430 based on the louver pitch. The fin pitch (Fp), the louver pitch (Lp), tube pitch (Tp) and the flow depth (Fd) are taken as 1.5 mm, 1.7 mm, 12.4 mm and 20 mm, respectively. The 2d and 3D results are compared with each other and the results reported in open literature at the same Reynolds number.

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MATHEMATICAL MODEL The analysis for 2D and 3D air side heat transfer and laminar flow in multilouvered fin heat exchanger is performed by considering fundamental conservation principles. The mass, momentum and energy conservation equations are written by assuming that the flow is steady, incompressible and single phase. The Reynolds number based on louver pitch is taken about 430. The working fluid is the air assumed as ideal gas. The governing equations are given as follow

⋅ V= 0 , (1)

ρ (V ⋅ V) = - p + μ 2V , (2)

ρ Cp ( V. T)= k 2 T , (3) Ideal gas law is used for the density solution of air

, (4) The dynamic viscosity of air is computed from the Sutherland Law given as follow

, (5)

where µo (1.716x10-5 kg/ms) is the dynamic viscosity at To (273.15 K) and Cs is the Sutherland temperature (110.4 K). Because of the little change of the air temperature over the computational domain thermal conductivity and specific heat are assumed to be constant at about mean air temperature. Finite volume method [15] is chosen to solve governing equations by using Ansys Fluent 14 code [16]. Simple algoritm is used for the calculations of pressure fields. The residuals are about 10-5 for all the quantities. Second ordary discretization sheme is used for all simulations.

Geometry and boundary conditions In this paper, the effects of the louver angle (Lα) on the air side heat transfer and laminar flow are numerically investigated for both 2D and 3D over multilouvered fin and flat tube heat exchanger. The solutions are obtained for total 5 geometries having different louver angles (Table 1). Figure 1 presents the geometry under consideration.

Figure 1. The geometry considered. Table 1. The details of multilouvered fin heat exchangers investigated Lα Lh Fp Fl Fd Lp t dh (deg) (mm) 20, 24, 28, 30, 32 8 1.5 10 20 1.7 0.1 2.8

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2D study Figure 2 shows 2D physical model and boundary conditions. The inlet temperature and velocity of air are given as 298 and 3.75 m/s, respectively. The temperature of the fin wall is assumed as constant (314 K). The outlet is opened the atmosphere. The periodic boundary condition is assumed on the upper and bottom sides of the geometry. The non-slip boundary condition is given on the fin walls.

Figure 2. The 2D physical model and boundary conditions. For the grid independency, the solutions of the grid structure having four different node numbers (32326, 40051, 90741 and 181207) are obtained and compared with each other. The grid structure applied to the geometry with Lα=20o is shown in Figure 3. As seen in this figure, the quadrilateral elements are applied to the geometry. For the grid structure, the sizes of quadrilateral elements are 0.02 mm for the fin and the fluid region around the fin, 0.05 mm for the entry and out fluid regions. The results are shown that the outlet temperature of air doesn’t considerable change for the bigger node numbers than 90741 (Figure 4).

Figure 3. 2D grid configuration considered.

Figure 4. The outlet temperature of air versus the number of nodes for 2D study. 3D study Figure 5 shows the module considered for 3D study. The module is obtained with a part of tube and a half fin. To reduce the computational mesh size, one half of the fin length is considered by using the symmetry. The height of the computational domain is equal to the fin pitch. Periodic boundary conditions are applied at the top and the bottom side. The computational domain is splitted of the entry, the fin and the outlet regions. The tube wall temperature is maintained at 314 K.

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Figure 5. The boundary conditions and computational domain of 3D geometry. For 3D study, three grid structures with quadrilateral cell type having different node numbers (225250, 1540496 and 4496752) are constructed to obtain grid independency solution. The results are shown that the outlet temperature of air and the pressure drop doesn’t considerable change for the bigger node numbers than 1540496. Figure 6 shows the 3D grid structure accepted. As seen in this figure, there is more grid density near the fin region.

Figure 6. 3D grid structure. Verification of the study The reliability of the numerical model and grid structure applied on the computational domain has an uthmost importance to be sure about the results. For this purpose, the accepted numerical model and grid structure are applied to the geometry which were investigated experimentally by Kim and Bullard [12]. The fin pitch, the fin depth and the louver angles are accepted as Fp=1.5 mm, Fd=20 mm and Lα =15o, 17o, 19o, 21o, respectively. The number of louvers is chosen as10.

The results of the 2D study are compared with the experimental results of Kim and Bullard [12] and indicated in Figure 7. The numerical Colburn and Fanning friction factor values are overestimated about 131% and 64% than those of the correlations of j and f derived from the experimental data of Kim and Bullard [12], respectively.

Figure 7. Comparisons of numerical and experimental [12] results. The discrepancies between the numerical and the experimental results are accordant with the results presented by Perrotin and Clodic [9] and Atkinson et. al [3]. Atkinson et. al [3] found that the 3D numerical results are better than those of 2D. They said that the Colburn factor overestimated about 100% are obtained for some geometry whereas the results of Fanning friction factor are generally good. Thus, it can be said to be reliable of this numerical study.

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Performance parameters In this study, Colburn number (j) and Fanning friction factor (f) are investigated as the performance parameters for the heat transfer and pressure drop, respectively. Performance parameters are given as [17] Colburn number, j

3/2PrStj , (6) where St is Stanton number and given as

pc

c

CuhSt

, (7)

Fanning friction factor, f

c

ac

AAu

Pf2

2

, (8)

Reynolds number based on louver pitch is given as

pc

Lp

LuRe , (9)

where uc is the mean velocity through minimum flow area and given by

c

frinc A

Auu , (10)

where Afr and Ac are the frontal area and minimum flow area, respectively. The heat transfer coefficient, hc is

mac TA

Qh

, (11)

where Aa is the total heat transfer surface area and ΔTm is the logaritmic temperature difference of air which is defined as

c

h

chm

TT

TTTln

, (12)

where temperature differences presented with and are given as ∆ , (13)

∆ , (14)

Both the j and f results have to be evaluate simultaneously with regards to obtain the thermo- hydraulic performance of heat exchangers studied. For this purpose, volume goodness factor values calculated as

3/1fj

, (15)

RESULTS AND DISCUSSIONS In this paper, the effects of the louver angle on the heat transfer and laminar flow are numerically investigated for the multilouvered fin type heat exchanger. For this purpose, the solutions are obtained for different louver angles (20o, 24o, 28o, 30oand 32o). The 2D and 3D j, f and volume goodness factor results are presented versus louver angles considered. They are compared with the results presented open the literature and each other. The Reynolds number value calculated for all configurations is about 430. Figure 8 shows 2D and 3D j and f values obtained for all geometries studied. It can be said from this figure that the increasing of the louver angle has no considerably effect on heat transfer rate but the friction factor increases after the louver angle of Lα=24o. Nevertheless, it can be seen that 2D and 3D f values are close with each other whereas 2D j values are bigger than that of 3D about up to 120%. This may be due to the fact that the air flow is subjected to bigger resistance increasing of the louver angle.

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Figure 8. Colburn – j and Fanning friction –f factors for Lα studied (for Re=430) The volume goodness factor (η) values are shown in Figure 9 for the all geometries studied. As seen in this figure, the best performance is obtained for Lα=20o. This may be because the increase of the louver angle considerably affects the presure drop. When Lα increases the resistanse against the flow also increases and the friction factor is more effective than the heat transfer rate to the variation of volume goodness factor. Thus, η decreases with increasing Lα.

Figure 9. Volume goodness factor results versus louver angles. In Figure 10, the results of 2D and 3D are compared with the results of literature presented by Perrotin and Clodic [9]. They said that the 2D numerical results overpredict up to 80% compared to the experimental data. They indicated that acceptance of the constant fin temperature and neglecting the tube surface effect overestimate the Colburn factor but the tendency of curves are similar. As seen in Fig. 10, 3D j results is less than that of 2D while 3D f results bigger than that of 2D. This situation is also suitable with the literature results. It can be also seen that the difference between 2D and 3D j results is bigger than that of f results. Nevertheless, it is seen from this figure that the experimental j results of Kim and Bullard [12] are quite small in comparison with the results of Perrotin and Clodic [9] accepted the constant fin temperature for both 2D and 3D study. Considering that the discrepancy about 2 times, it can be said to be compatible this situation with the results of this study. Thus, it can be said to be reliable of this numerical study.

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(a) (b)

Figure 10. Comparison of the j (a) and f (b) results of 2D/3D (this study) and literature [9]. CONCLUSION In this study, the effect of the louver angles on the air side heat transfer and laminar flow in multilouvered fin heat exchanger has been presented numerically. 2D/3D j, f and results are obtained and compared with each other and the open literature. Different geometries (Lα=20, 24, 28, 30 and 32o) are solved by Ansys Fluent 14 [16]. Major conclusions are summarized as follows:

1. The Fanning friction factor increases with the increasing louver angle, especially for Lα>24o. 2. The heat transfer rate doesn’t considerably change versus louver angle. 3. The maximum performance is obtained for Lα=20o. 4. The fan power increases with increasing louver angle, especially for Lα>24o. 5. 2D numerical results are bigger than that of 3D. 6. 2D j results are bigger than that of 3D about up to120% . 7. 2D f results are closely found with those of 3D. 8. 2D and 3D results are in good agreement with that of literature.

acknowledgment

The financial support of this study by SANTEZ (Project no: 00865.STZ.2011-1) and ARCELİK Inc. is greatly appreciated.

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REFERENCES 1. V.P. Malapure, S.K. Mitra, A. Bhattacharya, Numerical investigation of fluid flow and heat transfer over multilouvered fins in compact heat exchanger, International Journal of Thermal Sciences 46 (2007) 199–211. 2. S. Akyüz, Investigation of Effect of Fin Height and Fin Pitch to Performance on Air Cooled Mini Micro Channel Condensers, master of science thesis, Eskisehir Osmangazi University, Eskisehir, Turkey, 2013. 3. K.N. Atkinson, R. Drakulic, M.R. Heikal and T.A. Cowell, Two-and three-dimensional numerical models of flow and heat transfer over louvred fin arrays in compact heat exchangers, International Journal of Heat and Mass Transfer 30 (1998) 3952-3979. 4. H. Yılmaz, , Numerical Investigation Of The Performance Of Condenser In Domestic Refrigerator, master of science thesis, Eskisehir Osmangazi University, Eskisehir, Turkey, 2013. 5. N. Ugurlubilek,, L. B. Erbay, B. Dogan, Bir ısı değiştiricide dış akış ısı transferi karakteristiğinin sayısal incelenmesi (in Turkish) in: Proceeding of ULIBTK’13, Samsun, 2013, pp. 380-385. 6. N. Ugurlubilek, L. B. Erbay, B. Dogan, Bir ısı değiştiricide dış akış basınç düşümü karakteristiğinin sayısal incelenmesi (in Turkish) in: Proceeding of ULIBTK’13, Samsun, 2013, pp. 386-391. 7. C.T. Hsieh, J. Y. Jang, 3-D thermal-hydraulic analysis for louver fin heat exchangers with variable louver angle, Applied Thermal Engineering 26 (2006) 1629–1639. 8. D.K. Tafti, J. Cui, , Fin-tube junction effects on flow and heat transfer in flat tube multilouvered heat exchangers, International Journal of Heat and Mass Transfer 46 (2003) 2027–2038. 9. T.D Perrotin, D. Clodic, Thermal-hydraulic CFD study in multilouvered fin-and-flat-tube heat exchangers, International Journal of Refrigeration 27 (2004) 422–432. 10. M.E. Springer, K.A. Thole, Experimental design for flowfield studies of multilouvered fins, Experimental Thermal and Fluid Science 18 (1998) 258-269. 11. A. Achaichia,T.A. Cowell, Heat transfer and pressure drop characteristics of flat tube and louvered plate fin surfaces, Experimental Thermal and Fluid Science 1(2) ( 1988) 147-157. 12. M.H. Kim, C.W. Bullard, Air-side thermal hydraulic performance of multi-louvered fin aluminum heat exchangers, International Journal of Refrigeration 25 (2002) 390-400. 13. J. Dong, J. Chen, Z. Chen, W. Zhang , Y. Zhou, Heat transfer and pressure drop correlations for the multi-louvered fin compact heat exchangers, Energy Conversion and Management 48 (2007) 1506–1515. 14. H. Aoki, T. Shinagawa, K.K. Suga, , An experimental study of the local heat transfer characteristics in automotive multilouvered fins, Experimental Thermal and Fluid Science 2 (1989) 293–300. 15. S.V. Patankar, Numerical Heat Transfer and Fluid Flow, Hemisphere, 1980. 16. Ansys Fluent. 2011. version 14.0.1 17. W.M. Kays, A.L. London,Compact heat exchangers, 3rd New York, 1984.

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[Abstract:0009][Heating, Climatization and Air-conditioning Applications in Buildings] NEXT GENERATION REFRIGERANT R-32 AND THE LATEST DEVELOPMENTS ABOUT

THE DISSEMINATION OF R-32 AROUND THE WORLD

Dr. Andaç YAKUT1 1Daikin Turkey

Corresponding email: a.yakut@daikin.com.tr ABSTRACT World’s first R-32 residential air conditioner was launched in November 2012 in Japan. Since 2013, an increasing number of other air conditioner manufacturers have been releasing R-32 air conditioners, mainly in Japan, as R-32 gains growing recognition as a next-generation refrigerant. In the end of 2013, Europe’s first R-32 residential air to air heat pump was introduced. Currently more than 6 million units using R-32 have been sold worldwide. R-32 units are on sale in more than 40 countries. Latest developments proven that R-32, which has a low global warming impact, is the most suitable next generation refrigerant for residential and commercial air conditioners. Compared to the commonly used refrigerant R-410A, the Global Warming Potential of R-32 is only one third (GWP: 675), while it allows for a smaller refrigerant volume and higher energy efficiency. In this paper, first, R-32 refrigerant properties and aspects of using R-32 in air conditioners & heat pumps are explained. Afterwards, the revised EU F-Gas Regulation and its connection to R-32 is introduced in detail. Also R-32 Installation & Service Aspects is explained. Finally, the latest developments about the dissemination of R-32 all around the world is investigated. Keywords: Refrigerant, R-32, Low GWP, F-gas regulation. INTRODUCTION Refrigerants are crucial to air conditioning, circulating inside the air conditioner and transporting heat. However, the Montreal Protocol and the Kyoto Protocol restricted the use of conventional refrigerants that deplete the ozone layer and contribute to global warming, and the world needs refrigerants that mitigate these harmful effects. Industrialized countries have already converted to HFCs like R410A that don’t deplete the ozone layer, but these refrigerants still have the problem of having a high global warming impact. In 2013, developing countries began phasing down the use of conventional HCFC refrigerants. Air conditioner demand is growing in developing countries, and if these countries follow industrialized countries in adopting R410A, global warming will accelerate. It is therefore crucial that the world convert to a next-generation refrigerant. Industrialized countries are also aiming to reduce HFC emissions and concerned parties are actively seeking to find next-generation refrigerants. Latest developments proven that R-32, which has a low global warming impact, is the most suitable next generation refrigerant for residential and commercial air conditioners. 1. What is R-32? R-32 is a single component HFC which is called “difluoromethane” (CH2F2). It is a refrigerant which has been used for many years as a component of the refrigerant blend R-410A (which is 50% R-32 and 50% R-125). Most of the industry manufacturers now recognise that using R-32 in its pure form instead of R-410A or other types of blends offers a number of advantages. What is GWP? Global Warming Potential (GWP) is a number which expresses the potential impact that a particular refrigerant would have on global warming if it were released into the atmosphere. It is a relative value which compares the impact of 1kg of refrigerant to 1kg of CO2 over a period of 100 years. Although this impact can be avoided by preventing leaks and ensuring proper end of life recovery, choosing a refrigerant with a lower GWP and minimising the volume of refrigerant will reduce the risk to the environment if a leak were to occur accidentally. What is ODP? Ozone Depletion Potential (ODP) is a number that refers to the harmful impact on the stratospheric ozone layer caused by a chemical substance. It is a relative value which compares the impact of a refrigerant to a similar mass of R-11. Thus, the ODP of R-11 is defined to be 1. 2. R-32 Refrigerant Properties R-32, R-410A, R-134a and other refrigerants currently used in the European Union do not deplete the ozone layer. The previous generation refrigerants such as R-22 had a detrimental effect on the stratospheric ozone layer because they contained chlorine. Since 2004, EU regulations have banned any new equipment using ozone-depleting refrigerants such as R-22. Since January 2015, servicing existing equipment with R-22, even with recycled R-22, has also been banned [1]. R-32 is a single component HFC and it does not deplete the ozone layer (ODP=0). Also the GWP of R-32 is only 1/3rd of the GWP of R410A. Basic properties of R32 in comparison with some other refrigerants can be seen in Table 1.

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Table 1: Properties of R-32 3. Aspects of Using R-32 in Air Conditioners & Heat Pumps Each manufacturer needs to make choices depending on the application and the needs of the market and taking into account energy efficiency, safety, affordability, local legislations and standards. There is no “one-size-fits all” refrigerant. R-32 is a refrigerant that can be easily handled by installers and service technicians as it can be charged in both gas and liquid state. In addition, R-32 can be recovered and reused without difficulty. And there is no need to be concerned about changes in composition in case of a leakage in the system. R-32 is also a better solution from environmental point of view: it has very high capacity and energy efficiency characteristics; and its Global Warming Potential (GWP) is only one third of the GWP of R410A (GWP of R-32 is 675*, and R410A is 2088*). We can summarize the advantages of R-32 as below [2]:

Not depleting the ozone layer (ODP=0) GWP only one third of R-410A (GWP of R-32 is 675, and R-410A is 2088) Reduced refrigerant charge possible Higher Energy Efficiency compared to R-410A More compact design possible Acceptably safe because of lower flammability (Class A2L) Refrigerant Production capacity is available (R-32 is a component of R-410A) Easier to recycle and reuse (single component refrigerant)

4. The Revised EU F-gas Regulation If released into the atmosphere, refrigerants can have an impact on global warming. In 2006, EU regulators implemented the so-called ‘F-gas regulation’ to minimise the risk of a certain group of fluorinated greenhouse gases, of which the most important are the HFC gases commonly used as refrigerants [3]. A certification system was introduced for installation and service companies. This, in combination with mandatory leak inspections for systems with a charge of 3kg or more, has successfully reduced emissions. (Note: the revised F-gas regulation still requires leak inspections, but the threshold changed to 5 Tonnes CO2 equivalent or more, which is equivalent to 2.4 kg of R-410A or 7.4 kg of R-32.) Despite the fact that F-gas emissions currently only represent 2% of total EU greenhouse gas emissions, EU regulators and the industry have recognised that more can be done in view of the EU roadmap towards a low carbon economy. That is why a revised F-gas regulation came into force at the beginning of 2015 [4]. This regulation encourages the design of equipment with lower CO2 equivalent refrigerant values. In other words, equipment with lower refrigerant GWP or a lower refrigerant charge, but ideally a reduction in both (commonly known as the ‘phase down’ on HFC consumption, expressed in CO2 equivalents). Thanks to the revised F-gas regulation, the EU’s F-gas emissions will be cut by two-thirds by 2030 compared with 2014 levels (Figure 1). *Global Warming Potential (GWP) values based on the Fourth Assessment Report by the Intergovernmental Panel on Climate Change.

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Figure 1: The impact of the EU legislation on F gas emissions

There is no refrigerant which can meet the needs of every kind of application. This means that when selecting refrigerant, taking into account not only reduction of GWP and the amount used, but also aspects such as energy efficiency, safety and affordability is important. For example, selecting a refrigerant with a lower GWP, but which uses more energy would not be a good choice, as it would be counterproductive for the total product’s global warming impact. The new F-gas regulation bans the use of refrigerants with a GWP above 750 in single split air conditioners with a refrigerant charge below 3 kg from 2025. R-32 is already compliant with this requirement (Table 2). Therefore the sooner the industry changes to lower GWP refrigerants like R-32, the earlier the environmental impact of HFC emissions can be reduced.

Table 2: R-32 and the new EU F-gas regulation

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5. R-32 Installation & Service Aspects The installation and service methods for R-32 are very similar to R-410A. Working pressures for R-32 and R-410A are similar (Design pressure for R-410A: 4.15 MPa, for R-32: 4.29 MPa). For charging requirements, R-32 is easier to handle as it can be charged in both a gas and a liquid state (not possible with R-410A which always needs to be charged in liquid state. Adding R-410A in gas form may cause the refrigerant composition to change, preventing normal operation). When installing or servicing R-32 equipment it is particularly important to check that manifolds, leak detectors and recovery pumps can be used with R-32. Tools are available which are permitted and suitable for both R-32 and R-410A. If in doubt, it should be checked with the tool supplier. For recovering R-32 an approved R-32 recovery cylinder is needed. Other installation tools such as charging hose, scale, torque wrench, flare tool, pipe bender and the vacuum pump are the same, so R-410A installation tools can be used (Table 3). Finally the equipment manufacturer’s and the refrigerant cylinder provider’s safety instructions must, of course, always be adhered to [5].

Table 3: R-32 installation & service aspects

6. Recent developments about R-32 World’s first R-32 residential air conditioner was launched in November 2012 in Japan [6]. Since 2013, an increasing number of other air conditioner manufacturers have been releasing R32 air conditioners, mainly in Japan, as R-32 gains growing recognition as a next-generation refrigerant [7]. In the end of 2013, Europe’s first R-32 residential air to air heat pump was introduced. R-32 models have been introduced in Australia, New Zealand, India, Thailand, Vietnam, the Philippines, Malaysia and Indonesia…etc. Currently more than 6 million units using R-32 have been sold worldwide. R-32 units are on sale in more than 40 countries [8]. After the release of first R-32 air conditioners in Europe, some other manufacturers also started to announce their new R-32 units this year [9], [10], [11]. Meanwhile several demonstration and testing projects including local manufacturers in China, Middle East, and Asia was also carried out. Increasing number of developing countries show interest to change from R-22 to R-32. For instance; a project to convert to R-32 in Thailand, where METI is offering financial aid as part of support for developing countries under the Montreal Protocol. R22 use will be banned in Thailand starting in 2017, and the Thai government’s policy is to convert from R22 to R32 as a next-generation refrigerant. On request from METI, one of the leading air conditioner manufacturer company helped Thai manufacturers convert to R-32 and is offered technical training to Thai service engineers. In April 2014, R-32 air conditioner was launched in Thailand [12]. Funded by Japan’s Ministry of Economy, Trade and Industry, One of the industry's leading Japanese company established “training for the trainers” program to expand the servicing network for R-32 in India. 3,600 local installers were trained in this demonstration project. Through this project, India has developed the necessary expertise to service R-32 equipment throughout the country. R-32 residential air conditioner was launched in March 2013 in India [13]. In UK, more than 200 installers have been trained on how to work with R-32 [14]. Turkey’s first residential air conditioner was launched in May 2015 [15]. In Turkey, more than 100 installers trained for how to work with R32. Finally, on 10 September 2015, one of the world's leading air conditioner and refrigerant manufacturer announced it is offering companies worldwide free access to 93 patents, to encourage companies to develop and commercialize air conditioning, cooling and heat pump equipment that use HFC-32 as a single component refrigerant [16].

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REFERENCES 1. http://www.daikin.eu/binaries/ECPEN15-017A_R-32%20white%20paper_tcm507-400605.pdf?quoteId=tcm:507-

400607-64 (Daikin, R32 The Next Generation Refrigerant for Air Conditioners and Heat Pumps White Paper) 2. Next generation refrigerants A Daikin perspective, Hilde Dhont, Daikin Europe N.V.

Environment Research Center, Chillventa 2012 3. Regulation (EC) No 842/2006 of the European Parliament and of the Council of 17 May 2006 on certain fluorinated

greenhouse gases 4. Regulation (EU) No 517/2014 of the European Parliament and of the Council of 16 April 2014 on fluorinated

greenhouse gases and repealing Regulation (EC) No 842/2006 5. Koelmiddel R-32, Hilde Dhont, Daikin Europe Environment Research Center, UCLL event, 28 April 2015 6. http://www.daikin.com/csr/report/2013/p15_18_feature1.pdf 7. JARN Magazine, October 25, 2013, Serial No.537 Volume 45, No.10 8. Daikin Media FAQ, http://www.daikin.com/press/2015/150910/Daikin_Media_FAQ_final.pdf 9. https://www.eseficiencia.es/noticias/kaysun-presenta-su-modelo-eficiente-de-aire-acondicionado-r-32 10. http://www.coolingpost.com/uk-news/second-manufacturer-offers-r32-units/ 11. http://www.coolingpost.com/world-news/panasonic-to-launch-r32-heat-pump/ 12. http://www.daikin.com/csr/report/2014/05_feature01.pdf 13. HFC32 (R32) for A/C Applications; Progress and Actual Use, Tadafumi Mikoshi, Senior Manager CSR & Global

Environment Center, Daikin Industries, LTD, Dubai, UAE (10-11 Sept, 2013) 14. http://www.coolingpost.com/uk-news/business-edge-to-offer-r32-training/ 15. www.daikin.com.tr 16. http://www.daikin.com/press/2015/150910/index.html

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[Abstract:0011][Sanitary System Applications] SCOPE AND COMPARISON OF STANDARDS WHICH REGULATE THE DESIGN OF

HYGIENIC AIR HANDLING UNITS

OrkunYılmaz1 , Hamza Sonkur2

1Systemair HSK, Turkey 2Systemair HSK, Turkey

orkun.yilmaz@systemairhsk.com.tr hamza.sonkur@systemairhsk.com.tr

ABSTRACT

Air handling units are of critical importance not only for hospitals, but also for hygienic environments such as in the food and pharmaceutical industry. With this perspective, the importance of air handling units is increasing with each passing day, with relevant standards being updated at the same time. The most widely accepted of the standards defining the design of hygienic air handling units are DIN 1946-4:2008, VDI 6022-1:2011, and the RLT 01 Air Handling Unit Guideline 01. The EN 13053:2006+A1:2011 standard also includes general information regarding Hygienic Air Handling Units. This study examines hygienic air handling unit requirements in comparison with the current versions of relevant standards as of the date of this statement, and presents the approaches of standards and the differences between the standards.

1. Introduction

The ever increasing air pollution has, in time, raised requirements regarding conditions of comfort as well as conditions of ventilation and air conditioning in hygienic environments. Furthermore, the rising awareness and consciousness in the area of health has brought about the fact that principles formerly accepted as hygienic now constitute design inputs as principles of comfort, and that hygiene requirements have increased further. This increases the importance of the use of hygiene equipment in ventilation and air conditioning systems as well as the quality of the equipment used. This importance becomes even more sensitive when hospitals are considered. The specifications of ventilation equipment used in hospital environments which require high hygiene, high particle density, and high indoor air quality are among the most important factors in achieving these conditions. The principles of production and application of hygienic air handling units have been determined with standards and reinforced with various guidelines. However, when we look at practical applications, major errors are observed in the production and application of hygienic air handling units due to a lack of oversight and differing interpretations. This study discusses with comparisons the properties that are required in ventilation and air conditioning equipment used in hygienic environments, according to EN 13053:2006+A1:2011, RLT Guideline 01, DIN 1946-4:2008, VDI 6022-1:2011 standards and guidelines. EN 13053:2006+A1:2011: Ventilation for buildings. Air handling units. Rating and performance for units, components and sections [1] This is a standard which classifies the testing methods of the performance of an air handling unit in general, the properties of sections and components; and defines testing methods. Within the standard, additional properties that are required in air handling units specifically for hygienic applications have been stated for each section of the air handling unit. VDI 6022-1:2011 Hygiene Requirements for Ventilation and Air-Conditioning Systems and Units [2]

This standard prepared by the Association of German Engineers defines the requirements for the production, planning, application, operation, and maintenance of hygienic air handling units. Although not extensive as the DIN 1946-4:2008 standard with respect to its content, it includes some matters only mentioned in the VDI 6022-1:2011 standard.

DIN 1946-4:2008 Ventilation and air conditioning systems used in healthcare buildings [3]

This standard describes the structure and qualification of ventilation and air conditioning equipment used in healthcare buildings. This standard defines not only design criteria, but also the subjects of system acceptance tests, flow visualization, room classification according to protection classes, and microbiological monitoring.

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RLT Guideline 01: General Requirements for Air Handling Units [4]

This guideline defines the structural requirements of air handling units. With respect to hygienic air handling units, it examines all air handling unit components individually, with reference to the DIN 1946-4:2008 and VDI 6022-1:2011 standards. It adds its own comments to issues not included by these standards, constituting a whole.

Other standards which reference the above mentioned standards, and constitute a reference for some information provided below are the following:

EN 1886:2007: Ventilation for buildings – Air handling units – Mechanical performance [5] EN 1751:2014: Ventilation for buildings - Air terminal devices – Aerodynamic testing of damper and valves [6] EN 779:2012: Particulate air filters for general ventilation. Determination of the filtration performance [7] DIN EN 1822-1:2011: High efficiency air filters (EPA, HEPA and ULPA) - Part 1: Classification, Performance Testing, Marking [8]

2. REQUIREMENTS OF HYGIENIC AIR HANDLING UNITS BY STANDARDS 2.1 Casing The casing structure of equipment used in hygienic ventilation and air conditioning systems is of great importance due to cleanability and accessibility. Furthermore, corrosion resistance of casing elements which come in contact with air flow have a direct influence on the conformance of systems to hygienic requirements.

2.1.1 EN 13053:2006+A1:2011

The EN13053 standard stipulates the use of abrasion resistant, odourless, emissions free, humidity-free material and material that does not allow the growth of micro-organisms for all equipment that come in contact with air flow. The base of the unit must also have a smooth structure to enable cleaning without any residue. In terms of accessibility, the standard states that all equipment must be accessible from both sides through doors and panels, or as an alternative to this it states that units with a height of up to 1,6 m shall be designed with removable equipment. 2.1.2 VDI 6022-1:2011

The VDI 6022-1:2011 standard also ascribes great significance to the subject of the casing. The standard similarly stresses the need for interior surfaces to be without intersections and the need to avoid the use of open cell materials. As is also the case in EN 13053, the standard stresses that removable access panels can be used for access in units with up to 800 mm interior height, but an access door must be present in units where interior height is more than 800 mm. This standard also states the requirement for providing access to all equipment from both the upstream and downstream side. 2.1.3 DIN 1946-4:2008

Since accessibility and cleanability is of great importance for hygienic air handling units, the subject of the casing is discussed in detail in this standard as well. In addition to the important matters in the above mentioned two standards, the DIN 1946-4 standard stipulates some detailed requirements. For instance, surfaces within the airflow must be at least sendzimir galvanized and coated for corrosive resistance. This can be 25 µm for coil coating, 60 µm powder coating or double layer wet coating. While high corrosion resistance is expected of these surfaces, not all surfaces are required to be stainless steel. The sections that must be of stainless steel have been specified in detail in the DIN 1946-4. For instance, condensate basins must be a minimum of 304 stainless steel while base panels, moving parts, and parts that will possibly come in contact with water must be a minimum of 304 stainless steel or AlMg. Furthermore, materials that are within the air current must also be resistant to corrosion that can occur due to cleaning with disinfectants. Another example of the detailed specifications of the DIN 1946-4 standard is door seals. The standard specifically states that these can be fitted, clamping, or foaming, but that the use of glued sealing gaskets is not permitted. Condensate basins are another example of the detailed design specifications in this standard. The standard states that a condensate basin must be installed on the fresh air intake and that the condensate pipe must be of stainless steel material. In fact this standard even stipulates that the width of the pan must be a minimum of 0,5 m and that

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the pan must be capable of draining a minimum of 95% of 5 litres of water per m2 of area within 10 minutes. The DIN 1946-4 standard requires the existence of condensate basins in some sections. These sections are the outdoor air intake section, the cooling coil, humidifier section, and heat recovery section provided that it is installed on both sides. Another requirement for accessibility is for the fresh air intake section to be accessible through a door or panel. This standard also cautions against subjecting the equipment to any contamination or damage during processes such as production, assembly, transport, and field installation.

According to the DIN 1946-4 standard, the mechanical specifications of the casing of a hygienic air handling unit must meet the following figures ‘as a minimum’ (with respect to EN 1886 standard): Mechanical strength D2; thermal transmittance class T3; if outdoor air temperature < -7°C or if the unit is an outdoor unit, thermal bridging factor TB3; if outdoor air temperature > -7 °C, thermal bridging factor TB4; casing air leakage class L2; and filter bypass leakage class F9.

2.1.4 RLT Guideline 01

The RLT guideline also provides for the same articles by referencing these standards to a large extent while it specifies some details separately. For instance it states that locking mechanisms of doors must enable cleaning and resistant to disinfectants. Another detail is that the use of hollow rivets within the air handling unit is not permissible. The RLT guidelines has also seen fit to stress some hygienic principles regarding cabling. According to these, all cabling should be made outside the air handling unit whenever possible, cable conduit must not be used for cabling within the unit, and in cases where cables are installed inside the unit, distances should be kept to a minimum.

2.2 AIR CONNECTIONS AND OPENINGS

The air inlet-outlet openings of a hygienic air handling unit as well as the equipment used in these sites and design specifications must meet certain criteria.

2.2.1 EN 13053:2006+A1:2011

The EN 13053 standard makes recommendations on air velocity in outdoor weather protection equipment. According to this, maximum air velocity for the fresh air side with louvers has been recommended as 2.5 m/s, 3.5 m/s with drop eliminator, and 4.5 m/s with rain hood, while on the exhaust air side maximum air velocity has been recommended as 4.0 m/s with louvers, 5.0 m with drop eliminator, and 6.0 m/s with rain hood.

2.2.2 VDI 6022-1:2011

The VDI 6022-1:2011 discusses air side connections and openings in great detail. A series of design measures are specified from keeping the duct lengths to the unit to a minimum to the provision of drainage, from the cleanability of the drainage to the need to avoid connecting it to the sewage drain. The necessity for installing an inspection opening was emphasized as well. The standard states that exhaust discharge must be made at a higher elevation than fresh air intake, and the following issues were discussed for the fresh air intake opening. According to VDI 6022, the fresh air intake opening must be placed in a location where it will not be affected by the exhaust related odours and emissions discharged from the space. Some of the interesting yet important details are, the need to avoid placing the opening downstream of the wind blowing down the cooling tower, and that the air intake should be installed at an elevation of 1.5 times the possible height of snowfall –a minimum of 0.3 m- in cases where air intake is made over the roof. Also according to VDI 6022, the distance from the fresh air intake opening to the exhaust discharge must be a minimum of 2 m, and the distance to the adjacent building a minimum of 8 m. As can be seen, VDI 6022 has dealt quite extensively with this matter.

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2.2.3 DIN 1946-4:2008

DIN 1946-4 similarly stresses the need for drainage for water that may seep into the air exhaust opening. It also provides details regarding hoods. Some examples of such details are that rain hoods should have a minimum angle of 45°, the need to install wire meshes with a maximum dimension of 20 x 20 mm in front of the hoods on suction and pressure sides of outdoor units, and that access should be provided for cleaning on one side. According to DIN 1946-4, weather protection must also be provided when the unit is not in use.

The material from which the elastic connection of the unit is made should have a closed cell structure without grooves and indentations. Another matter specifically stated by the DIN 1946-4 standard is the necessity to have the outdoor intake opening a minimum of 3 m above ground level.

2.2.4 RLT

The provision set forth on this subject in the RLT standard in addition to the provisions of the EN 13053:2006+A1:2011, DIN 1946-4:2008, and VDI 6022-1:2011 standards is that surface quality is required to be a minimum of hot dip galvanized sheet sheeting for exterior surface, and a minimum of hot dip galvanized and painted/coated for interior surface.

2.3 DAMPERS AND MIXING SECTIONS

Dampers used in the event of power outages, or repair and maintenance in order to isolate the hygienic air handling unit from ducts and the external environment, or when impermeability is required as a result of a change in the status of occupation of a space with a certain cleaning class, are of critical importance in systems. Therefore relevant standards also include articles pertaining to material and leakage requirements of dampers used in hygienic systems.

2.3.1 EN 13053:2006+A1:2011

Dampers are discussed with special detail in the EN 13053 standard. EN 13053 stipulates that the leakage class of dampers used in hygienic air handling units must be at least Class 2 according to the EN 1751:2014 standard. The leakage class of dampers used in locations with special hygiene requirements must be Class 4 according to the same standard.

2.3.2 VDI 6022-1:2008

The VDI 6022-1:2011 standard does not deal extensively with the subject of dampers. The standard stresses that when the unit is brought offline, the damper must cut off air flow from the unit and that it must be fully isolated from outdoor air and the duct.

2.3.3 DIN 1946-4:2008

One of the important details of dampers described in the DIN 1946-4 standard with regard to hygiene is that gear wheels of dampers are not permitted in the air flow. This standard stipulates that dampers must be installed on the intake and outlet openings of the air handling unit, and prescribing the material for dampers that take in fresh air, required the use of stainless steel (SS304) or aluminum (AlMg). Other requirements stated include the automatic closing of outdoor air dampers in the event of a power cut. Unlike other standards, the DIN 1946-4 standard also defines the placement of dampers inside and outside the unit, and stated that dampers in indoor units can be placed inside the unit or outside it with a double layer of insulation, and that dampers must be placed inside the unit for outdoor units.

2.3.4 RLT

In addition to the articles stated in EN 13053:2006+A1:2011, DIN 1946-4:2008, and VDI 6022-1:2011, the RLT Guideline emphasizes that dampers must be hot dip galvanized or painted/coated.

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2.4 FILTERS

Filters are among the most important elements of hygienic air handling units in terms of accessibility and cleanability. This subject is discussed in every standard that is analysed within the scope of this study in detail since it directly affects the number and size of particles in hygienic sites.

2.4.1 EN 13053:2006+A1:2011

EN 13053 standard does not include specific filter requirements for hygienic air handling units, but includes filter requirements for air handling units in general. This standard states that the filter area should be at least 10 m2 per 1 m2 cross section area. The first filter stage in an air handling unit must be fitted on the intake side, and the second filter stage must be arranged on the output side under EN 13053. It also emphasizes that the seals used shall be of a closed cell type, and the filter material shall not form a nutrient substrate for micro-organisms. 2.4.2 VDI 6022-1:2011

As per VDI 3022-1:2011, All air filters must conform to either EN 779, or DIN EN 1822, and such conformity must be visible on labels placed any filter. Another point that is indicated under this standard is the fact that the filter must maintain its class throughout its lifetime, and that it must be replaced on the dusty air side. Under VDI 6022, the outdoor air in air handling units must pass through at least class F7 filter, recirculation air through at least class F5, and the mixture air must pass through at least class F7 filter in hygienic applications.

VDI 6022 includes recommendations for circumstances that have high relative humidity in filter sections in terms of protection of filters and compliance with the hygienic criteria as well. For instance, it states that in case air conditions that have relative humidity exceeding 80% at an air temperature of above 0°C, or a relative humidity that exceeds 90% prevail for a long time, the relative humidity must be decreased to an acceptable level by pre-heating by 3K. VDI 6022, unlike other standards, includes details about filter placement and connection types. According to this, filter pockets must not lie flat on the floor in air handling units. It states that sealing must be ensured by a permanent compression member, and if there is a compression member placed on the upstream of airflow, an additional seal must also be used. Under VDI 6022, regardless of the pressure losses of the filters, the first stage filter must be replaced after 1 year or 2,000 hours of operation, and the second stage filter must be replaced after 2 years or 4,000 hours of operation. Furthermore, pressure drop of the filter must be displayed and monitored. Filter chambers must be fitted with an inspection window with an inside diameter of minimum 150 mm. 2.4.3 DIN 1946-4:2008

DIN 1946-4:2008 standard includes filter sections by discussing common requirements with VDI 6022 in general. DIN 1946-4:2008 standard distinctively defines number of filter stages and filter efficiency requirements by room types. An inspection window with a diameter of no less than 150 mm and a lighting fixture must be provided, and they must have cleanable surfaces. DIN 1946-4 standard specifically highlights that ship’s fittings with metal grid cover are not permissible. As for the airtightness, using double sided sealing equipment (push fit, compression) or foam is allowed, glued seals on the other hand are only permitted on filter insert. The filter, according to this standard, must only be replaced from the dusty-air side, and there must be a space of at least the length of the filter pocket. It states that the filter material must be water repellent. Electrostatic filters are not permitted under DIN 1946-4.

DIN 1946-4 states that room Class 1a and 1b require three-stage filtration, with at least class M5 at the first stage (F7 filters are recommended), class F9 at the second, and H13 at the third stage; room Class 2 requires two stages of M5 (F7 is recommended) and F9; and H13 filters must be used at the air outlet of infection rooms, at least F9 in the final filter stage at quarantine rooms or a class H13 filter, if necessary.

DIN 1946-4 states that the first filter stage can be omitted if there’s no humidification at the cooling unit. Another detail that is provided is the fact that the filter pressure drop gauge must not have barrier fluid. DIN 1946-4 differentiates from other standards in this respect.

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2.4.4 RLT

RLT guideline generally refers to EN 13053:2006+A1:2011, DIN 1946-4:2008 and VDI 6022-1:2011 standards for air filters. Another point it states in addition to the abovementioned requirements is that the filter frame material must be made up of hot-dip galvanized steel.

2.5 HEAT RECOVERY

Energy recovery is of great importance since hygienic ventilation and air conditioning systems operate on a continuous basis. The heat recovery equipment used in these systems must also conform to hygienic specifications set forth in standards.

2.5.1 EN 13053:2006+A1:2011

The EN 13053 standard does not deal separately with the matter of hygiene in its section on heat recovery, and stated that the issues of air impermeability and the use of condensate basins where needed should be considered.

2.5.2 VDI 6022-1:2011

The VDI 6022-1:2011 standard emphasizes that the requirements for heating and cooling sections apply in general to heat recovery systems as well. Some points it adds are the need to avoid mixing outdoor air with return air in heat recovery sections except in units with recirculation, and that heat recovery types with no risk of mixture should be used in 100% fresh air units.

2.5.3 DIN 1946-4:2008

All articles discussed under the heading of heat exchangers in the DIN 1946-4:2008 standard also apply to heat recovery sections. Furthermore the condensate basins installed in heat recovery sections must be on the supply and return side, and manufactured of stainless steel (SS304) or aluminum (AlMg). This standard goes so far into detail as specifying that the diameter of drainage of the condensate basin must be 1 1/2". Since the mixing of air between spaces is not permitted in spaces with a high hygiene requirement, also according to this standard, heat recovery systems where air mixing does not occur should be used in such applications. The heat recovery section should be installed downstream of the filter at the supply and return, and the filter class must be a minimum of M5.

2.5.4 RLT

In addition to the EN 13053:2006+A1:2011, DIN 1946-4:2008, and VDI 6022-1:2011 standards, the RTI Guideline focuses on several additional points. For example, according to the RLT Guideline, the frame in rotary and plate type heat recovery exchangers must be hot dip galvanized and painted/coated, fins must be painted/coated or aluminum, and moving parts such as mounting rails must be stainless steel or aluminum (AlMg).

2.6 HEATING AND COOLING SECTIONS

The equipment of heating and cooling sections are important in terms of cleanability. The standards have dealt predominantly with the fin spacing of heat exchangers, and the qualities of material due to the risk of corrosion in the event of condensation.

2.6.1 EN 13053:2006+A1:2011

According to EN 13053, materials used in heating-cooling sections must be corrosion resistant pursuant to hygienic requirements. This standard also defines fin spacing. For example fin spacing in cooling coils without dehumidification must be a minimum of 2.0 mm, while it must be a minimum of 2.5 mm for cooling with dehumidification, a minimum of 4.0 mm for outdoor air pre-heaters, and a minimum of 2.0 in other heat exchangers. The EN 13053 standard has defined the depth of the heat exchanger in proportion to fin spacing, considering the requirements of accessibility and cleanability. For example for a fin spacing of 2 mm, this depth musty be 300 mm for offset tubes, and 450 mm with tubes in line, and for requirements higher than these, the heat exchanger should be split. Another issue that should be considered in the design is the necessity for heat exchangers to be accessible

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from both sides. The standard also stresses the fact that units with height up to 1,6 m must be cleanable without removing any parts. There are also some provisions in this section regarding drop eliminators. One of these is that drop eliminators should only be used where needed, and cooling coils without drop eliminators should be selected when possible. With respect to materials and design, only corrosion-proof drop eliminators with pull-out function for cleaning, with access via door or access panel must be used and fins must be demountable for cleaning. As seen, the EN 13053 standard has dealt quite extensively on this subject. Some provisions regarding drainage and siphoning include the requirement to avoid connecting the drainage directly to the wastewater network and the requirement to use self-filling siphons. The requirement that cooling coils with dehumidification must not be located immediately upstream of filters or silencers, and that heaters or fans must be installed between are some points set forth in the EN 13053. On the basis of all these details, it is observed that the EN 13053 standard is quite sensitive to the details pertaining to heating and cooling sections, as required by the rules of hygiene.

2.6.2 VDI 6022-1:2011

One of the matters emphasized regarding this subject in the VDI 6022-1: 2011 standard is that materials used in cooling sections must be corrosion resistant. This standard places emphasis on requirements whose importance for hygiene can be understood such as fins being flat and easily cleanable, on the condensate basin being manufactured of stainless steel (SS304) or aluminum (AlMg), and preventing the carry over of water to downstream sections for the purpose of preventing corrosion and the growth of micro-organisms.

2.6.3 DIN 1946-4:2008

The DIN 1946-4:2008 standard is for the most part parallel to VDI 6022-1:2011 in this regard. However, coil materials are also described in detail in this standard. For example, the mounting rails of the cooling coil must be manufactured of stainless steel (SS304) or aluminum (AlMg) material. Requirements specified for coil materials include the requirement that the fins in the heating coil be aluminum or copper, the frame be hot dip galvanized, tubes be copper, collectors be black coated steel, galvanized steel or copper. In the cooling coil, corrosion resistance has been kept to a higher grade in general, and aluminum has been required for fins, stainless steel (SS304) or corrosion resistant aluminum (AlMg) for the frame, copper for the tubes, and copper material for collectors. Another issue stipulated in the DIN 1946-4 is that all drainage directions should be on the same side. Also as a hygiene requirement, all parts in wet areas have been required to be cleanable. According to this standard as well, the cooling coil and drop eliminator must be placed upstream of the second filter stage. The fin spacing in the cooling coil must be a minimum of 2.5 mm. Another point specifically stated in this regard is that the cooling coil must be visible from both sides. Along with all these details, the DIN 1946-4 standard can be said to describe the subject in more detail as compared to other standards.

2.6.4 RLT

In addition to the articles stated in EN 13053:2006+A1:2011, DIN 1946-4:2008, and VDI 6022-1:2011, the RLT Guideline emphasizes the drop eliminator frame must be made of corrosion resistant material such as stainless steel (SS304) or aluminum (AlMg).

2.7 SOUND ATTENUATORS

Using sound attenuators becomes compulsory since systems in most of the hygienic areas require to keep the sound below a certain level. The silencers that are used in hygienic systems must be broadly cleanable and have corrosion resistant materials.

2.7.1 EN 13053:2006+A1:2011

Sound attenuators must be installed directly downstream of the first and upstream of the second filter stages, however they must not be placed immediately downstream of dehumidifiers under EN 13053. The splitters must be removable for cleaning, without the need of removing other equipment.

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2.7.2 VDI 6022-1:2011

As per VDI 6022-1:2011 standard, surface material of the sound attenuator must withstand the pressure during the cleaning, and no fibres must be loosened during service.

The requirements regarding humidity in this standard are identical to those for the filters, it states that in case air conditions that have relative humidity exceeding 80% at an air temperature of above 0°C, or a relative humidity that exceeds 90% prevail for a long time, the relative humidity must be decreased to an acceptable level by pre-heating by 3K as a protection measure for sound attenuators as well. 2.7.3 DIN 1946-4:2008

Although the DIN 1946-4 standard does not emphasize many points regarding silencer sections, it states that surface qualities of silencer sections must be a minimum of coated hot dip galvanized steel for section sheeting and for profiles that are within the air flow.

2.7.4 RLT

In addition to the articles stated in EN 13053:2006+A1:2011, DIN 1946-4:2008, and VDI 6022-1:2011, the RLT Guideline emphasizes that mounting rails must be stainless steel (SS304) or aluminum (AlMg).

2.8 HUMIDIFIERS:

In some cases of hygienic air conditioning systems, precise control of air temperature and relative humidity may be needed; and the control of humidity along with temperature is possible through the use of humidifier equipment. Humidification systems used in hygienic systems should be considered with care due to the high risk of corrosion.

2.8.1 EN 13053:2006+A1:2011

Humidifiers have been described in detail in the EN 13053 standard. According to this standard, humidifiers must not be placed directly upstream of filters or attenuator. With respect to cleanability and accessibility, all components must be demountable; All parts in contact with water must be accessible for cleaning and inspection, and be manufactured of material that is resistant to corrosion due to disinfectants. It must be possible to completely empty the section of the equipment when needed. According to this standard, a sight glass of a minimum diameter of 150 mm as well as illuminating equipment must be installed in humidifier sections, and a condensation basin with a gradient must be installed. All components that come in contact with water must have sufficient slope. EN 13053 also defines the material properties of humidifier sections. According to this air washer and high pressure humidifiers must be manufactured of stainless steel (SS3014) or aluminum, and the condensation pan must be manufactured of stainless steel or aluminum (AlMg).

The humidifier must stop automatically when the system stops. This standard recommends UV for de-germination.

2.8.2 VDI 6022-1:2011

The VDI 6022 standard specifies critical levels for the quantity of bacteria in the circulation water in detail. For example, interior surfaces of humidifier must be corrosion resistant; the section must include a condensation basin connected to a non-return siphon; and relative humidity at the section outlet must not exceed 90%. For homogeneous distribution, the length of the humidifier section must be provided by manufacturers. According to the VDI 6022 standard, the humidifier section must be equipped with sufficient lighting and a sight glass. It must be able to monitor the operating status of the light from the outside.

2.8.3 DIN 1946-4:2008

To mention some design criteria specified for humidifier sections of hygienic units by the DIN 1946-4 standard, one of the most important points is that the surface quality must be stainless steel (SS304). This standard also emphasizes the use of only steam humidification for operating theatres. The humidifier must be placed upstream of the second stage filter.

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Like the VDI 6022 standard, the DIN 1946-4 standard specifies details pertaining to the basin. According to this, a condensation basin with a connection diameter of at least 40 mm, manufactured of 304 quality stainless steel or AlMg material must be used in humidifier sections.

As stated in VDI 6022, humidification must stop if the system stops, and condensation must not be allowed to occur. The approach of both standards to this subject is the same.

Also in line with VDI 6022, an inspection window with a diameter of no less than 150 mm and a lighting fixture must be installed in the humidifier cell, and they must have cleanable surfaces. DIN 1946-4 standard does not permit the use of ship’s fittings with metal grid cover.

2.8.4 RLT

The RLT guideline generally refers to the requirements of EN 13053:2006+A1:2011, DIN 1946-4:2008 and VDI 6022-1:2011 standards for humidifier sections and equipment.

2.9 FANS

Fans used in hygienic systems are components that are exposed most to air flow, since they are the equipment that create the air flow. Therefore they are the equipment that will be affected the most from unfavourable air conditions such as high humidity. It is hygienically important for fans to be easily cleaned and resistant to corrosion.

2.9.1 EN 13053:2006+A1:2011

Due to hygienic reasons, and for preventing the discharge of unfiltered air to the environment, suction side air leaks of the fan must be minimized. The fans must be placed between two filter stages.

2.9.2 VDI 6022-1:2011

According to the VDI 6022-1:2011 standard, plug fans should be preferred. If the system is belt driven, a stage of filter must be placed upstream of the fan. An easily removable inspection openings must be installed for fans with a diameter of 400 mm and higher. According to VDI 6022, a sight glass with a minimum diameter of 150 mm and illumination must be installed on fan sections. The accumulation of water within the fan must also be prevented.

2.9.3 DIN 1946-4:2008

Materials that must be used in fan sections have been described in detail in the DIN 1946-4: 2008 standard. According to this, fan blades must be corrosion resistant, the fan casing must be coated/painted sendzimir galvanized, fan and motor bases must be painted/coated sendzimir galvanized, and mounting rails must be coated sendzimir galvanized.

As stated in other standards, a sight glass with a minimum diameter of 150 mm and illumination must be installed on the fan section and surfaces must be cleanable. Ship’s fittings with metal grid cover are not permissible. Good access must be provided for service and maintenance, and a flow rate measuring element must be installed with a display.

2.9.4 RLT

In addition to the EN 13053:2006+A1:2011, DIN 1946-4:2008, and VDI 6022-1: 2011 standards, the RLT Guidelines stipulate that snail fans that have a casing height of up to 1.0 m must be possible to pull out the fan/motor unit, and their gliding surfaces must be a minimum of stainless steel.

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2.10 OTHER REQUIREMENTS

2.10.1 EN 13053:2006+A1:2011

According to the EN 13053 standard, filter sections must be marked with a label which contains information such as nominal air flow rate, the number of filters, type of filter, measurements, filter class, type of medium, initial-final pressure.

2.10.2 VDI 6022-1:2011

An issue stated in addition to a description of the casing and component specifications of a hygienic unit in the VDI 6022-1:2011standard is that hygienic air handling units must be transported in fair weather conditions to avoid exposure to unfavourable conditions until their commissioning. Units must also be protected from dust and moisture at building sites.

2.10.3 DIN 1946-4:2008

The information that must be declared on the unit have been described in detail in the DIN 1946-4: 2008 standard. According to this, a durable name plate must be installed on the filter and fan sections on the unit, to provide the service-maintenance staff with the information needed before working on the unit. The information required in fan sections are the same as in the EN 13053 standard. The plate on fan sections must state the type and year of manufacture, nominal air flow, operating pressure, nominal and maximum speed, nominal motor power, and turning direction.

2.10.4 RLT

RLT guideline refers to the EN 13053:2006+A1:2011, DIN 1946-4:2008 and VDI 6022-1:2011 standards for the labelling and documentation of hygienic air handling units.

3. CONCLUSION In this study, the most important standards which determine the principles of hygienic air handling units have been examined in detail, and the overall properties of air handling units and all requirements for each section have been compared. As mentioned in the introduction, the EN 13053 standard is one which discusses all structural properties of air handling units in general and also describes requirements specific to hygienic applications. For instance, the standard specifies cleanability and accessibility as well as classes some components must conform to according to relevant standards. Since its basic purpose is not to describe hygienic air handling units, it does not go into much detail in some subjects. These details have been provided in the VDI 6022 and DIN 1946-4 standards. The VDI 6022 standard is a standard specific to hygiene applications, which describes hygiene requirements for ventilation and air conditioning systems and equipment. Therefore some requirements that were not included in the EN 13053 standard are explained here. To state an example, the hygienic damage of high relative humidity that can occur in filter sections have been described along with measures that can be taken. Many such details can be found in this standard. The DIN 1946-4 standard presents all properties required in ventilation and air conditioning systems used in buildings and rooms in the healthcare sector in fairly extensive detail. In addition to the above, it can be said to be the standard which provides the most comprehensive information and detail for the hygienic air handling unit compared to the other relevant standards. For instance while other standards only require surfaces exposed to the air flow in hygienic units to be corrosion resistant, the DIN 1946-4 standard describes the sheet metal that must be used in these sections in detail from the type of coating to the thickness of the sheet. Also in terms of certification, DIN 1946-4 standard is one of the most important documents referenced. Although the RLT Guidelines is not essentially a standard like the above mentioned standards, it is a guideline which also reveals the hygiene requirements of air handling units by referencing these standards. It even includes details that have been left out of these standards on some subjects. For example, only in the RLT guidelines is the requirement stated that spiral fans located within unit casings of up to 1 m in height must be removable. All four documents examined in this study state the requirements of hygienic air handling units with various details. According to this, hygienic air handling units must be designed and manufactured in accordance with the principles of relevant standards,

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and in line with the application and the intended use of the unit. In this context, it is important to raise the awareness, information, and consciousness of relevant sectors on the basis of manufacturer, practitioner, and user; to take the necessary steps correctly with respect to oversight; for ensuring the implementation and sustainability of the principles set forth in these documents.

4. DISCUSSION While the above mentioned standards involve many of the subjects regarding the design, application, planning, production, maintenance, and operation of hygienic ventilation units, this study only encompasses the ‘Hygienic Air Handling Unit’ and equipment. Within this scope, only the pertinent parts of the standards have been utilized. Although this study can be used as a general guide for air handling units, it would be best to obtain and consult the original current version of the standard to access all requirements of the standard. Since the study was prepared on a comparative basis, it would be beneficial to review and revise this and similar studies in the event of revision of the standards utilized. The preparation of guidelines on the subject are underway as this study is being prepared. As the ongoing studies are completed, they can be included in this and similar studies.

5. RESOURCES

[1]EN 13053:2006+A1:2011: Ventilation for buildings – Air handling units – Rating and performance for units, components and sections [2]VDI 6022-1:2011 Hygiene requirements for ventilation and air-conditioning systems and units [3]DIN 1946-4:2008 VAC systems in buildings and rooms used in the health care sector [4]RLT Guideline General Requirements for Air Handling Units [5]EN 1886:2007 :Ventilation for buildings – Air handling units – Mechanical performance [6]EN 1751:2014 : Ventilation for buildings - Air terminal devices – Aerodynamic testing of damper and valves [7]EN 779:2012 : Particulate air filters for general ventilation. Determination of the filtration performance [8]DIN EN 1822-1:2011 :High efficiency air filters (EPA, HEPA and ULPA) - Part 1: Classification, performance testing, marking; German version EN 1822-1:2009

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[Abstract:0012][Urban Transformation] URBAN TRANSFORMATION AND QUALITY OF LIFE IN BURSA DOĞANBEY

Miray Gür1 and Neslihan Dostoğlu2

1Uludag University, Faculty of Architecture, Dept. of Architecture, Bursa 2Istanbul Kültür University, Faculty of Architecture, Dept. of Architecture, İstanbul

Corresponding email: miraygur@yahoo.com SUMMARY Urban transformation has been playing a significant role recently in the planning policies of Turkey, especially in relation to developing solutions for deprived regions. As an intervening process and action that paves the way for the reconstruction of cities, transformation has the most significant impact on the users of the area. Therefore, monitoring the changes in the quality of life can be helpful in shaping the transformation policies for their suitability to sensitivity, flexibility and efficiency concepts, and also in producing considerable data. In this paper, Bursa Doğanbey Urban Transformation project is discussed from this perspective, by the quality of life scale developed in order to determine the effects of transformed environments on the quality of life. In relation to the importance of the right to choose for increasing quality of life, this paper discusses the interaction among different dimensions related with urban transformation from the perspective of user participation in Doğanbey Project, and deductions are made for the improvement of quality of life based on the results.

INTRODUCTION Urban transformation has played a significant role in the planning policies in Turkey over the past decade, in an attempt to develop solutions for the deprived regions of Turkey. The importance of the policies and implementations within this scope has increased with the Law. No 6306 about transforming the areas under disaster risk. It can be observed that the transformation implementations, which demolish and regenerate the living quarters of people, develop in the direction of the decisions made centrally, apart from the preferences of people. In this context, urban transformation which is a comprehensive action providing the configuration of city parts, in an expectation of sustainability, appears as an intervening action having long-term effects on the environment and both on the area and city users.

As seen in different cities, urban transformation that creates serious changes in the environment and public spaces in the city, is influential mostly on the social structure. In fact, transformations remodeling the area within the frame of dimensions related to the physical environment, finance, infrastructure, transportation, affect the quality of life of the users. In this context, long-term effects emerge in the lives of users that can not be seen directly at the beginning of the process until the use phase. This paper analyzes urban transformation as experienced at Bursa Doğanbey Implementation, related to its effects on the quality of life in the area. Doğanbey, which is an example of the transformations that are redefined with Law No. 6306, is an interesting case for monitoring the effects of the organizational actors on users, and shows how transformation can change the quality of life.

THE RELATIONSHIP BETWEEN URBAN TRANSFORMATION AND QUALITY OF LIFE Urban transformation, which has social, environmental and economical dimensions, remodel the city parts at the end of a process that progress in the direction of decisions made in the legal-administrative platform. The underlying purpose of urban transformation is to revitalize and reconfigurate the areas that have become neglected due to different socio-economic conditions and urban deprivation. Thus, healthy and livable conditions are expected through demolishing and reconstructing [1]. Urban transformation aiming to make the area utilizable, pursuant to current necessities is an extensive and integrated vision and action that seek to bring permanent improvement to the economic, physical, social and environmental conditions of the area, and to provide solutions for urban problems [2]. In this context, it should be emphasized that urban transformation must be designed so as to be responsive to the necessities and expectations from the future. The sustainability in the area to be transformed is related to the balance among some factors. These factors, as summarized by Turok [3], create extensive opportunities and welfare level for people, strengthening local business opportunities and employment performance, and increasing the attraction of the area in order to be preferred by people and companies that have the opportunity to choose the place to settle.

In this context, it is important that the community has the chance to choose the place to live and the area is reconfigurated properly to their life style. The factors that enable urban transformation to succeed are related to the increase in the quality of life in terms of social and economic dimensions as well as physical environment with the contribution of user participation. In this context, the physical dimension of urban transformation that is a multidimensional action and process, is related with design and planning from subscale to upper scale, economic dimension, related to financing and employment opportunities, and administrative and legal dimensions related to the actors partaking in the legal framework and decision mechanisms. As well as neighborliness, safety, accessibility to public services etc., which directly concern area users, are also included in the social

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dimension. Within the context of associating quality of life to urban transformation, user participation, which is part of the administrative dimension, will be analyzed in this paper.

QUALITY OF LIFE FROM THE PERSPECTIVE OF SUSTAINABILITY AND URBAN TRANSFORMATION Quality of life, consisting of many components, is an issue that is constantly investigated by specialized indicators in different disciplines, such as environment, economy and health, as well as being a multidisciplinary research area. Quality of life surveys are kept up-to-date because people’s welfare level is important for many sectors. Literally quality of life can be defined as a personal judgment that is reached by evaluating the characteristics of an individual’s life and the different dimensions of the environment according to his/her own experience and perception [4, 5, 6]. If focused on the quality of life in relation to the environment, environmental quality of life emerges as a more specialized area, and individual components constituting environmental quality of life contain the individual’s environmental perception alongside the perception of individual life. Quality of life is closely related to sustainability as a concept involving components of livability, justice and sustainability [7]. It is possible to mention quality of life in policies developed integrally with economic, environmental and social dimensions in sustainable communities. While sustainability evaluates person-environment relationship by referring to the future, quality of life and livability are focused on the “here and now” [5]. The concept of quality of life has enhanced its influence in direct proportion to the increasing importance of sustainability and studies within the scope of sustainability. One of the basic factors for sustainable development is using natural sources wisely in order to provide an infrastructure for long-term quality of life [6]. In this context, action plans should be constituted from high scale decisions and country policies to local scale. In this context, quality of life studies are favorable tools based upon scientific data in order to actualize sustainable implementations, providing benefit when evaluated holistically with transformation policies and implementations to create sustainable environment. The significant factor in urban transformation and urban quality of life is based on the related users taking advantage of the facilities generating increase in the value of the area. Analyzing the quality of life in urban transformation would provide more planned and efficient usage of the sources that organizational components have, by determining the effects of the policies on users and the city. Thereby it could be possible to make investments more wisely and equitable and to increase the wealth level of the users.

METHODOLOGY OF MEASURING QOL in DOĞANBEY URBAN TRANSFORMATION Bursa Doğanbey Urban Transformation Implementation Doğanbey Urban Transformation Implementation which started with the protocol made in 2006 among TOKİ (Mass Housing Administration), Bursa Metropolitan Municipality and Osmangazi Municipality, involve four quarters called Doğanbey, Tayakadın, Kiremitçi and Kırcaali. No renovations took place in the area until the decision of Osmangazi Municipality to implement the project because the area is one of the oldest settlements of Bursa, where there are a vast number of property owners due to reasons such as death, inheritance, individual joint-owned sales, etc. [7]. Following the decision that public authority would play an important role in organizing the transformation process, the municipality decided to modify the plan, ownership and function in accordance with the transformation model that they designed by means of the involvement of public authorities because they believed that partial and multiple ownership in the area, legal reasons, social problems, housing rights would make implementations in the area difficult. The area where transformation took place is a special region including monumental buildings and examples of civic architecture registered by the Bursa Cultural and Natural Heritage Preservation Board. Prior to demolishment, there were single or two-storey buildings in the area in general. Registered buildings on the site, planned to be transformed, and the adjacent heritage site made the project different from many attempts of transformation, and required planning decisions to consider conserving past values and to be useful in different dimensions in the long term. However, a number of the abovementioned examples of civic architecture were demolished during the process. In the area, which was approximately 200.000 m2, 96% of the ownership was private and 3784 shareholders came up during the process where the number of stakeholders was known to be 1910. Based on the ownership structure in the project area, it was found out that 50% of the property in the area was less than 50 m2, and 80% was less than 100 m2. This fact affected the size and number of houses to be designed and resulted in the high density of the area after transformation. A numerical share model was developed in accordance with the goal of Osmangazi Municipality not to aggrieve owners in the area, and this became a transformation model affecting planning, design and economic dimensions, which constituted the essence of the implementation, thus directly affecting the city. The decision that affected the environmental design and economic situation the most was to make owners with a land share of 5 m2 and higher a homeowner and to give a completed house of 3 m2 per a land of 2 m2. In the numerical share model (Table 1), the size and number of houses given from 100 m2 on varied according to the groups of ownership.

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Table 1. Numerical Share Model [8] OWNERSHIP GROUPS TYPE OF DWELLINGS (m2)

Up to 50 m2 75 Up to 75 m2 112,5

Up to 100 m2 150

According to the information given by the authorities on the project design, TOKİ decided to build low blocks to the east of the area and high blocks to the west of the area on the grounds that it was compatible with the urban heritage site to the east of the area and the MIA (Central Business District) to the west of the area, and blocks were designed according to the gradually increasing heights between these two areas. The project designed by an architect contacted, was not found appropriate by TOKİ due to financial concerns and revised by TOKİ architects so as to add more housing units. The number of housing units was 2407 in the first project (Fig. 1), and the number reached 2729 in the second project, depicted in Fig. 2.

Figure 1. Layout plan of the first project

Figure 2. Layout plan of the project implemented in transformation

Figures 3-5. Views from Doğanbey

40% of the overall area consists of 3-4 storey buildings called A Blocks. The storey height increases towards the west, with 23-storey blocks being located to the far west. 2729 housing units in 3 different types (75, 112.5, 150 m2) were constructed by TOKİ and 2338 belonged to the owners of these units. While A Blocks consisting of 75 m2 houses are 3-4 or storied, the other blocks are all 23-storeys, B Blocks with 112,5 and 150 m2 flats; C Blocks with 75, 112,5 and 150 m2 flats, D Blocks with 150 m2 flats (Fig. 3-5). The density, which was 75-100 person/hectare in the housing area prior to transformation, was raised to 500 person/hectare afterwards. In short, the problem was that the planning decisions were made based on the numerical data rather than by focusing on people.

After a delay of 3 years, which highly upset the owners, housing units were started to be delivered in July 2012, the units being allocated according to the drawing of lots carried out in October 2010. Problems arose because users’ choices about the blocks or size of housing units were not taken into account, and at this stage, two different associations were established in order for the communication of the locals with the officials and the generation of solutions, however these associations did not achieve their goal in terms of user participation. Quality of Life Measurement in Bursa Doğanbey Urban Transformation The quality of life is shaped by the satisfaction of individuals according to choice and preferences, and this is where quality of life is associated with environmental design. In the production of the physical environment, the development of architectural approach in accordance with user preferences, in other words, bringing together architecture and user preferences resides in adopting participatory design method. Urban transformation related to environmental design, is included in dimensions of individual and social quality of life that can be intervened by organizational actors. Thus, in this study a quality of life scale including parameters specific for Doğanbey was developed with indicators in literature in order to identify dimensions of quality of life regarding the housing area and to determine the effects of transformed environments on the quality of life. In consideration of local characteristics, it is important to examine the transformation of Doğanbey from the perspective of user participation because the source of problems was based on the miscommunication between the officials and local users, and the resulting architectural decisions that created the current state. In the measurement of quality of life process conducted in Doğanbey, the participation dimension was analyzed in addition to the characteristics of houses and housing area, transportation-accessibility, neighbor relations, security, adherence, belonging, and economic components.

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Questionnaires and in-depth interview methods were used simultaneously in the measurement. Data obtained from questionnaires and in-depth interviews aimed to to assess the situation before and after the transformation. Since 1700 inhabitants had settled in Doğanbey at the time of study, a sample volume was proportionally distributed to the user group of 1700. The questionnaire was carried out with 325 participants, which was more than required, and an in-depth interview was conducted with 35 participants selected from those participants. There were a large number of owners who rejected to answer the questionnaires because many people were psychologically and adversely affected by the problems of the transformation process. The data from the questionnaires were analyzed by using SPSS 18, and in-depth interviews were evaluated by analyzing voice records.

EVALUATION In examining what the concept of quality of life meant to users, it was found out that most of the users perceived it as a whole, i.e. with individual characteristics, the physical and social environment, and economic dimensions. In Doğanbey, satisfaction with housing was found to be related to the satisfaction with drawing lots, which directly affected the quality of life. The users’ social structure, needs, requests and expectations were disregarded during the allocation of flats by drawing lots. In fact, they did not have the right to choose the sizes, location or number of rooms of their houses and therefore, the satisfaction and quality of life decreased. Most of the users expressed that their new house after transformation did not conform to their life style although the rate of users who were satisfied with the number and size of rooms, climatic comfort, etc. parameters was higher than those who were not satisfied.

Although those, who were the most satisfied with the overall size of the flat, were the users who lived in the largest flat of 150 m2, the highest general level of satisfaction was among the owners of middle sized flats of 112,5 m2. On the other hand, both the general level of satisfaction and the satisfaction with the size of small flats of 75 m2 was low. Thus, it can be concluded that the flats of 112,5 m2 appear to be the most suitable housing units for the social structure and the life style of Doğanbey dwellers, while the flats of 75 m2 appear to be the most inconvenient ones. Differences in satisfaction with sizes are associated with different income levels and social structures in the blocks. In this context, in C Blocks consisting of flats of 75, 112,5 and 150 m2, differentiation in user groups due to varying sizes of the units decrease the satisfaction. Although the level of satisfaction with flats of 75 m2 is low, the level of satisfaction with A Blocks, which consist of such flats, is not poor, which can be an indication that the level of satisfaction is positively affected because these blocks, which are low-rise, comprise people with similar life styles.

In regard to the quality of construction, nearly 70% of users are dissatisfied both with the workmanship and materials used. During the field study, participants mentioned that the most important problems are the construction quality and the unqualified workmanship. It was observed that many users had to have their houses repaired, and almost all participants indicated that although officials had promised that the housing units would be built with qualified materials at the beginning of the transformation process, their expectations in this regard were never met. Thus, it would be correct to conclude that using low-quality materials for construction to reduce the costs has adverse effects on satisfaction and quality of life, and causes problems during the utilization phase (Fig. 6-8). It was observed that since almost all of the users were dissatisfied with the construction quality and since the implementation was much below their expectations, the quality of life was adversely affected.

Figures 6-7. The workmanship problems that reduced construction quality

Figure 8. The uncared areas in the blocks

In this survey, it was seen that well-kept housing area is important for environmental quality and quality of life, and that satisfaction with the care of gardens and roads in Doğanbey is low. While the opinion on the satisfaction with green spaces and care of the area affect the perception towards the physical appearance of the area, architecture of the housing area is also important. The housing environment, which was fully changed after transformation, had adverse effects on belonging and conformance to life style. Various analyses have revealed that users who are disturbed by the physical appearance of Doğanbey are higher than 70%, and many users express their dissatisfaction with the height and proximity of blocks, or the difference of height among some blocks. In fact, more than 80% of participants have underlined that the area has an adverse influence on the silhouette of the city. The results indicate that Doğanbey, which contains registered buildings and is located adjacent to the urban heritage site and in close proximity to the historical Khans Area, which has been included in UNESCO World Heritage List in 2014, is an unfavorable irreversible implementation for the inhabitants, the city and the city dwellers in

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Bursa. Therefore, it can be stated that the transformation of Doğanbey has had adverse effects on the quality of life of users, and urban quality of life for city dwellers. Within the social dimension of quality of life, neighbor relations and perception of security are associated with each other in connection with the changed physical appearance. The physical environment fully changed in comparison to the previous state in Doğanbey. The social life of users and whether they could adapt to the new environment were disregarded. At present, Doğanbey has become a housing area of high blocks and the concept of neighborhood has been forgotten. Vertical life in blocks of apartments does not allow neighbor relations like the horizontal site organization before transformation. Before transformation, neighbor relations formed the life style as a reflection of neighborhood culture, but after transformation, the rate of the participants who reported that they saw each other everyday decreased from 80% to 5%. Similarly, the rate of participants feeling adhered and belonging to the neighborhood and neighbors significantly decreased after transformation. In addition to this, perception of security significantly decreased in the area after transformation, which is associated with problems of architectural design, decreased neighbor relations and inadequate illumination of roads, having a direct effect on the quality of life.

In examining the process from the perspective of user participation, all decisions in various stages were made in collaboration with organizational actors and they were capable of affecting the users’ quality of life directly. The users stated that they were not given the right to speak at any phase of the transformation process and their opinions were not taken into consideration by the authorities and therefore more than 80% of the participants expressed that their satisfaction was adversely affected. More than half of the participants were dissatisfied with the lot drawing and they were not satisfied with the fact that they could not choose their flats. The results in Doğanbey where the trust of people in local administration significantly decreased, reveal that the deficiency in user participation has adverse effects on satisfaction and quality of life. CONCLUSION Urban transformation from the perspective of quality of life can be discussed in relation to user satisfaction and environmental aspects. Since the quality of life is associated with the environment and satisfaction with choices in this context, it can be indicated that an urban transformation process which does not include user participation will not be successful. In Turkey, organizational actors are dominant in the law shaping the current urban transformation implementations, while a framework regarding user participation is not defined. The critical lesson to learn from Doğanbey Urban Transformation process is that users’ quality of life is adversely affected and trust in the local administration decreases when the transformed environment does not physically and socially conform to their life style.

In the context of improvement of quality of life, it is in fact an invaluable opportunity to demolish and reconstruct an urban area through urban transformation. Since urban transformation is a future action, short-term plans should not be generated; goals should be set well by all stakeholders, from different industries, academic institutions, NGO’s, and area users, and political infrastructure should support this; the process should be incrementally planned and developed focusing on people, protecting local and urban identity, improving social integration; and providing new business and economic opportunities. Meeting fundamental needs is not sufficient for urban transformation and implementations should be planned considering the benefits they will yield in the long term.

As for generalizations over the example of Doğanbey, it can be concluded that it is possible to succeed in increasing quality of life by analyzing the region in detail technically and socially in the process of making decisions, determining user opinions and trends and forming design parameters accordingly and enabling the right to choose in urban transformation. In addition to this, taking into consideration the effect of the region on the city is important, as well as, foreseeing density, public equipment, connections, and decision-making not only by organizational actors, but also with the participation of the local community. It is important to provide such an infrastructure because future cities to be shaped by urban transformation should embrace the society and are expected to improve the quality of life of the society.

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Dergisi, TMMOB Mimarlar Odası Antalya Şubesi 2. Roberts P, Sykes H. 2000. Urban Regeneration A Handbook, Sage Publications. 3. Turok I. 2004. Kentsel Dönüşüm: Neler Yapılabilir ve Nelerden Kaçınılmalı?, Uluslararası Kentsel Dönüşüm Uygulamaları

Sempozyumu, Küçükçekmece Belediyesi Yayını, İstanbul. 4. Szalai, A., 1980. The Meaning of Comparative Research on the Quality of Life. The Quality of Life: Comparative Studies,

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Landscape and Urban Planning, 49: 163-178 8. World Commission on Environment and Development 1987. Our Common Future, Oxford: Oxford University Press. 9. Osmangazi Belediyesi. 2008. Yeni Merkez Uygulama İmar Planı Revizyonu Değişikliği Plan Açıklama Raporu (New Center

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Raporu, Bursa: Osmangazi Belediyesi.

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[Abstract:0013][Economy of Energy and Environment] ENERGY SAVING APPLICATIONS IN FLUE GAS INSTALLATIONS

Muammer Akgün

BACADER, Baca İmalatçıları ve Uygulayıcıları Derneği muammerakgun@bacader.org

SUMMARY Energy saving has been an important issue from the decreasing of the world’s energy resources, the pollution of the environment and the global warming. Because of the fact that energy saving is really important and the fuel in our country is uneconomical, the energy loss in flue gas installations should be considered and prevented.If the external ambient temperature is colder than the boiler room the flue gas installation draws air from the heating device. This situation occurs more frequently if the burner stops or is in stand-by mode. It draws heat energy from boiler or its heat exchanger. It decreases the temperature of the water in the boiler or heat exchanger. Depends on external ambient temperature. The density difference and the chimney draught increases as the external ambient temperature is colder.These kinds of losses can be reduced by correct design to type of fuel and special equipment. There are different systems designed for different devices such as secondary air appliances, thermal and motor-driven flue gas dampers. In this paper, the equipments for energy saving in flue gas installations, their areas of utilization, advantages and applications are explained. INTRODUCTION In this essay, the methods which are appropriate for energy conservation in the chimney and waste gas systems, the advantages and the disadvantages of utilising these equipments will be evaluated. 1- FLUE GAS REGULATOR AND DUST CATCHER

Waste gas regulator is connected to the waste gas exit or vertical connection part to increase the flow resistance on the waste gas way. Waste gas regulator does not cover the radius of the flue 100 percent. At least 3 percent of the area or 20 cm2 should be space. The regulator should be visible with the help of the adjusting machine that is located near the regulator. The waste gas regulator can be utilitized with the natural gas using devices that works with fan if they are proved to be safe(with a type test or an expertise report). Dust catcher is connected to the flue horizontally or vertically and makes full impereability in the flue. Dust catcher can be utilized with the devices that works with stiff or liquid fuel. 2- FLUE DAMPERS

Flue gas dampers are connected to the vertical connection part of the flue or to the exit of the boiler, to close the gas way. Thus it hinders the device to cool down as it is in stand-by mode. Multiple devices that are connected blocks the access of cold air to the devices that are excluded and by that, enhances the ejaculation of the waste gas. Moreover it contributes to the decrease of the sound and the vibrations that can occur through the flow direction from the boiler room.

These valves can be implemented with; Liquid/Gas fuel using devices with fans Gas fuel using devices without fans Open fireplaces which burned wood or gas

The limitation of the air flow to the waste gas walves should only be enough to dry the wet parts which are due to the condensation that occurs when the device is off. The thermal resistance of the flues or horizontal was gas system should be enduring to moist and 0.65 m2K/W. Also, the air supply should be made with the help of secondary air appliances if the was gaste valve is off. There are two types of flue dampers:

Thermal flue dampers Motorised flue dampers

2.1- THERMAL FLUE DAMPERS Thermal flue dampers are used in line with the devices that are for DIN 3388-4. These dampers can be erected according to the flow resistance. They work with bimetal. The flue damper takes action with the heating of the waste gas or the cooling of the environment that is around the bimetal. The damper turns on as the burner steps in and the valve actively starts to work as the temperature reaches to 80-100 C. These dampers can pose danger if the annual maintenance is not made or if there is fatigue of materials.

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Fig. 1- Thermal Flue Damper

2.2. MOTORISED FLUE DAMPERS The motor driven flue dampers are erected to fuel exit of the gas/fuel oil using heating systems. The lid that is controlled with the motor closes down the flue gas port as the heating systems gets into the standby mode. Thus the energy that is produced in the system is stored and the unnecessary coolings are prevented by hindering the gas waste with chimney draught and entrance of cold air. As this system lowers the fuel outgoing by energy conservation (which varies from system to system but is between 5-12%), it also lowers the occurance of faults in store equipment by protracting the heating system. Motor driven flue dampers start working with additional energy. The dampers start working with the electricity in a short amount of time (5-7 seconds) before the starting of the burner. The mechanism is controlled with a switch. Motor driven flue dampers are connected to the flue or should be connected to the front or the back of the working device if the device does not use a fan. If it is connected to the front, it is dependent to the internal design, the installation and the utilization of the device. The document that is for the safety of the valves is DIN 388-2. This standart categorizes the devices according to three different installations:

The minimum gap with the damper (the released diameter) Non-leaked dampers (the endurance to the metal) Secondary-air appliances are non-leaked dampers, with the essential setting system or thermal flue dampers.

100% non-leaked damper is only used with burner devices.

Fig.2- Motorised Flue Dampers

2.3. SECONDARY AIR APPLIANCES Secondary-air appliances are valves that provides air with the mechanism that they have inside. They are used for the increase in the condensation speed of the flue gas with the entrance of additional air. The independent operating secondary-air appliances are also known as draft regulator. Draft regulator also keeps the burning in an average value for the differences in air intake. Moreover, there should be an air circulation for the entrance of additional air to the flue since there can be a pressure difference between the boiler room and flue. According to DIN 4795 the flues that are negative pressure designed can be utilized only if they implement these rules;

The waste gas road should be open There should be no danger incase of an occlusion or leaking out in the flue It should not handicap the cleaning or the maintenance of the waste gas system

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The flue should be designed in a way that is enduring to the corrosion and able to block any leaking out due to the condensation.

Fig.3- Operating principle of secondary air appliance

Secondary air appliances can be utilized in boiler rooms or the rooms which has air flow with the room next to it. All devices should be in the same boiler room if there are multiple connections in the flue gas system. Seconder air valves can be grouped into three:

Independent seconder air appliances Seconder air appliances that mechanicly adjusted Combinated seconder air appliances

The upwards pressure of the waste gas in draught limiter which is an regulations made specifies adjustment weight. The weights of openings and the closing cause an equilibrium. Because of this equilibrium, the hot waste gases of the waste gas installations gets mixed with the cold air automatically. Inflowing air through transpiration so the formation of dew point tend to be reduced. If the combustion unit stops additional to the upwards pressure above the adjustment valve of draught regulator, the ventilation of the system is made with the air in the room. However this drying effect decreases with upward pressure and gets cut out completely if the adjustment value of the draught regulator is not reached. The air flap is opened in an appropriate amount with the help of the excitation current running motor when the adjustable working secondary air flap is deactivated. Motorised damper is placed on the mounting or installation of the exhaust side connection part or applications. Motorised dampers need more pressure for larger flue gas installations. This requires additional air assembly.

Fig. 4- Self-employed secondary air appliance. When the burner stops in combined systems, it makes the complete ventilation. The ventilation duration can be limited to 10 minutes according to the choice engine control. If the heater should be on first the motor control starts. Than the setting weight releases the adjustable disk. Thus the upwards pressure can be adjusted. However the ignition switch is release when motor control reaches final position. Forced control and combined systems are not suitable for solid fuels burning devices. The waste gas temperature between the device and installation because of secondary air appliances. Also condensation temperature decreases and the speed of waste gas increases. This dries the chimney. If the draught regulator works in an inappropriate situation there too much waste gas lose and there may be some difficulties with the optimal adjustment of the burner with fan.The draught regulator provides optimum working conditions for burning by increasing pressure.

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Figure 5. The secondary air appliance with adjustable mechanical flue gas damper.

2.3.1. SELECTION AND SIZING The sizing of the secondary air appliance by the manual which is determined by the manufacturer and according to height, diameter, structure of the chimney and capacity of the device. The amount of additional air (m3/h)=A x H Qmax [kW] A: Chimney cross-section [m2] H: Heigth [m] Qmax: Chimney thermal capacity [kW] The draught regulator should have at least 40 cm from the floor when it is attached to the chimney. It should not endanger the fire safety of chimney. There should be at least 40 cm. between the secondary air appliance which are 2,5 meters above the ground and combustible materials. Secondary air appliance should be positioned differently according to the requirements. Issues that must be considered according to DIN4795;

Fig. 6-. Types of draught regulator assembly

(Installation of additional air assembly facilities). Position 1: Very good regulation, ventilation effect limited in long flue gas pipes or small flue gas pipe cross-section Q1 in relation to chimney cross-section Q2. Position 2: +3: Very good ventilation effect, good regulation, only retrofitable at installation place 3 in brick-built chimneys. Position 4: and ventilation limited. However, due to the low accumulation of soot this is a good place of installation for solid fuel boilers and lined chimneys. 3. DAMPERS FOR LARGE PLANTS in large installations draught limiter with pressure flap and hydraulic damper for the setting weight which are for chimneys that are higher than 20 meters must be used for operation of the installation (Fig.7). The speed of waste gas occurs importance during the selection of secondary air appliance. When the ignition stops, the flow of waste gas stops and the sub pressure at the waste gas installation increases quickly. As a result, draught limiter works with pulse. This may cause deformation on chimney. The draught limiter should have minimum waste gas speed. In the summer and in the transition period is upwards pressure is low and adjustment disc is close because of that. However, when the burner gets operated for long time it reaches the set value of the draught limiter, thus the adjustment disc opens. Released cross section grows with the steadily rising up the pressure.

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Fig.7- Draught limiter for large installations (Hydraulic Shock Absorber).

4. CONCLUSION The equipment for flue gas systems are used to prevent deformation at the chimney and recover the losses which may occur in the flue. The selection and the utilization of these systems for solution are important. When the payback period is considered, thermal valves pay off their prices in two heating seasons. Motorized dampers provide fuel savings of 5-12% per annum based on the applied system. Draught regulators vary according to the amount of annual fuel savings chimney pressure. Also the draught regulators are essential for the prevention of deformation which may occur in chimneys if the burners in the central heating system suddenly close. REFERENCES [1] POSTRENRIEDER, E. and SCHLEE, G., Abgasanlagen, Gentner Verlag, Stuttgart. 1992. [2] SCHAFER, W., Schornsteinfragen in der heizungstechnik, Kramer Verlag, Düseldorf. 1994. [3] MILTON Keynes., Mechanical venting of flues and chimneys, Exhausto Ltd., 2000. [4] Schiedel GmbH & Co., Schornsteintechnik, München. 2001. [5] DIN-Taschenbuch 146., Schornsteine planung berechnung ausführung, Beuth. 1993. [6] Kutzner Weber GmbH & Co., Abgastechnische produkte geratebau, Maisach. 2000.

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[Abstract:0015][Modeling and Software]

DETERMINING THE OPTIMUM FINANCIAL AND ENERGY SAVING SOLUTIONS FOR BUILDINGS IN DIFFERENT CLIMATES

Aslıhan Şenel Solmaz Dokuz Eylul University, Department of Architecture, Izmir, Turkey

Corresponding email: asenelsolmaz@gmail.com; aslihan.senel@deu.edu.tr SUMMARY

In this study, a simulation based optimization approach, the combination of building energy performance simulation tool and an optimization program, is employed to assess energy saving alternatives and concurrently optimize building heating and cooling energy savings and a financial measure, Net-Present Value (NPV). A hypothetical building is used to illustrate the applicability of the approach in different climate zones in Turkey. Sketch-up Open Studio plug-in is utilized for building energy modelling. EnergyPlus dynamic simulation tool and a GenOpt optimization package are used together for the execution of the approach. The optimization algorithm does the search within a wide solution space including multiple alternative materials ranging from external wall insulation to different window types. The optimization results clearly show that the interaction between the conflicting objectives and their trade-off relationships should be analysed while determining the optimum energy and cost saving solutions for buildings. INTRODUCTION Determining the optimum solution for building energy efficiency is a vastly complex process because it requires taking into account a large number of solution alternatives to optimize multiple criteria (e.g. energy, environment and cost) simultaneously [1]. The criteria mostly conflict with each other and the optimum solution is a trade-off among them. The other main difficulty is the great number of design variables [2] ranging from the geometry to the selection of energy efficient building materials; hence a large solution space is required for finding an optimum solution. Today, building performance simulation is the most common method used for developing building energy saving solutions. Using these tools alone is naturally time consuming and the search may only result in partial performance improvements due to searching within the limited group of solution alternatives. To find an optimal solution to a problem with much less time, decision-support approaches that combine building performance simulation with optimization engines are needed [3]. In this study, a simulation based optimization approach that integrates a simulation program with optimization tool is presented to identify the optimal building energy efficiency solutions. METHOD In this study a simulation based optimization method is used to optimize both building heating and cooling energy consumption while minimizing NPV. The aim of the GenOpt is to minimize an objective function extracted from external simulation programs [4]. GenOpt can be integrated to all simulation programs that work with text file (.txt) output. In this study, a validated dynamic simulation program, EnergyPlus is selected to calculate building energy consumptions. GenOpt optimization program is selected because it successfully converges to global optimum solutions and can give close enough results to brute-force approach [5]. GenOpt is defined as kernel in which many algorithms are integrated inside. In this study, a population based meta-heuristic algorithm, Particle Swarm Optimization (PSO), is chosen for optimization because this study requires assigning discrete values to input parameters in this study [4]. The general steps for applying simulation based optimization method in this study are listed below:

1. Create a base-case building energy model with Sketch-up Open Studio and convert it to EnergyPlus input file (.idf) for template.

2. Identify the energy efficiency solution alternatives and their related data (material thermo-physical properties, thickness, unit cost etc.)

3. Define an objective function with three objective criteria (heating energy consumption, cooling energy consumption and NPV)

4. Run optimization program and get optimal solutions per defined objective function.

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CASE STUDY Building energy modelling A multi-objective simulation based optimization method is applied to a hypothetical office building to illustrate its applicability in different climate zones in Turkey. The case building energy model was created with Sketch-up Open Studio plug-in (Figure 1).

Figure 1: The hypothetical office building energy model created with Sketch-up Open Studio plug-in The case building is oriented north-south direction. There are eight thermal zones (six office spaces and two circulation areas) on each floor. It was assumed that the building was built before existing national standards, there is no insulation on the any part of the building and it needs energy retrofit. All windows with PVC frames have single glazing with high SHGC value (0.82) and there is no shading component on any façade of the building as well. The building has central system for both heating and cooling, and the thermostats set points are 22°C and 26°C respectively. The base-case building was analysed in both Izmir and Ankara cities that represent the hot-humid climate and temperate-dry climate regions of Turkey respectively in order to show the effects of different climate conditions on building energy analysis. The annual heating and cooling energy consumptions of the building are calculated 30100.69 kWh and 38581.03 kWh in Izmir and 82000.48 kWh and 16737.61 in Ankara respectively. Identification of the energy efficiency solution alternatives In this study, the focus is on the building envelope to generate energy efficiency measures. The selected material alternatives and their thermo-physical properties and unit costs are presented on Table 1-2-3. Table 1. Generated energy efficiency solution alternatives regarding insulation materials: roof, exterior wall and ground floor [6].

Envelope Component Material Name ID Thickness

(mm) Conductivity

(W/mK) Specific

Heat Density (kg/m3)

Cost (TL/m2)

ROOF (A-B)

XPS Extruded Polystyrene Foam Board (A)

A1-A10

20-25-30-40-50-60-70-80-90-100

0.035 1500 30 4.64-25.60

Glass Wool (B) B1-B7 80-100-120-140-160-180-200

0.040 840 14 3.32-8.40

EXTERIOR WALL (E-F-G)

Rock Wool (E ) E1-E7 30-40-50-60-80-100-120 0.037 840 150 6.15-

24.53 EPS Expanded Polystyrene Foam Boar (F)

F1-F9 30-40-50-60-70-80-100-120-140

0.039 1500 16 2.65-12.25

XPS Extruded Polystyrene Foam Boar (G)

G1-G8 30-40-50-60-70-80-100-120 0.035 1500 30 5.0-23.0

GROUND FLOOR

(H)

XPS Extruded Polystyrene Foam Boar (H)

H1-H10

20-25-30-40-50-60-70-80-90-100

0.035 1500 30 4.64-25.60

Table 2. Generated energy efficiency solution alternatives for window types [6].

Envelope Component Material Name ID U Value

(W/m2K) SHGC Vis. Tran.

Cost (TL/m2)

Single Glazing, 4mm C1 5.2 0.87 0.9 23.5

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WINDOW (C)

Low-e single glazing, 4mm C2 4.2 0.65 0.79 26.5 Tinted single glazing, 4mm C3 5.2 0.54 0.71 25.5 Tinted low-e single glazing, 4mm C4 4.2 0.54 0.71 28.0 Clear double glazing, air-filled, 4-12-4mm C5 2.9 0.75 0.8 36.0 Clear double glazing, air-filled, 4-16-4mm C6 2.7 0.75 0.8 36.5 Clear double glazing, argon-filled, 4-12-4mm C7 2.7 0.75 0.8 37.5 Clear double glazing, argon-filled, 4-16-4mm C8 2.6 0.75 0.8 38.0 Low-e double glazing, air-filled, 4-12-4mm C9 1.6 0.56 0.79 38.0 Low-e double glazing, air-filled, 4-16-4mm C10 1.3 0.56 0.79 38.5 Low-e double glazing, argon-filled, 4-12-4mm C11 1.3 0.56 0.79 39.5 Low-e double glazing, argon-filled, 4-16-4mm C12 1.1 0.56 0.79 40.0 Tinted low-e double glazing, air-filled, 4-12-4mm C13 1.6 0.44 0.71 40.0

Tinted low-e double glazing, air-filled, 4-16-4mm C14 1.3 0.44 0.71 40.5

Tinted low-e double glazing, argon-filled, 4-12-4m C15 1.3 0.44 0.71 41.5

Tinted low-e double glazing, argon-filled, 4-16-4mm C16 1.1 0.44 0.71 42.0

Clear triple glazing, air-filled, 4-12-4-12-4mm C17 1.1 0.73 0.78 43.0 Clear triple glazing, air-filled, 4-16-4-16-4mm C18 1 0.73 0.78 44.0

Table 3. Generated energy efficiency solution alternatives regarding shading materials of windows [6].

Envelope Component Material Name ID Length (m) Cost (TL/m2)

SHADING (D) Horizontal fixed overhang D1-D9 0.2-0.3-0.4-0.5-

0.6-0.7-0.8-0.9-0 30

Define an objective function In this study, three objective criteria, heating energy consumption, cooling energy consumption and NPV, are used in the multi-objective optimization process. In order to integrate these three objectives into GenOpt, a “weighted-sum” approach is used. According to this approach, each criterion has an assigned weight coefficient in the optimization process. The objective function is shown in Eq. 1:

(1)

where a, b and c are the weight coefficients of heating energy consumption, cooling energy consumption and NPV respectively. Each objective formula is entered into the relevant GenOpt input file. Heating and cooling energy consumption data are gathered from EnergyPlus, NPV is calculated according to Eq. 2 [6]:

(2)

In Eq. 2, i is the nominal discount rate, t is the duration of the cash flow, Rt is the net cash flow at time t including inflation rate for the energy prices hikes. The NPV is calculated for 10 years with 4.5% nominal discount rate and 10% inflation rate. After setting the objective function within the GenOpt, several optimization runs are made for both Izmir and Ankara separately. RESULTS Two different optimization runs were done for Izmir to get optimal solutions for heating and cooling energy savings separately. It is done by adjusting the weight factors of each criterion. The assigned weights for each criterion on different optimization runs are given below:

Optimization for maximum heating energy saving: a=20, b=1, and c=-0.05 Optimization for maximum cooling energy saving: a=1, b=10, and c=-0.1

f = a f1 + b f2 + c f3

NPV R t

(1 i)tt1

N

IniInv

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As mentioned before, a, b and c are the weight coefficients of heating energy consumption, cooling energy consumption and NPV on objective criteria respectively. So firstly, in order to get an optimal solution for maximum heating energy saving, coefficient of heating saving (a) is given 20, while the smaller values are given to other coefficients. Optimal heating and cooling energy saving solutions for Izmir The optimization results for maximum heating and cooling energy saving are presented in Figure 2a-2b separately. On Figure 2a-2b, the individual objectives are plotted against each other with NPV on y-axis, cooling savings on x-axis and heating savings on z-axis with colorbar. According to Figure 2a-2b, there are positive and negative correlations among objective criteria. For example, while cooling savings are negatively correlated with heating savings, they are positively correlated with NPV. Similarly, there is a negative relationship between heating savings and NPV. This means, solution alternatives to obtain more heating energy savings reduce the NPV. On the contrary, the solutions to obtain more cooling savings affects NPV positively because of their relationship. Three data points marked on both Figure 2a-2b represent the maximum savings related to heating, cooling, and NPV with their values. The maximum NPV points on both Figures 2a-2b are near to maximum cooling energy saving points while they are far away from maximum heating energy saving points.

a) b) Figure 2: Optimization results for Izmir a) results for maximum heating energy saving, b) results for maximum cooling energy saving. On Figure 2a, compared to the base-case scenario, the maximum heating saving is 61.36 % with 77.94% in cooling loss and 72424.4 TL NPV loss at the end of a 10-year period. On the same figure, the maximum cooling saving is 10.61% corresponding to 39.24% heating saving and 54905.93 TL NPV gain at the end of 10-year period. On the maximum NPV point on Figure 2a, 10-year NPV gain is 57670.55 TL, the heating saving is 48.71 % and the cooling saving is 4.38 %. On Figure 2b, compared to the base-case scenario, the maximum cooling saving is 34.90% corresponding to 20.18% heating saving and 67247.25 TL NPV gain at the end of 10-year period. The maximum heating saving is 57.44 % with 46.33% in cooling loss and 24147.3 TL NPV loss at the end of 10-year period. On the maximum NPV point on Figure 2b, 10-year NPV gain is 74914.67 TL, the heating saving is 41.31 % and the cooling saving is 18.62 %. Each data point on Figure 2 has a set of energy efficiency solution alternatives. A set of solution alternatives of each data point on both Figure 2a and Figure 2b are shown in Table 4 under the title of “Result I” and “Result II” separately. According to Table 4, Result I shows the optimization results for maximum heating energy saving and Result II shows the maximum cooling energy saving results. When looked at the set of alternatives on Result I-, the window material with ID C18 (see Table 2) that have low U-value and high SHGC was assigned to all windows of building in order to increase heating saving. Parallel to this selection, the shadings were not assigned to any windows (D9). The insulation material with the highest thickness (H10) was assigned to ground floor and the insulation material with the highest thickness (B7) was assigned to roof to obtain maximum heating energy saving. As for the exterior walls, the insulation material with lowest conductivity and highest thickness (G8) was assigned to walls on north, east and west directions, and the wall insulation with ID of F9 having 140mm thickness was assigned to wall on south direction. This set of solution is the rational choice for increasing the heating energy saving.

Cooling Energy Saving (%)

Net

Pre

sent

Val

ue (N

PV) (

TL)

Heating Energy

Saving (%)

Max. Cooling Saving Point HS: %39.24 CS: %10.61 NPV: 54905.93 TL

Max. Heating Saving Point HS: %61.36 CS: -%77.94 NPV: -72424.4 TL

Max. NPV Point HS: %48.71 CS: %4.38 NPV: 57670.55 TL

I zmir: Result I -Optimization for Maximum Heating Energy Saving

Cooling Energy Saving (%)

Net

Pre

sent

Val

ue (N

PV) (

TL)

Heating Energy

Saving (%)

Max. Heating Saving Point HS: %57.44 CS: -%46.33 NPV: -24147.3 TL

Max. NPV Point HS: %41.31 CS: %18.62 NPV: 74914.67 TL

Max. Cooling Saving Point HS: %34.90 CS: %20.18 NPV: 67247.25 TL

Izmir: Result I I -Optimization for Maximum Cooling Energy Saving

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Table 4. Optimal energy efficiency saving solutions both heating and cooling for Izmir

IZMIR Result I: Optimization for Max

Heating (a=20; b=1; c=-0.05)

Result II: Optimization for Max Cooling

(a=1; b=10; c=-0.1)

Building Envelope Components Max Heating Saving (%) Max Cooling Saving (%) 61.36 20.18

Window-South C18 C13 Window-North C18 C3 Window-East C18 C13 Window-West C18 C13 WindowShading-South D9 D8 WindowShading-North D9 D8 WindowShading-East D9 D8 WindowShading-West D9 D8 RoofInsulation B7 B7 FloorInsulation H10 H1 WallInsulation-North G8 F1 WallInsulation-South F9 E6 WallInsulation-East G8 G8 WallInsulation-West G8 F7

When looked at the set of solutions on Result II-maximum cooling saving solution, window material with ID C13 (see Table 2) that have lowest the SHGC and low U-value was assigned to windows on south, east and west directions, while the window material with ID C3 was assigned to north direction. Considering the negative impact of SHGC on the cooling energy savings, this selection is rational. Similarly, the shading alternative with ID of D8 having 0.9 m height was assigned to all windows on each direction. The insulation material with lowest thickness (H1) was assigned to ground floor and the insulation material with highest thickness (B7) was assigned to roof to obtain maximum cooling energy saving. On the exterior walls, while the insulation material with lowest thickness (F1) was assigned to walls on north direction, the different insulation materials with high thicknesses were assigned to walls on the other directions. This set of solution is seems rational for increasing the cooling energy savings. Optimal heating and cooling energy saving solutions for Ankara The optimization results for maximum heating and cooling energy saving are presented in Figure 3a-3b separately. On Figure 3a, compared to the base-case scenario, the maximum heating saving is 49.68 % with 96.73% in cooling loss and 98091.9 TL NPV savings at the end of a 10-year period. On the same figure, the maximum cooling saving is 19.61% corresponding to a 36.56% heating saving and 136982.5 TL NPV gain at the end of 10-year period. On the maximum NPV point on Figure 3a, 10-year NPV gain is 148014.3 TL, the heating saving is 40.45 % and the cooling saving is 14.69%. On Figure 3b, compared to the base-case scenario, the maximum cooling saving is 35.57% corresponding to 26.29% heating saving and 112907.5 TL NPV gain at the end of 10-year period. The maximum heating saving is 46.94% with 69.71% in cooling loss and 104439.8 TL NPV gain at the end of 10-year period. On the maximum NPV point on Figure 3b, 10-year NPV gain is 145553.0 TL, the heating saving is 42.24 % and the cooling saving is 5.66%.

a) b) Figure 3: Optimization results for Ankara a) results for maximum heating energy saving, b) results for maximum cooling energy saving.

Cooling Energy Saving (%)

Net

Pre

sent

Val

ue (N

PV) (

TL)

Heating Energy

Saving (%)

Max. Heating Saving Point HS: %49.68 CS: -%96.73 NPV: 98091.9 TL

Max. Cooling Saving Point HS: %36.56 CS: %19.61 NPV: 136982.5 TL

Max. NPV Point HS: %40.45 CS: %14.69 NPV: 148014.3 TL

Ankara: Result I -Optimization for Maximum Heating Energy Saving

Cooling Energy Saving (%)

Net

Pre

sent

Val

ue (N

PV) (

TL)

Heating Energy

Saving (%)

Max. Heating Saving Point HS: %46.94 CS: -%69.71 NPV: 104439.8 TL

Max. Cooling Saving Point HS: %26.29 CS: %35.57 NPV: 112907.5 TL

Max. NPV Point HS: %42.24 CS: %5.66 NPV: 145553.0 TL

Ankara: Result I I -Optimization for Maximum Cooling Energy Saving

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A set of solution alternatives of each data point on both Figure 3a and Figure 3b are shown in Table 5 under the title of “Result I” and “Result II” separately. According to Table 5, Result I shows the optimization results for maximum heating energy saving and Result II shows the maximum cooling energy saving. When looked at the set of solution on Result I-maximum heating saving solution, similarly the solution on Table 4, window material with ID C18 (see Table 2) having low U-value and high SHGC was assigned to all windows of building in order to increase heating saving. Parallel to this selection, the shadings were not assigned to windows (D9) except for the windows on east direction. Table 5. Optimal energy efficiency saving solutions both heating and cooling for Ankara

ANKARA Result I: Optimization for Max

Heating (a=20; b=1; c=-0.05)

Result II: Optimization for Max Cooling

(a=1; b=10; c=-0.1)

Building Envelope Components Max Heating Saving (%) Max Cooling Saving (%) 49.68 35.57

Window-South C18 C3 Window-North C18 C3 Window-East C18 C13 Window-West C18 C13 WindowShading-South D9 D8 WindowShading-North D9 D8 WindowShading-East D6 D8 WindowShading-West D9 D8 RoofInsulation B7 B7 FloorInsulation H8 H1 WallInsulation-North G8 G1 WallInsulation-South G7 F1 WallInsulation-East F9 E6 WallInsulation-West F8 E2

The insulation material with a high thickness (H8) was assigned to ground floor and the insulation material with the highest thickness (B7) were assigned to roof in order to obtain maximum heating energy saving. On the exterior walls, the insulation materials with almost the highest thicknesses (G7-G8-F9-F8) were assigned to walls. Looking at the set of solutions on Result II-maximum cooling saving solution, while the window material with ID C3 was assigned on north and south directions, window material with ID C13 (see Table 2) that have lowest SHGC and low U-value was assigned to windows on east and west directions. Similarly, the shading alternative with ID of D8 having 0.9 m height was assigned to all windows on each direction. The insulation material with the lowest thickness (H1) was assigned to ground floor and the insulation material with the highest thickness (B7) was assigned to roof in order to obtain maximum cooling energy savings. On the exterior walls, contrary to the result of Izmir, the insulation material with lowest thickness (G1-F1-E2) was assigned to exterior walls. DISCUSSION AND CONCLUSION By applying a simulation based optimization method, the optimal solutions for building envelope were developed for different climate zones in Turkey (Izmir and Ankara) by considering heating and cooling energy consumptions, and NPV criteria simultaneously. The optimization algorithm did the search within a wide solution space including multiple alternative materials for building envelope. The optimization results clearly show the interaction between the conflicting objectives and their trade-off relationships. For example, there is a trade-off between heating and cooling energy savings, and there is a negative relationship between heating savings and NPV. While there are losses from both cooling energy and NPV on the maximum heating energy saving points in Izmir, this is not valid for Ankara. This result possibly arises from the differences between climate characteristics of Izmir and Ankara. Cooling energy saving is positively correlated with NPV, so the maximum NPV points on both optimization results for Izmir and Ankara are near the maximum cooling energy saving points while they are far away from maximum heating energy saving points. It can be concluded that the optimization algorithm completely can find the suitable set of energy saving solutions for each city with economical returns. Finally, the interactions between the objectives should be analysed while determining energy and cost saving solutions for buildings.

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REFERENCES 1. Kolokotsa, D, Diakaki, C, Grigoroudis, E, Stavrakakis, G, et. Al. 2009. Decision support methodologies on the energy

efficiency and energy management in buildings. Advances in Building Energy Research. Vol. 3(1), pp. 121-146. 2. Evins, R. 2013. A review of computational optimisation methods applied to sustainable building design. Renewable and

Sustainable Energy Reviews. Vol. 22(0), pp. 230-245. 3. Nguyen, A T, Reiter, S, and Rigo, P. 2014. A review on simulation-based optimization methods applied to building

performance analysis. Applied Energy. Vol. 113 (0), pp 1043-1058. 4. Wetter, M. 2011. GenOpt – Generic optimization program user manual v3.1.0. technical report-LBNL- 2077E. U.S.

Lawrence Berkeley National Laboratory (LBNL). 5. Hasan, A, Vuolle, M, and Siren, K, 2008. Minimisation of life cycle cost of a detached house using combined simulation

and optimisation. Building and Environment. Vol. 43 (12), pp 2022-2034. 6. Senel Solmaz, A. 2015. Bina enerji performansını geliştirmede optimum çözümleri belirlemeye yönelik simülasyon ve çok

amaçlı optimizasyon tabanlı bir karar destek modeli. PhD. Thesis, Dokuz Eylul University, Izmir, Turkey.

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[Abstract:0017][Sanitary System Applications] ENERGY EFFICIENT CLEANROOM DESIGN – IS IT AN OXYMORON?

Nejat Babür, PE, LEED AP Anel Group, Istanbul Turkey

Nejat.babur@anel.com.tr

SUMMARY

Cleanrooms seen in Semiconductor and Pharmaceutical facilities are one of the largest energy users in advanced technology manufacturing facilities. Design criteria such as cleanliness levels, air changes, filtration levels, temperature and humidity conditions are driven by regulations and process needs. The risk of losing a batch of wafer or dumping a campaign of drugs prevents owners and designers from considering new design concepts. In cGMP driven facilities in particular, there is little acceptance for doing anything outside of common practices.

After process, HVAC is the largest energy consumer in advanced technology fabrication facilities. For example making educated design decisions can reduce the amount of air that is recirculated within the space, as well as outside air quantities, selecting the right air management concepts and using high efficiency equipment can reduce energy consumption. Reducing duct, pipe, filter and coil face velocities can reduce the pump and fan energy.There are multiple opportunities to reduce the energy consumption and carbon print of your facility, as long as risks vs. benefits are evaluated during design.

INTRODUCTION Green Building Counsel in the United States and similar organizations around the world as well as The American Society of Heating, Refrigeration and Air Conditioning Engineers (ASHRAE) and International Society of Pharmaceutical Engineers (ISPE) are focusing on reducing the energy consumption of the advanced technology facilities. Data Center owners and designers came up with a metric (Power Use Index – PUI) to benchmark and reduce the energy use of their facilities and challenge designers and owners to achieve better overall efficiency of power use. If we think along the same lines, facilities employing cleanrooms should be able to do the same. However, the challenge would be creating a metric for each type of facility since their process needs and cleanroom requirements would be different. In an aseptic fill facility or semiconductor fabrication building might need different cleanroom geometry. We cannot just take a design concept used for semiconductor and apply to a pharmaceutical facility. However, designers should be able to adapt the design concepts from one type to another type by using their engineering knowledge and experiences. This paper will approach the energy efficiency holistically and review,

Primary Energy Users ASHRAE Energy Standard 90.1 – How to comply? Cleanroom Design Parameters, Temperature, humidity and air change rates HVAC equipment selections and their impacts Air management concepts and their impacts.

PRIMARY ENERGY USERS The construction cost for the advanced technology facilities are three to four times higher than any commercial building. While the process equipment represents about half of the total project cost, mechanical systems comprise the second biggest share, about 25% of the total project cost. On the other hand, the utility cost to operate the facility to manufacture products can influence the overall manufacturing cost from as much as 20% to 30%. Natural gas or fuel oil is used to generate heating mediums such as steam and hot water. Electricity is consumed to run chillers, cooling towers, and pumps to generate and distribute the cooling medium. Electricity also operates air-handling units to introduce airflow to control space temperature-humidity, filter air for cleanliness and provide a comfortable working environment for occupants.

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Figure 1. Energy end use in Cleanrooms

ENERGY STANDARS

It is difficult to comply with energy standards for process driven facilities since the primary goal is to manufacture products properly and safely. ASHRAE Standard 90.1 “Energy Standard for Buildings Except Low Rise Residential Buildings” mandates performance standards for envelope components, mechanical systems and lighting. This standard was updated in 2013. Building envelope, window area, mechanical equipment efficiency, ductwork and piping insulation, sizing electric feeders, piping, ductwork with lower losses and commissioning requirements can be easily met. However, some of the mandates cannot be complied with;

Level of lighting cannot be compromised due to safety reasons. Outside air economizer is not recommended since it can upset the space pressure control Heat recovery cannot be achieved if corrosive chemical influents exist in the air stream.

Energy usage cannot be easily compared since benchmarking in the industry is not easily available for different type of facilities. Manufacturers keep these kinds of information confidential. United Kingdom’s Energy efficiency program gathered and published some of their own energy use values. This can be seen in Figure 2.

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Figure 2: Summary of energy performance indicators from the UK pharmaceutical Survey (ECG 33) Legend: Delivered Energy; Energy used in the facility Primary Energy; Energy used to generate and deliver to the facility. On the other hand, energy modeling of pressure controlled and air changed based facilities cannot be done by using commercially available software which is developed for standard buildings. One needs to develop and validate home grown spreadsheets to perform energy calculations by using Bin Data. This will not provide consistent calculations from one facility to another and from one designer to another.

CLEANROOM DESIGN PARAMETERS

Critical parameters for Cleanroom validation are: temperature, humidity, and particulate counts depending onto the process areas. Control tolerances are very important since they drive the quality of the control devices and eventually the project cost. ASHRAE comfort conditions are provided for spaces other than cleanrooms. Cleanroom temperature and humidity requirements are driven from the gowning requirements, activity levels, electrostatic build-up and process needs. When deciding the internal design conditions, using lower than necessary values might drive the overall project’s initial and operational cost. If the space temperature and relative humidity levels are selected as low as 16 degrees C +/-1 degree C and 30% Relative Humidity +/-2%, which results in -1.7 degrees C dew point. To be able to achieve this space dew point, either very cold water and glycol mixture at -4 degrees C or desiccant dehumidification using high pressure steam is required. In both cases, the operational cost will be unnecessarily increased due to inherent inefficiency with their system operation. Chiller will operate at lower COP and pumping energy will be increased due to glycol solution and its higher density and viscosity. On the other hand, steam used to regenerate the desiccant wheel will dry the air but will also heat it up so post cooling would be required. Design engineer should evaluate the impact and explain this to the process engineers to determine the optimum space temperature humidity pair. This will still do the work, but will also reduce energy consumption and simplify the mechanical systems. It is also valid for winter operation, as using higher relative humidity will introduce the need for humidification that increases preheating and steam needs. The location of the humidifier is also critical. Placing them in wrong spot without evaluating the psychometric process might result in a non-functioning system and flood air handling units or ductwork, as well as potential damage to cleanrooms and bacteria growth in the air system. Another critical parameter is the particulate count experienced within the space, which defines the cleanliness level of the process areas. ISO 14644-1 defines the particulate levels at different sizes. See Figure 3 below for details.

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It is important to specify what state the cleanroom will be certified in. Cleanrooms can be certified:

As-built: No equipment or people in the room. After construction, before any tool or equipment installation. At Rest: Equipment installed but no process or activity in the space. In Operation: The space is in operation and people are in the room.

Most of the semiconductor cleanrooms are certified “As-built” condition. While in the US, Pharmaceutical cleanrooms are certified “In Operation”, in European Union, they are certified both “At-rest” and “In Operation” states. In pharmaceutical facilities, airflow rates for different types of cleanrooms are recommended by the regulatory bodies. FDA’s Aseptic Manufacturing Guideline recommend 0.45 m/s air velocity for ISO 5 cleanrooms and min 20 ACH for ISO 8 cleanrooms. On the other hand, while the European Union does not set any flow rates, it requires that cleanrooms can recover from “In Operation” condition to “At Rest” condition in 15 minutes. Even though this gives some leeway for designers to decide what air change can be used, most pharmaceutical manufacturers set their air change rates for each cleanliness level. In Semiconductor, the air change rates are more flexible since there is no regulatory mandate for the air changes. The product yield defines the air change rates. See Figure 4 for the air change rates seen in pharmaceutical and semiconductor cleanrooms. Classification Pharmaceutical Semiconductor ISO 4 (Class 10) N/A 0.3 – 0.5 m/s (60 -100 FPM) ISO 5 (Class 100) 0.45 m/s (90 FPM) 0.2 – 0.45 m/s (40 – 90 FPM) ISO 6 (Class 1,000) N/A 0.1 – 0.3 m/s (20 – 60 FPM) ISO 7 (Class 10,000) 40 – 60 ACH 30 – 70 ACH ISO 8 (Class 100,000) 20 – 30 ACH 10 – 20 ACH Figure 4 – Air Change Rates seen in different industries to achieve cleanliness levels Note: Table is compiled from multiple manufacturers given design criteria. ACH is used for non-unidirectional and mixed airflows. Designers should not use these values, but do calculations to decide the appropriate air changes for their cleanrooms. Recovery time, particulate generation, and type of HEPA filtration influence the air change rates. ISPE Aseptic Facility Guideline recommends using the graph shown in Figure 5, to calculate the air change rates depending on the recovery time. The formula given below can also be used to calculate the air change rates.

Crrest = (Crop – Cs) exp (-Nt)+Cs Where; Crrest: Final room concentration Crop: Initial room concentration Cs: Supply air concentration N: Room air change t: Time

Supply air concentration is reduced depending onto the HEPA filter efficiency. Room particulate concentration can be calculated using the following;

Cs = Crop * (1- Eff) * 100 Where, Eff = Filter efficiency such as 99.99% at MPPS (Most penetrating particulate size), let’s assume it is at 0.3 micron.

Room particulate quantity at a given time can be calculated by using the following formula, Crop = Cs+((PGRp * Np) + PGRm) Where, PGRp = Particulate generated by people – See Figure 6 Np= Number of People

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PGRm= Particulate generated by process

Figure 5 – Cleanroom Recovery Performance

Figure 6 – Particulate generation by people following different gowning protocol. Optimizing and reducing air change rates instead of just using some industry standards will reduce both airflow quantities, and, eventually the energy consumption of the fans. There are also studies are performed by Lawrence Livermore Laboratories on reducing the airflow rates depending on the space particulate quantities. This requires active particulate measurement and can be used for laboratory environments. It would be both challenging and risky to track particulate count, actively and modulate fan speed due to time lag between measurements and modulation of the fan. However, changing air flow rate during occupied (In Operation) and un-occupied periods (At Rest) would be more achievable because that would be a schedule driven operation, similar to occupied – unoccupied temperature resets which are commonly used by office building owners.

HVAC EQUIPMENT SELECTIONS and THEIR IMPACTS

HVAC equipment selection and their energy use at part load operation have a great impact on energy consumption of the HVAC equipment. When selecting chillers, designers should consider the part load efficiencies as given in kw/ton (COP) and as their IPLVs. Temperatures used in heating water for reheat and preheat can impact the type of the boiler and their efficiencies. If a condensing boiler at 93% efficiency can be used in lieu of a fire tube boiler at 82% efficiency, this can also realize great energy savings depending on the facility’s geographical location. The type of humidifier can play a great role in energy consumption. In hot, dry climates, using an adiabatic humidifier such as an air washer or compressed air powered spray can provide free cooling as well as humidification during the summer. In cold climates that require humidification only during winter, it makes sense to use steam humidification because it would preheat the air and humidify from the heat coming from the steam injection. One of the largest energy users in cleanrooms are moving devices – fans. Using the right fan for the right application can reduce the electrical consumption of the air delivery process. Type of the fans used in air handling units can be listed as,

Centrifugal fans – Double İnlet Double Width – DWDI. Plenum fans – Single Inlet Single Width (SWSI) Vane Axial Fans

Double Inlet Double Width (DWDI) fans are mostly used in make-up air. Efficiencies can vary between 70% and 82%. These fans use belt drives and generate particulates. Therefore, they are not recommended for air handling units doing the final cleanroom air moving. At the same time, these fans run at high static pressure due to the high level of filtration and number of elements inside the unit such as pre-filter, after filter, preheat coil, humidifier, cooling coil, which might have a desiccant wheel and post cooling depending on the dew point control, and final filtration. Air handling units moving large quantities of airflow can use vane axial fans since they have high efficiency around 80% to 86%. They have direct drive fans that are connected to the motor without any belts. These types of air handling units provide sensible cooling and final cleanroom filtration. The other

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limitation for this type of fans is that they are not AMCA certified for high static pressures. They also generate more noise than other fans. Hence, they require sound attenuation which increases the first cost and operational energy consumption. This would cancel out the gain from the high efficiency gain. The plenum fan can be used up to 8 inch WC (1900 Pa), and is also available in a direct drive arrangement. However, its efficiency is lower than both Vane Axial and DWDI fans; it runs between 60% and 75%. Plenum fans are commonly seen in semiconductor recirculation air handling units and pharmaceutical mixed airflow air handling units that use both return and supply fans. Owners and designers use variable frequency drives (VFD) on pumps and fans. Most of the time they are being used to modulate the flow depending on the demand. However, due to price reduction, they are being used for hard balancing purposes, too. We should be conscience about using VFDs for all applications. If the equipment is operating mostly on part loads, VFD provides great benefit and reduce energy consumption. However, if the motor is running mostly at full speed, due to heat losses inherent with VFDs can increase the energy consumption. Designers should evaluate the use of the VFD s and should not use them to fix the inefficient, oversized system designs. AIR MANAGEMENT CONCEPTS and THEIR IMPACTS Cleanroom airflow management concepts as shown in Figure 7 is commonly seen in pharmaceutical facilities. In this arrangement, the unit employs both return and supply fans, outside air and recirculation air is mixed and processed in one unit. Since the outside air is mixed with the return air the air mixture has to be cooled down to space dew point temperature to wring out the moisture introduced from the outside air. Therefore, the supply air temperature will be very cold and has to be reheated to maintain space temperature. This brings the simultaneous cooling and heating which will introduce the inefficiency of the supply air processing. The psychometric process is shown in Figure 9. Also, the total air has to go through multiple levels of filtration, which will increase the fan horsepower and energy consumption. On the other hand, this system will be using less equipment, and using it brings the benefits of this air management concept’s simplicity and flexibility. The biggest benefit of this system is the temperature recovery time because the supply air will be cooled down and reheated as required.

Figure 7 – Mixed Air Flow Air Management Schematic The other air management concept that is widely used in semiconductor cleanrooms is called a primary secondary air handling unit system, as seen in Figure 8. In this scenario, the outside air and cleanroom air components are separated from each other. Outside airflow, which is a lot less than the recirculation, air is cooled down or humidified to maintain space dew point. Treated outside air is introduced into the cleanroom via return air inside the secondary air handling unit. Since the outside air component is already treated, this unit only provides sensible cooling as required by the space and the reheating can be either eliminated or reduced. The psychometric process is shown in Figure 10. Only the cleanroom recirculation air passes through the HEPA filters which reduces the stages of filtration. The Primary secondary air handling concept is also used in aseptic processing areas since they are similar to semiconductor fabrication facilities when it comes to containment and cleanliness levels. Each air management concept has its uses in different type of cleanrooms. However, designers should be open minded and try to use the primary secondary air concept as much as possible in both semiconductor and pharmaceutical facilities.

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Figure 8 – Primary – Secondary Air Management Schematic

Figure 9 – Mixed Airflow Figure 10 – Primary Secondary Airflow

CONCLUSION

As a statement of principles for sustainable HVAC design in advanced technology buildings, Don’t apply standards blindly Utilize enough air to produce the desired results, don’t focus solely on air change rates. Air flow should be able to reduce if the area is not occupied Filtration levels and types should be considered carefully Space air should be recirculated whenever it is safe and practical Broadened temperature and Relative Humidity limits should be used as much as it is allowable by the products and

process, Carefully evaluate the use of VFDs Using primary and secondary air management concepts will reduce the energy consumption since it breaks down the

dehumidification and sensible cooling processes.

68

References

1. Mark Maguire, Compliance of Pharmaceutical R&D Facilities with Updated ASHRAE Standard, Pharmaceutical Engineering, Nov/Dec 2009, Vol 15, No. 5

2. ISO 14644-1: 2015 Classification of Air Cleanliness by Particle Concentration 3. ISO 14644-4:2001 Design, Construction and Start up 4. Steve Conchise, Mıke Murray, Energy and Pharmaceutical Industry, Pharmaceutical Engineering, Nov/Dec 1995, Vol

15, No.6 5. Dave Goswami, Mark Butler, Energy Savings in Pharmaceutical Facilities, Pharmaceutical Engineering, Nov/Dec 2007,

Vol 15, No.6 6. Norm Goldschmidt, First Steps for Sustainable Bio-Pharma HVAC, ES Magazine, August 009 7. ISPE – Pharmaceutical Engineering Guidelines for New and Renovated Facilities Volume 3 – Sterile Manufacturing 8. ISPE Good Manufacturing Guide, Heating Ventilating and Air Conditioning (HVAC) 9. William R. Acorn, Nejat Babür, Total Energy Management Strategies for Wafer Fabs, IESH Conference, May 29, 1996 10. Nejat Babür, Design Concepts in Air Management Systems, Controlled Environment, September 2008 11. Phil Naughton, HVAC Systems for Semiconductor Cleanrooms Part 1: System Components, ASHRAE Transactions,

SL-90-5-3 12. High Performance Cleanrooms, A Design Guideline Sourcebook, PG&E, January 2011

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[Abstract:0018][Exergy] THERMODYNAMIC ANALYSIS OF GROUND SOURCE HEAT PUMP

*Fatih YILMAZ, ** M.Tolga BALTA, *** Reşat SELBAŞ

*Aksaray University, Vocational School of Technical Sciences, fatiyilmaz7@gmail.com ** Aksaray University, Faculty of Engineering, mustafatolgabalta@hotmail.com

*** Suleyman Demirel University, Faculty of Technology, resatselbas@sdu.edu.tr

ABSTRACT: In this study is described in detail includes a brief description of the Ground Source Heat Pump as heating system, concentrating on thermodynamic analysis. The main objective of this paper is to do thermodynamic analysis on performance of the heat pump system. The system is selected hydrocarbon (HC R600) refrigerant as working fluid. The heating coefficient of performance for the overall system is found to be 3,40 while the corresponding exergy efficiency is 43,49%. The effect of ambient temperature exergy efficiency is investigated for the overall system. Keywords: energy, exergy, HC, heat pump, heat,

1. INTRODUCTION Nowadays, the concerns about energy sources have increase in the world. Also the depletion of fossil resources and the problem of environmental pollution have become major global issues. But fossil fuels is still being used various purposes, although has raised some environmental concerns as global warming, climate change, melting of polar ice sheets, pollution, ozone layer depletion, acid rains and rise in sea levels [1]. Therefore efficient energy utilization is getting very important many applications. In the world buildings are highly energy consumer in all energy consumptions. In this context, a high thermal efficiency heat pump has been proposed as a new heating apparatus [2]. Usually have been used two types of heat pumps; air-source heat pumps and ground source heat pumps (GSHP) [2]. GSHP is provided clean way for heating and cooling buildings applications. Also GSHP system usage passive energy stored in the ground. This system, due to higher coefficient of performance (COP), is suitable for many applications building types [3]. GSHP systems are a promising technology that is able to exploit the ground as a heat source. Once heating and cooling loads are almost balanced, this systems work efficiently. In literature, there are many of studies about to GSHP system. Some studies have applied the exergy concept to GSHPs [4-5]. Sanaye and Niroomand [6] evaluated the ideal design process of a GSHP system based on characterization and operational requirements imposed on the ground heat exchanger. Naili et al., [7] have investigated energy and exergy analysis of horizontal ground heat exchanger for hot climatic condition of northern Tunisia. In their study, the results showed that the energy and exergy efficiencies are found to range between 52 and 18% and 36 and 12%, respectively. A study on modeling and performance assessment of a heat pump system for utilizing low temperature geothermal resources in buildings conducted by Hepbasli and Balta [2]. The study deal with the modeling and performance evaluation of a heat pump system utilizing a low temperature geothermal resource, which is approximated to a geothermal reservoir. Energy and exergy efficiency values on a product/fuel basis were found to range from 73.9% to 73.3% and 63.3% to 51.7% at dead (reference) state temperatures varying from 0 to 25 1C for the heat pump unit and entire system, respectively. Liu et al., [8] studied that investigation on the feasibility and performance of ground source heat pump (GSHP) in three cities in cold climate zone, China. This study deal with the purpose of quantitative investigation on the feasibility and performance GSHP, three cities located in cold climate zone, Qiqihaer, Shenyang and Beijing, were sampled. The numerical modeling techniques for a GSHP: The Bolu case, have performed by Camdali et al., [9]. Numerical modeling of the GSHP system with horizontal ground heat exchangers operating in heating (1.4 kW) mode was carried out for a room, with 9.68 m2 floor area, located in Bolu, Turkey. The GSHP simulation model is studied. Balbay and Esen [10] studied experimental and computational studies by using vertical GSHP system to analyze the melting of the snow/ice occurred in bridge and pavement surfaces in winter. This study make out that thermal properties related to ground structure were important parameters in design of vertical GSHP system. The number of studies performed on exergetic assessment of GSHP systems in the literature [11-14]. In this paper, based on thermodynamic analyses of GSHP system. The coefficient of performance (COP) of the heat pump and exergy efficiency the overall system are computed by using Engineering Equation Solver (EES) software. In this system used to as a working fluid hydrocarbon (HC) R600. The performance of GSHP system will be determined and evaluated under various parameters.

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2. SYSTEM DESCRIPTION The brief description of the process components and operating conditions have been given to evaluate the thermodynamic assessment of the combined GSHP, whose compressor, condenser, expansion valve, evaporator, radiator and ground source. A schematic of the GSHP system is given in Fig.1. Also GSHP system of temperature-entropy (T-s) and pressure-enthalpy (P-h) diagram are given in the Fig.2.

Fig.1. A schematic of the GSHP system

Fig. 2. GSHP system of T-s and P-h diagram As seen in Fig.1 GSHP system can be separated in to three main circuits; (a) radiator heating circuits for heating, (b) heat pump system and (c) ground heat exchanger circuit for heat demand. This GSHP system provides heating for buildings. The system assumed that as a working fluid hydrocarbon (HC) R600. The refrigeration of (R600) thermophysical properties are given Table 1. Because of R600 refrigeration low Ozone Depletion Potential (ODP) and Global Warming Potential (GWP), in this study selected as working fluid.

0,0 0,5 1,0 1,5 2,0 2,5 3,0-100

-50

0

50

100

150

200

s [kJ/kg-K]

T [°

C]

0,2 0,4 0,6 0,8

R600

T[i]T[i]

1

2

3

4

0 100 200 300 400 500 600 700 800100

101

102

103

5x103

h [kJ/kg]

P [k

Pa]

0,2 0,4 0,6 0,8

R600

P[i]P[i]1

23

4

71

Table 1. Thermophysical properties of R600 Properties R600 Critical Pressure (Bar) 3,8 Critical Temperature (oC) 152 Boil Temperature (oC) -0,4 Molecular Weight (kg/kmol) GWP ODP

58,13 3 0

The GSHP system components shown in Fig. 1 can be treated as a control volume. The thermodynamic analysis of the GSHP system was performed based on the following general assumptions; The changes in kinetic and potential energy are negligible. Adiabatic and irreversible compression an isentropic efficiency of 0.8 Negligible pressure and heat losses in the pipe networks or system components All system components operate under steady-state conditions. Ambient (dead state) temperature (T0) 21 oC and ambient pressure (P0) 101,325 kPa The heat loss to the surroundings and pressure reduces in the pipes connecting the components are ignored For the steady-state process, the mass, energy and exergy balance for the each control volume can be expressed as follows; ∑m ∑m (1) where m is the mass flow rate, and the subscript in stands for inlet and out for outlet. The mass balance equation can be expressed in the rate form as with all energy terms as follows[15]; Q W ∑m h ∑m h 0 (2) The general exergy balance can be expressed in the rate form as; Ex Ex Ex (3) Ex Ex Ex , Ex , Ex (4) Using Eq. (4), the rate form of the general exergy balance can also be written as ∑ 1 Q W ∑m ψ ∑m ψ Ex (5) Where Q is the heat transfer rate through the boundary at temperature, Tk at location k, W is the work rate, ψ is the flow (specific) exergy, h is enthalpy, s is entropy, and the subscript zero indicates properties at the restricted dead state of P0 and T0 The specific exergy (flow exergy) of refrigerant (or water) is calculated by; ψ h h T s s (6) The exergy rate is calculated by; Ex mψ (7) Energy (or first law) efficiency of the HP unit (COPHP) and the whole HP system (COPsys) are determined as follows, respectively, COP (8) COP

, , (9)

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The exergy efficiency for whole system calculated as fallows; ε , ,

, ,

(10) 3. RESULTS AND DISCUSSION Temperature, pressure, and related to this properties data for working fluid (R600) and water are given in Table 2. Also thermodynamic states for GSHP system have been illustrate give in Table 2. In this paper, the reference states are assumed 0 oC temperature and atmospheric pressure 101,325 kPa. Table 2. Stream data at an operating and reference environment and thermodynamic assessment results for flows GSHP. Number Worki

ng Fluid

Phase Temperature T (oC)

Pressure P(kPa)

Specific Enthalpy h(kJ/kg)

Specific Entropy s (kJ/kgK)

Exergy rate

(kW) 0 water Dead state 0 101,325 0,06 -0,00015 - 0 R600 Dead state 0 101,325 584,6 2,41 - 1 R600 Super heat

vapor 10 110 600,9 2,458 0,1854

2 R600 Super heat vapor

67,41 600 686,6 2,509 4,133

3 R600 Liquid 40 600 296,6 1,327 0,4474 4 R600 Mixture 1,658 110 296,6 1,351 0,07496 5 Water Liquid 50,4 101,325 211,1 0,7089 6,262 6 Water Liquid 46,4 101,325 194,4 0,6569 5,353 7 Water Liquid 46,91 101,325 169,5 0,6635 5,466 8 Water Liquid 7,8 101,325 32,87 0,1183 0,1607 9 Water Liquid 14,5 101,325 60,92 0,217 0,5453 10 Water Liquid 15,8 101,325 66,36 0,2358 0,6453 Fig 3. displays the variation of condensation r temperature on COP and exergy efficiency cycle. As can be seen Fig 3. the cycle of ground source temperature 15,8 oC, by increasing condensation temperature the system of COP is decreases but exergy efficient has increased. While condenser temperature 40 oC, the system of COP and exergy efficiency are 2,904 and 0,3709, respectively.

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Fig.3 Variation of condensation temperature with COP The COP and exergy efficiency of the GSHP system have been found as 2,904 and 0,37, respectively. For the GSHP system, COP and exergy efficiency of the system for ambient temperature ranges from 0 to 30 oC are seen in Fig.4. As shown in this figurate the exergy efficiency of the GSHP system decreases with increasing ambient temperature. The COP of GSHP system is not function of ambient temperature as seen in the Fig.4

Fig.4. COP and exergy efficiency of GSHP system with various ambient temperatures Fig. 5 demonstrates the variation in the exergy destruction of different components of the GSHP system. According to this figure, the highest amount of exergy destruction rate takes place in the condenser (2,8 kW). The lower amount of destruction rate take place in radiator is calculated.

Fig.5. The exergy destruction values in different components of GSHP system

00,5

11,5

22,5

33,5

0 5 10 15 20 25 30 35T0 (OC)

COPsys ε

0,7654

2,889

0,37240,09726

0,6576 0,773

0,032050

0,51

1,52

2,53

3,5

Exer

gy d

estru

ctio

n ra

te (k

W)

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4. CONCLUSION İn this paper, a GSHP thermodynamic analysis of a GSHP system has been carried out using energy and exergy analyses. According to condenser and ambient temperatures exchange the system of COP and exergy efficiency are investigated. This paper as a refrigerants environmentally friendly fluidsR600 is used. The main concluding remarks have been given as follows;

While the system of condenser temperature at 40(oC) is fixed, ground source temperature 15,8 (oC), the system of COP and exergy efficiency are respectively 3,4 and 43%.

As the system of condenser temperature is increased, COP is decreases and exergy efficiency is increases. As the seen that with ambient temperature is increase, the system of exergy efficiency is decreases. The components involving the highest exergy destructions is the condenser

References [1] Dincer., I, 2000. Renewable energy and sustainable development: a crucial review, Renew. Sustain. Energy Rev. 4, (2), 157-175. [2] Hepbasli, A., Balta, M.T., 2007. A study on modeling and performance assessment of a heat pump system for utilizing low temperature geothermal resources in buildings, Building and Environment, 42, 3747–3756 [3] Öztürk, M., 2014. Energy and exergy analysis of a combined ground source heat pump System, Applied Thermal Engineering, 73, 362-370 [4] Ozgener L., Hepbasli A., 2007. Dincer, I., Exergy analysis of two geothermal district heating systems for building applications, Energy Conversion and Management, 48, 1185–1192. [5] Akpinar, E.K., Hepbasli, A., A comparative study on exergetic assessment of two ground-source (geothermal) heat pump systems for residential applications, Building and Environment, 42, 2004–2013. [6] Sanaye, s., Niroomand, B., 2009. Thermal-economic modeling and optimization of vertical ground-coupled heat pump, Energy Conserv. Manage, 50, 1136-1147. [7] Naili, N., Hazami, M., Kooli, S., Farhat, A., 2015. Energy and exergy analysis of horizontal ground heat exchanger for hot climatic condition of northern Tunisia, Geothermics, 53, 270–280 [8] Liu, Z., Xu, X., Qian, C., Chen, X., Jin. G., 2015. Investigation on the feasibility and performance of ground source heat pump (GSHP) in three cities in cold climate zone, China, Renewable Energy, 84, 89-96 [9] Camdali, Ü., Bulut, M., Sozbi, N., 2015. Numerical modeling of a ground source heat pump: The Bolu case, Renewable Energy, 83, 352-361 [10] Balbay, A., Esen, M., 2013. Temperature distributions in pavement and bridge slabs heated by using vertical ground-source heat pump systems. Acta Sci Technol,35,(4), 677-85. [11] Hepbasli A, Akdemir O. Energy and exergy analysis of a ground source (geothermal) heat pump system. Energy Conversion and Management 2004;45:737–53. [12] Ozgener O, Hepbasli A. Experimental performance analysis of a solar assisted ground-source heat pump greenhouse heating system. Energy and Buildings 2005;37:101–10. [13] Ozgener O, Hepbasli A. Performance analysis of a solar-assisted ground-source heat pump system for greenhouse heating: an experimental study. Building and Environment 2005;40(8):1040–50. [14] Hepbasli A. Thermodynamic analysis of a ground-source heat pump system for district heating. International Journal of Energy Research 2005;7:671–87. [15] Cengel, Y.A. and Boles, M.A. 2008. Thermodynamics: An Engineering Approach, 6th edition, McGraw-Hill, NY.

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[Abstract:0021][Heating, Climatization and Air-conditioning Applications in Buildings] THE COMPARISON OF DIFFERENT AIR-CONDITIONING SYSTEMS IN AN INDUSTRIAL

COMPANY

Nazım KURTULMUŞ1 and İlhami HORUZ2

1Adana Science and Technology University, Faculty of Engineering and Natural Science, Mech. Eng. Dept., 01180, Adana/Turkey

2Gazi University, Engineering Faculty, Mech. Eng. Dept., 06570, Ankara/Turkey Corresponding email: nkurtulmus@adanabtu.edu.tr

SUMMARY Air-conditioning systems use quite a lot of energy. This is because, the air-conditioning systems mainly use Vapor Compression (VC) Systems and compressing vapor consumes really a big amount of energy. However, if there is a waste heat and/or geothermal energy available or solar energy capacity is good enough or natural gas prices are reasonable, Vapor Absorption (VA) Air Conditioners are offering a promising alternative. This research aims to compare these two air-conditioning systems. For this purpose, the office building of an industrial company at Adana/Turkey is chosen and the natural gas driven VA and VC Air Conditioners for this office building were being compared. Firstly, the cooling load was calculated. Then, in order to provide the chilled water to fan-coil systems, the VA and VC Systems are investigated individually and compared. Next, the investment and running costs of the systems are determined, compared, presented in tabular form and discussed. INTRODUCTION The rapid increasing demands for energy directs researchers and engineers to use energy effectively. Air-conditioning systems being used in buildings consumes quite a lot of energy. The mainly used air-conditioning system is Vapor Compression Systems which has high running costs because of compressing vapor consumes really a big amount of energy. This means that it requires the mechanical work to compress vapor. Unlike Vapor Compression Systems, Vapor Absorption Refrigeration (VAR) Systems requires heat to drive the system. But the main disadvantage is that the VAR system has the lower COP value compare to the VCR system. Natural gas fired VAR systems may be an alternative to meet air-conditioning demands. Natural gas fired VAR systems has advantageous as follows [1]. -The natural gas is clean and environmental friendly compared with refrigerants as CFCs. - Elimination the peak electric demands. -The natural gas prices are low. Elsafty and Al-Daini [2] investigated the terms of economical the two systems which are the solar powered single/double effect VAR and VCR systems for air-conditioning the hospital. They used the present worth and the equivalent annual methods for comparison the cost of the air-conditioning systems. According to the present worth method, they found that while the total cost of VCR system is 11% lower than the cost of single effect VAR system, the cost of double effect VAR system is 30% lower than the cost of VCR system. And also the cost of double effect VAR system is 45% lower than the cost of the single-effect VAR system. According to the equivalent annual method, while the total cost of VCR system is 6% lower than the cost of single effect VAR system, the cost of double effect VAR system is 37% lower than the cost of VCR system. And also the cost of double effect VAR system is 45% lower than the cost of the single-effect VAR system. Horuz [3] compared the performance of ammonia-water and water- lithium bromide solutions in VAR systems. The VAR system with water-lithium bromide solutions has higher performance than the VAR system using ammonia-water system. While the VAR system with the water-lithium bromide solution is suitable for air-conditioning applications, the VAR system with the ammonia-water solution is suitable for industrial applications requiring low evaporating temperatures in evaporator. Kurtulmus [4] investigated the industrial application of the VAR system. The researcher used water –lithium bromide solutions as working fluid in the VAR system and made economical comparisons of the air-conditioning system. Because of the corrosion effect of the water- lithium bromide solutions, the material used at the VAR system must be resistible. Also the thickness of the tubes must be chosen properly to prevent collapsing because of the vacuum pressure at the VAR system. The effect of these considerations can be effective on the capital cost of the VAR system. In the light of these investigations, this paper compares the natural gas driven Vapor Absorption and Vapor Compression Air Conditioners for the office building of an industrial company. After the cooling load of the office building was calculated, to provide the chilled water to fan-coil systems, the Vapor Absorption and Vapor Compression Systems are investigated and then investment and running costs of these two systems are determined, compared each other, presented in tabular form and discussed. DESCRIPTIONS VCR System Descriptions VCR system has four basic components: condenser, evaporator, expansion valve and compressor. The refrigeration cycle is shown in Figure [1] and lnP-h diagram is shown in Figure [2].

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Figure 1. Vapor compression refrigeration system At the VCR system, the refrigerant vapor enters the compressor and is compressed up to high pressure. The temperature is also being increased. Then, the refrigerant vapor enters to the condenser where heat releases to the environment and it condenses. Condensing refrigerant passes through the expansion valve where the pressure gets reduced and then it enters the evaporator. Low pressure refrigerant evaporates and rejects heat from the medium then it goes to compressor as saturated vapor [5, 6].

Figure 2. Vapor compression refrigeration system lnP-h diagram VAR System Descriptions VAR system, shown at Figure [3] basically has an evaporator, a condenser, a generator, an absorber and a solution heat exchanger. The VAR system cycle uses a refrigerant- absorbent solution. The duty of absorbent is absorb the primary fluid which is refrigerant. Water-lithium bromide (water-LiBr) solution where water is the refrigerant and water-LiBr is the absorbent and ammonia- water solution where ammonia is the refrigerant and ammonia-water is the absorbent are most commonly used solutions in VAR systems at the market.

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Figure 3. Vapor absorption refrigeration system At the VAR system, the solution is pumped from absorber to the heat exchanger where the solution temperature increases. It passes through the heat exchanger then enters to the generator. Some refrigerant vapor is removed from the solution by heat supply from a heat source. The remaining solution goes back to the heat exchanger to give heat the solution coming from the absorber. Next, it enters to the absorber. The superheated refrigerant vapor goes to the condenser to release heat to the medium. After condensing process, it passes through the expansion valve then enters to the evaporator to reject heat from the medium. So comfort cooling is executed. After evaporating process, it enters to the absorber and is being absorbed by the solution which comes from the heat exchanger [5, 6]. Ln P-h diagram of the VAR system is shown at Figure [4].

Figure 4. Vapor absorption refrigeration system lnP-h diagram AIR-CONDITIONING APPLICATON TO AN INDUSTRIAL COMPANY The aim of this section is to apply air-conditioning systems to produce chilled water that will be sent to the fan-coil system at an industrial company. The office building needs to be air-conditioned and it has 85kW comfort cooling load. The drawing of the office building to be air-conditioned can be seen at Figure [5].

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To compare the VAR system application and VCR system, three different models are chosen; the water cooled VAR system using water-LiBr as working solution, the air cooled VAR system using ammonia-water as working solution and air-cooled VCR system. The obtained chilled water from the chiller units is sent to be the fan-coil system which is laid on the office buildings. Some of technical specifications can be found at Table [1].

Figu

re 5

. Coo

ling

load

of t

he ro

oms

79

Table 1. The technical specifications of the air-conditioning systems VCR system VAR system using water-

LiBr VAR system using

ammonia-water Cooling type

Air-cooled Water cooled Air-cooled

Air-conditioning capacity

108 kW 105 kW 107,58 kW

Electric consumption (including fan-coil system

and air-conditioning system)

~45 kW

~4 kW

~7.6 kW

Natural gas consumption - 103 kW 150 kW As it can be seen from Table [1], the air-conditioning capacity of the three systems is close to each other. The electric consumption of the VCR system which has highest electric consumption because it includes the electric consumptions of the compressor, the condenser fans, the fan coil system and the chilled water circulation pump. The selected VCR system’s COP is about 3. The electric consumption the VAR system using ammonia-water solution includes the electric consumptions of the fan-coil system, the chilled water circulation pump and the fan using for cooling the VAR system’s condenser and absorber. The electric consumptions of the VAR system using water-LiBr solution includes the fan-coil system, the chilled water circulation pump. The VAR system using water-LiBr solution consumes natural gas and electric power lower than that the VAR system using ammonia-water solution. While the COP of the selected VAR system using water-LiBr solution is about 1 and the COP of the selected VAR system using ammonia-water solution is about 0,72. The fan-coil system inlet and outlet temperatures are 7º C and 12 º C respectively. During the cost analysis, the life cycle cost technique which includes all cost factors as first cost, operating and maintenance costs. To apply this technique, equivalent annual method is chosen. At this method all the costs taking place over a period are converted to an equivalent uniform yearly amount. The equations below are being used [2].

1 ∗1 1

, 1

annualoperatingcost, 2

Where EAC means equivalent annual cost and Pc means present cost. Assumptions: i : 10% n : 20 year Electric price : 0.0576 Euro/kW Natural gas price : 0.2671 Euro/m3

Where i means annual interest rate and n is life of the components. Table 2. Cost comparison of the air-conditioning systems

Initial cost (Euro) Equivalent annual cost for initial cost (Euro)

Annual operation cost (Euro)

Total equivalent annual cost (Euro)

VCR system with fan-coil system

36300

4264

3110

7379

VAR system using water-LiBr solution and fan-coil system

89336

10494

3718

14212

VAR system using

ammonia-water solution and fan-

coil system

67014

7872

5536

13408

80

As can be seen from Table [2], the VCR system with fan-coil system has the lowest initial and annual operation costs. While the VAR system using ammonia-water solution and fan-coil system has lower initial cost compared to the VAR system using water-LiBr solution and fan-coil system, the latter one has lower annual operation cost because of its higher COP value. The VCR system with fan-coil system has the lowest total equivalent annual cost. Although it seems to the VAR system using water-LiBr solution and fan-coil system has the highest total equivalent annual cost and both VAR systems with fan-coil systems have close total equivalent annual cost. CONCLUSION In this study, to meet the demands effectively, the investigations of the cost comparison were applied to different air-conditioning applications conducted for an industrial company. To achieve this, firstly the comfort cooling load of office building was calculated and the fan-coil system applied to the building appropriately, then to send the chilled water to the fan-coil system, the appropriate chillers as VCR and VAR systems were chosen. After the comparison of these systems, it is concluded that the VCR system with fan-coil system has the lowest total equivalent annual cost and the VAR system using water-LiBr solution and fan-coil system has the highest total equivalent annual cost. This is due to the fact that since there is no manufacturers of VAR systems in Turkey and the initial costs are quite high. The other reason is that industrial electric power rate is quite low. REFERENCES 1. El-Gohary, M M. 2013. Economical analysis of combined fuel cell generators and absorption chillers. Alexandria

Engineering Journal. Vol. 52, pp 151-158 2. Elsafty, A. and Al-Daini, A J. 2002. Economical comparison between a solar powered vapour absorption air-

conditioning system and a vapour compression system in the Middle East. Renewable Energy. Vol. 25, pp 569-583 3. Horuz, I. 1998. A comparison between ammonia-water and water-lithium bromide solutions in vapor absorption

refrigeration systems. International Communications Heat Mass Transfer. Vol. 25, pp 711-721 4. Kurtulmus, N. 2014. The Application of Absorption Systems to an Industrial Company. MSc Thesis: Gazi University,

Graduate School Of Natural And Applied Sciences 5. Yamankaradeniz, R. Horuz, I. Kaynaklı, O. et al.2009. Soğutma Tekniği ve Isı Pompası Uygulamaları:DORA press

(InTurkish). 6. Çengel, Y.A. Boles, M.A. 2011. Thermodynamics: An Engineering Approach. McGraw Hill Press.

81

[Abstract:0022][Structure Physics] SAP2000 ANALYSIS OF STEEL SUPPORTS USED FOR SUSPENDED MEP SERVICES

Alp Yücerman1, Merve Türkoğlu2 and Anıl Çağatay3 1Ulusyapi Tesisat AS, İstanbul 2Ulusyapi Tesisat AS, İstanbul 3Ulusyapi Tesisat AS, İstanbul alp.yucerman@ulusyapi.com merve.guney@ulusyapi.com anil.cagatay@ulusyapi.com

SUMMARY Steel supports are being used for suspended and floor-mounted equipment and distribution lines in pipe galleries, mechanical plant rooms, roofs, tight corridors, equipment inlets & outlets located in facilities such as skyscrapers, hospitals, industrial facilities etc. The problem is designing these steel supports by only considering the static vertical load, which is a combination of MEP-service weight and the dead load of the support. Actually, there are many different factors like earthquakes, operating pressure, thermal deformation etc. that can cause significant additional loads on the support. This paper aims to provide information about what type of loads effect on the steel supports and which kind of load combinations, calculation methods and design criteria should be used in design phase. Keywords: Steel support, MEP service, support design, non-structural components, earthquake load INTRODUCTION In today’s world, density of MEP-services in buildings reached serious proportions due to the increase in demand and expectations. These services are essential for a building and/or facility to be functional to its full potential. Depending on the design seismic performance level of the building, some degree of seismic restraint of these non-structural systems is required. Sometimes it is just small precautions for collapse prevention, but sometimes it is higher level of protection to keep the facility operational just after the seismic activity. Once the design of seismic force on a mechanical component (piping, equipment etc.) is calculated, the next step is to make the necessary precaution. This can be done either by fixing it to the structure or by restraining it with specifically designed, manufactured and tested seismic restraint assemblies (snubber, seismic isolator, steel cable etc) [1]. In many cases, due to the intensity of the distribution lines and site conditions installing the above mentioned seismic restraint assemblies is being impossible. When facing this kind of problems, first thing that comes into mind is using high-load capacity steel supports. These supports usually fabricated on site by only considering the section properties. Meaning by that, only the steel support shape and size is being determined. In addition to that, these analyses are being made only by considering the static vertical load. Another common mistake is, not properly designing and checking welding strength, anchor bolts, bolt-nut connections which are the weakest points in the load path. LOAD COMBINATIONS Vertical Loads In construction business, steel structural supports are generally designed by using a specialized software. This kind of softwares has many common steel section shapes in their libraries. In addition to that, the designer can create custom material properties and sections if a job requires. These softwares enable us to do analyses in three dimensional space by just setting the material properties, sections, load values, directions and combinations. In Figure 1 below, you can see the static load distribution on a generic steel support. This is a combination of vertical loads at -Z direction which includes pipe weights and a dead load of the support itself. Unfortunately, most of the steel supports are being designed by only considering this vertical load combination. In addition to that, to make matters worse, the dead load of the support held exempt from the calculations in many cases.

82

COMBINATION 1: Dead+Static

Figure 1.Vertical Static Loads Seismic Forces In addition to the vertical loads, seismic forces are one of the additional forces that should be taken into consideration to avoid using additional seismic restraint assemblies. Actually, in many cases, it is the combination of seismic restraint requirement and space constraints of site that gives birth to a requirement of rigid steel supports. Design seismic forces can be calculated thru empirical formulas, and most commonly used ones are from IBC (International Building Code) code. Following lateral and vertical seismic design force formulas can be found in ASCE 7-10 (American Society of Civil Engineers) [2] which is being referred by IBC 2012 [3] (Equation 1 & 2).

,

1 2 , (1)

0,2 , (2) Where;

Fp is the design lateral seismic force, Fpv is the design vertical seismic force ap is the component amplification factor, Rp is the component response modification factor, Ip is the component importance factor, SDS is the design spectral response acceleration at short periods, Wp is the component operating weight, z is the height in structure at point of attachment of component, h is the average roof height of structure.

83

Figure 2. X-Y-Z Axis. The seismic forces are being taken into consideration in three directions while performing an earthquake protection design. Eq X-X and Eq Y-Y are the design seismic lateral forces, Eq Z-Z is the design vertical seismic force which affect in the direction of X,Y and Z axes respectively. They all can be considered in analysis by defining new load combinations as follows: COMBINATION 2: Dead+Static+Eq X-X+Eq Z-Z COMBINATION 3: Dead+Static+Eq Y-Y+Eq Z-Z COMBINATION 4: Dead+Static+Eq X-X+Eq Y-Y+Eq Z-Z

Figure 3.Defining New Load Combinations.

CONNECTIONS & ATTACHMENTS

After running analysis we can display the frame element forces and moments in three dimensions of attachment points. These forces and moments could and should be used to calculate the welding thickness, anchor strength and plate size. Figure 4 below shows an example of structural attachment of a suspended support. Technical details like welding area, steel plate thickness & dimensions, distance between anchors etc. should be specified by design engineer.

84

Figure 4.Example of an Attachment Detail. All the connections should be taken into consideration and separate calculations has to be performed for each connection. As seen in the Figure 4, first step is to attach the steel profile to the plate. Therefore, the welding thickness should be calculated [4]. If the size of the welding around steel section isn’t enough, the connection can be strengthened by using additional reinforcing plates. After this, designer can move on to the selection of proper anchor bolts, which connects the plate to the structure. Plate design and the distance between anchor bolts have influence on tension and shear loads acting on anchors. Sometimes, the analysis results with a loads in very high values ocurring on the connections. In such cases selecting a proper connection can be unreasonable for a given support. Calculations could require using M50 anchor bolts, 1000x1000 mm base plates, 50 mm thick reinforcing plates, meters long weldings etc. Designing this kind of supports are relatively unfeasible or even impossible in many cases. To avoid this situation, the designer can reduce loads by reducing the spacing between supports. Also if the axial loads are too high, than the pipe stress designer should consider changing the design in order to reduce the loads.

85

Local Axes of Steel Sections

Figure 5.Local Axes of Steel Structural Profile. 1: Neutral Axis 2: Weak Axis 3: Strong Axis During the design phase, designer should pay attention to direction of the supporting steel profile according to the weak axis and strong axis which are shown in the figure above. In case of ignoring the direction of supporting steel and choosing the weak axis to take the loads by accident, strength of the supporting steel profile significantly decreases. In some cases, if the loads are low enough, the designer might chose the weak axis to take the loads on purpose. The purpose of doing this, could be making pipe intallation easier by placing them on the longer part of the steel section or achieving a height constraint which is usually determined by the site conditions. Last but not least, during the construction of supporting steel profile on site, site engineers and foremen should follow the actual design. It is a very common mistake to change the direction of the main support frame during manufacturing phase. If the site engineer change the direction of supporting steel profile and make the weak axis take the vertical loads, support might collapse while the acting loads getting closer to the design loads (which are calculated by considering the strong axis) and the consequences could be disastrous.

86

Figure 6.Steel Design Results (Strong Axis).

Figure 7. Design Results (Weak Axis). Figure 6 & 7 above show the difference between strong and weak axis. Both of the generic supports are identical: Same sections, same height and width, same connections etc.. Acting loads on the supports are also identical: same values, same directions, same locations etc.. The only difference is the direction of the main lateral steel frame which is placed in its strong axis against vertical forces in Figure 6 and in its weak axis against vertical forces in Figure 7. When we run the software and test the frame in question, strong axis (Figure 6) is working at 56% of its capacity, while the weak axis (Figure 7) is being forced to 79% of its capacity. We can say that the strong axis results about 1.5 times stronger than the weak axis fort his kind of generic set-up. CONCLUSION While designing high-capacity steel supports, all of the acting loads should be considered to prevent failure and/or collapse of the support which can be disastrous both from the life-safety and cost efficiency perspective. In this study we only covered the seismic forces since it is the least considered or even remembered during the design phase. There may be additional forces caused by wind, thermal shape change, working pressure of the service carried etc.. Sometimes the problem is manufacturing or installing it on site. Because of the limiting site conditions, not every design on the paper is suitable to manufacture. Finding this kind of fact the hard way, meaning after the completion of the design, is clearly unfavourable for time and cost efficiency. To prevent this to happen, site engineer and the designer of the support should be in a constant and healthy communication since the pre-design phase. Last but not least, design perimeters should be followed properly during the manufacturing and installing phases. Forgetting or disregarding small details like installing sections in the right directions, maintaining right welding thickness & length, following anchor installation instructions etc. may result in high carrying capacity loss. . ACKNOWLEDGEMENT This study is supported and sponsored by Ulus Yapi Tesisat Malzemeleri ve Sanayi Tic. AS. REFERENCES 1. Kalafat, E, Yucerman A, Goksu B. 2015. Seismic Restraint For Non-Structural Components. 2. American Society Of Civil Engineers (ASCE). 2010. ASCE 7-10 Minimum Design Loads for Buildings and Other

Structures 3. International Code Council (ICC). 2012. IBC-2012 - International Building Code. 4. Turkish Standards Institute. 1979. TS 3357 - Turkish Standard. Building Code for the Design and execution of

Welded Connections in Steel Structures

87

[Abstract:0023][Economy of Energy and Environment] ANALYSIS OF THE CHIMNEY RELATED CARBON MONOXIDE POISONING

Muammer Akgün

BACADER, Baca İmalatçıları ve Uygulayıcıları Derneği muammerakgun@bacader.org

SUMMARY Every year in Turkey, thousands of people are exposed to flue based carbon monoxide poisoning and hundreds of these people lose their lives. These poisoning cases and deaths are especially because of the stoves, water heaters and the combo boilers that are used. In this article, the data that is collected on this topic since 2011 is analyzed according to months, years, provinces, areas and devices that have caused the problems. Also the data on the chimney based carbon monoxide poisoning and the deaths are evaluated. Moreover, some solution suggestions on these problems are developed. 1- INTRODUCTION Every year, hundreds of people are poisoned due to the chimney related carbon monoxide. There are no data in our country about chimney related carbon monoxide poisoning. In this subject, especially cities with higher poisoning risk rate are getting some protection methods such as informing people about poisoning. There are local studies about this issue however there are no governmental solutions. In this paper, poisoning cases in Turkey are reviewed first generally then regionally after then city based. Also devices which cause poisoning are explained.

Generally, chimney related carbon monoxide poisoning increase in winter time because of the beginning of the heating period. 2- THE CONDITIONS IN TURKEY[1,2] Every year in Turkey, thousands of people are exposed to flue based carbon monoxide poisoning and hundreds of these people lose their lives[1]. In this study, data are processed until the beginning of 2016. In Turkey, chimney related poisoning data are very high when compared to European countries. In our country, chimney related poisoning and death data are shown in Fig.2 and Fig.3

Fig.1- Annual chimney related poisoning and death data between years 2011-2015

919

3316

2398

9861248

216 221 221 174 209

0

500

1000

1500

2000

2500

3000

3500

4000

2011 2012 2013 2014 2015

POISONINGDEATH

88

Fig.2- Monthly chimney related poisoning data between

years 2011-2015.

Fig.3- Monthly chimney related death data between

2011-2015 3- CONDITIONS IN REGIONS

Chimney related poisoning and death data are sorted according to regions. 3.1- MARMARA REGION

Fig.4- Annual chimney related poisoning and death data between years 2011-2015

in Marmara Region. The region with highest rate of carbon monoxide poisoning & death is Marmara. In Marmara region, Bursa is the city that has the highest rate. 3.2- AEGEAN REGION

Fig.5- Annual chimney related poisoning and death data between years 2011-2015 in Aegean Region.

In Aegean region, the carbon monoxide poisoning & death rate are lower compared to other regions.

204 206 111

116 19

0 4 0 0

18

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22 26 22

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POISONINGDEATH

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3.3- BLACK SEA REGION

Fig.6- Annual chimney related poisoning and death data between years 2011-2015 in Black Sea Region. In Black Sea region, the carbon monoxide poisoning & death rates are higher because of geographic conditions.

3.4- MEDITERRANEAN REGION

Fig.7- Annual Chimney related poisoning and death data between years 2011-2015 in Mediterranean Region.

In Mediterranean region, the carbon monoxide poisoning & death rates are lower than other regions.

115

174 19

6

137 15

4

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42

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90

3.5- CENTRAL ANATOLIAN REGION

Fig.8- Annual chimney related poisoning and death data between years 2011-2015 in Central Anatolian Region.

Central Anatolian Region is one of the regions that have the highest rate of the carbon monoxide poisoning & death.

3.6- EAST ANATOLIAN REGION

Fig.9- Annual chimney related poisoning and death data between years 2011-2015 in East Anatolian Region.

East Anatolian Region is one of the regions that have the lowest rate of the carbon monoxide poisoning & death.

431

856

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3.7- SOUTH EASTERN ANATOLIAN REGION

Fig.10- Annual chimney related poisoning and death data between years 2011-2015 in South Eastern Anatolian Region.

South Eastern Anatolian Region has a high risk of carbon monoxide poisoning but the death rates are pretty low.

4- CONDITION IN THE CITIES[1,2]

Fig.11- Annual chimney related poisoning and death data

at 2011 in most effected cities.

Fig.12- Annual chimney related yearly poisoning and death

data at 2012 in most effected cities.

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Bursa

Gaziantep

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Zonguldak

37

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İçel

İzmir

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Manisa

Samsun

Trabzon

Tunceli

92

Fig.13- Annual chimney related yearly poisoning and

death data at 2013 in most effected cities.

Fig.14- Annual chimney related yearly poisoning and death

data at 2014 in most effected cities.

Fig 15- Annual chimney related yearly poisoning and death data

at 2015 in most effected cities.

40

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0 25 50 75 100 125 150 175 200

AdıyamanAmasya Ankara

Bolu Bursa

Denizli Eskişehir

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İzmir Kahramanmaraş

KayseriKırıkkale

KilisKonya

Manisa Samsun

SinopTekirdağ Trabzon

Yozgat

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5- DEVICES THAT CAUSE CHIMNEY BASED POISONING

Fig.16- The data of chimney based poisoning and death according to devices

between years 2011-2015. 6- CONCLUSION AND RECOMMENDATIONS

These poisoning cases and related deaths are due to stoves, combi boiler, water heaters etc. These death cases can be protected with some precautions such as;

- Ventilation of boiler rooms - The utilization of CO detectors. - The cleaning of chimney biannually. - Keeping the minimum height between the chimneys and the roof top at least 40 cm.

The reason why carbon monoxide poisoning risk gets higher in winter then other seasons is the temperature of the weather. The precautions mentioned above are essential. But there are also ways to make the situation even better. The way to prevent carbon monoxide poisoning is to have chimneys according to the regulations. If the chimneys are not according to the regulations, if the chimney can’t draught enough, then the precautions will be insufficient. If the chimney is according to regulations then even if there occurs carbon monoxide due to burning the chimney will remove gas out of the building. The regulation that makes sure that the chimney must be used is the building regulations. The suggestion for decreasing the deaths are written below,

1- The controls for the new constructed buildings should be more careful and all control engineers should know chimney regulations and standards.

2- The chimneys must be produced by firms that have factory production control document with declaration performance. 3- The periodic control regulations must be made immediately.

REFERENCES 1- Akgün M, Carbon monoxide poisoning and Statistical Data, The Meeting for Prevention of carbon monoxide

poisonings in Turkey, Türkiye Halk Sağlığı Kurumu, 10-12 April 2014 Antalya. 2- Akgün M, Statistical analysis of chimney related carbon monoxide poisonings, 12.National Congress of Installation, 8-

11 April 2015, İzmir.

0 500 1000 1500 2000 2500 3000 3500

POISONING

DEATH

POISONING

DEATH

POISONING

DEATH

POISONING

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MBI

BOIL

ERST

OV

E

WA

TER

HEA

TER

SO

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S

88

2

768

168

25

16

38

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84

33117

192

35

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80

20

64

15

2253

163

20

14

61

29

37

8

901

126

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1203

183

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201320122011

20142015

94

[Abstract:0024][Heating, Climatization and Air-conditioning Applications in Buildings] CUI – CORROSION UNDER INSULAION

Fulya Hamidiye

ODE Yalıtım Sanayi ve Ticaret A.Ş. f.hamidiye@ode.com.tr

SUMMARY Corrosion under thermal insulation has become a major problem in the above-ground pipelines and industrial facilities during the recent years. As a result of corrosion, especially at industrial plants, the occurred damages may result with temporary or permanent shut downs or costly repairment and maintenance activities or replacing of the complete equipment, thus it precludes the completion of the intended service life. Corrosion under insulation (CUI) is the corrosion of piping and vessels at industrial plants that occurs beneath insulation as a result of unexpected water penetration. To fight against the corrosion under thermal insulation and unintended consequences all required precautions should be taken from the design stage.This paper will focus on the reasons of corrosion occurrence under insulation and necessary precautions to fight against this problem and its unintended consequences. Keywords: corrosion under insulation, CUI, pipeline, industrial facility, insulation INTRODUCTION Corrosion under insulation (CUI), is one of the major difficulties sighted in industrial facilities, petrochemical plants, power plants and pipelines. Especially at industrial plants, the occurred damages due to corrosion may result with temporary or permanent shut downs or costly repairment and maintenance activities or replacing of the complete equipment, thus it precludes the completion of the intended service life. Although it is difficult to determine the start point of corrosion under insulation at industrial plants, after removal of complete insulation material and finishing the repair work, all over installation of new insulation material is not an economical solution. At the design stage of industrial plants, the first prevention method against CUI (Corrosion Under Insulation) is to apply high quality protective coatings. The second prevention method is to apply jacketing against vapor at industrial plants exposed to atmospheric conditions. The third and the most important technique is to choose and apply the appropriate insulation material with accurate thickness. METHODS Since 1980's oil and gas industry has realized CUI as an important problem and continued research activities to eliminate the negative effects. API 580 (Risk-Based Inspection Recommended Practice) was developed by American Petroleum Institute, to determine the risk-based inspection methods such as radiographic inspection, ultrasonic inspection and thermographic inspection for usage during the service life of petro-chemical plants. Some approaches and methods against CUI at industrial plants:

1. Risk-Based Inspection (RBI) : This method is generally used for examination of tanks, pressure vessels, heat exchangers and pipe-lines at industrial plants. In general, by this method it is possible to determine current and potential risks by periodically examinations. But the disadvantage of this method is that this method is generally applied without taking in consideration the status and examination history of equipment. in some cases, due to disadvantage point, although it is not necessary the equipment can be left out of service in accordance with examination result and this may cause wastage of money and time.

2. CUI detection works usually start at later periods at the life cycle of industrial plant and due to lateness; corrosion may be seriously in an advanced stage.

3. Many of pipelines are usually coated with protective coatings at installation stage. Application of some special types of protective coatings before installation of insulation material over steel equipment take precaution against CUI and corrosion proceed.

95

A. General Definition and Mechanism of CUI In the most general sense, CUI is corrosion based on condensation over the surface or within material due to increasement of water vapor at exterior side of insulated equipment. CUI in usually caused by the penetration of unexpected water and chloride ions to system. Water vapor and high temperature usually trigger the probability occurrence of corrosion. Also the presence of oxygen may trigger the CUI for materials like carbon steel. Oxygen is already present in nature. Also chloride ions are abundant presence at sea water, potable water and process water. The presence of acids also may cause corrosion and accelerate the process of corrosion. At structures built with high initial investment ,for protection against the corrosion and undesired results of it, direct or cappillar penetration of water to the system should be avoided by taking necessary measures at design stage. although penetration of water either directly or indirectly to the system, may damage the insulation material it also may devastate the equipment under insulation. Mechanism of corrosion under insulation results from presence of below mentioned three reasons:

- The presence of oxygen - High temperature - Intensely presence of some dissolved elements like chloride

Under normal conditions, as the temperature increases, the amount of oxygen dissolved in solution decreases by the boiling point is reached resulting in reduced corrosion rates. However, on the surface covered by insulation, a poultice effect is created which holds in the moisture which essentially makes it s closed system. In fact the measured corrosion rates associated with corrosion under insulation follow trends to higher corrosion rates commonly associated with only pressurized systems. Furthermore, in cases where precipitation becomes trapped on the metal surface by insulation, corrosive atmospheric constituents such as chlorides and sulfuric acid can concentration to also accelerate corrosion. In some cases, chlorides are present in the insulation which greatly promotes corrosion of the underlying surface which it becomes laden with moisture. As mentioned in API 570 there are certain temperature ranges for formation of CUI. Especially continuous condensation caused by service temperatures and re-evaporation of moisture in the atmosphere, leads to determine this temperature range between 25 °F to 250°F in carbon alloy steel pipeline systems. Although it is not continuous, in certain periods, service temperature above 250 ° F pose a risk for CUI formation in carbon alloy steel pipe lines. The following table summarizes the risk of CUI at various service temperatures. Tablo 1. The risk of CUI at various service temperatures

Service

Temperature

Risk of CUI

Pipe- line (Painted)

<10 Years

Pipe- line (Painted)

<10 Years

Paintes/ Pipe-line

>10 years

Intermittent

Operation Status

Continuous Operation

Status

Intermittent Or

Continuous Operation

Status

< -30 °C Low-Medium Low Medium

-30°C to 30 °C Medium Low-Medium High

30 °C to 120 °C High High High

>120 °C Medium Low Medium

96

B. Insulation Materials Can be Used Against CUI: Insulation materials may be categorized in to part as the insulation materials used at high service temperatures and low service temperatures. Insulation material types are decided by taking in consideration the service temperature and ambient temperature at design stage. Also by selecting the protective coating a second measure would be taken.

1. Insulation materials to be used at low service temperatures: Plastic materials such as polyurethane and poliisocyanurate having a porous cell structure may be used at structures having low service temperature. However, when these types of insulation materials are used, acidic solution may occur due to penetration of water and this may cause corrosion. also production technique of phenolic foams may cause acidic reaction if the water penetrates to system in any way. in conclusion while using the plastic insulation materials at low service temperate systems, the material should be chosen as closed pore material or the insulation material should be covered with an appropriate material to prevent the penetration of water.

2. Insulation materials to be used at high service temperatures : insulation materials like aerojel, mineral wool, calcium silicate and pearlite can be used at high service temperature systems. Although mentioned insulation materials are appropriate for high service temperature systems, all necessary preventive actions should be taken against the water penetration due to porous structure of material. although it is possible to take the necessary precautions at production stage (lamination) of material, it is also possible to prevent the water penetration at site installation stage by sheet metal coating over the insulation material.

C. Determination Methods of CUI: For identification of CUI there are some non-destructive testing methods that gives reliable results. CUI may be determined by visual inspection after removal of insulation material. But for a pipe line with 8-10 km length the complete removal of insulation material is not an economical solution. Also it is possible to use removable insulation materials at installation stage. But at partially removal stage, water may penetrate to the system and this may cause corrosion. It is possible to determine CUI by using X-Ray inspection method without removing the insulation material installed over the system. The important issue in using this method is that the X-Ray source and detector should be appropriate to check all the surfaces of pipe. Since some insulation materials behave as barrier against X-Ray beams it is better to use this method for low diameter pipes. Another method for determination of CUI is AE (Acoustical Emission Method) method. In this method, for placing the AE sensor a very small part of insulation should be remover. This method determines the amount of water vapor between the circular surface of steel pipe and insulation material. By measuring the water vapor amount it is possible to evaluate the existing CUI and future probability of occurrence of CUI. CONCLUSION Considering the above-mentioned explanations, especially at recent years CUI has become an important problem at structures with high initial investment cost. By technological developments although it is possible to use appropriate insulation materials every passing day against and there are new inspection methods, CUI is still a big problem that leads to the loss of millions of dollars. It is possible to fight against CUI and negative impacts of corrosion by determination of appropriate insulation material and protective coating at the design stage. In addition, by using the periodic inspection methods and maintenance activities in accordance with standards, it is possible to prevent from corrosion and unexpected results due to CUI. REFERENCES: [1]. T.H.karakoç, O. Turan, E.Binyıldız, E.Yıldırım, Isı Yalıtımı Kitabı, Rota Yayın Yapım Tanıtım Tic. Ltd. Şti., İstanbul, 2010. [2]. J.F. Delahunt, Corrosion Under Thermal Insulation and Fireproofing an Overview, NACE makale no. 03022, Houston, 2003. [3]. T. Hanratty, Corrosion under insulation is a hidden problem, Hydrocarbon Processing, Mart 2013, s 51-52. [4]. M. Funahashi, Solution to CUI with Three Layered Control and Warning Systems, NACE makale no. 4079, 2014. [5]. “Standard Practice for Inner and Outer Diameters of Rigid Thermal Insulation for Nominal Sizes of Pipe andTubing (NPS System)”, ASTM C585-90, 2004. [6]. ODEcalc Calculation Programme (Sürüm 1.0) , ODE Yalıtım Sanayi ve Ticaret A.Ş., İstanbul, 2010.

97

[Abstract:0026][Indoor Air Quality and Comfort Conditions] EXPERIMENTAL INVESTIGATION OF VERTICAL TEMPERATURE DISTRIBUTION

FOR A HYBRID HEATING SYSTEM

M. Fatih Evren1, Abuzer K. Özsunar2, Birol Kılkış3 1Hezarfen Energy Co., Ankara

2Gazi University, Ankara 3Baskent University, Ankara

Corresponding email: fatih.evren@hezar.net SUMMARY In this study, vertical temperature distribution, of a hybrid heating system are investigated experimentally in order to get how combined radiant and convective heating system to effects thermal comfort. Controlled experiments were carried out in a special test chamber which is established according to ANSI/ASHRAE Standard 138. Each interior surface temperature and dry-bulb air temperature of the chamber can be controlled independently in the test chamber. Two different electric fan heaters, which have equal heating capacities but different fan powers, were settled in the test chamber and these heaters were respectively hybridized with floor heating. In the experiments; fan heater and floor heating were operated with different heating capacities simultaneously and hereby radiant to convective split ratio was adjusted indirectly. Vertical temperature distribution was investigated and criticized according to ANSI/ASHRAE Standard 55 for stand-alone floor heating, stand-alone convective heating and hybrid heating. INTRODUCTION Modern human spend their time generally indoor environment and HVAC loads are almost half of the total energy consumptions of buildings. Therefore, maintaining the indoor thermal comfort properly with using low energy and low exergy systems is become a real necessity [1] [2]. Supplying sensible and latent HVAC loads from different and most relevant systems is enable to maximize thermal comfort and minimize both power consumption and exergy destruction [3] [4]. Thus, next generation HVAC systems will be more complex in order to maximize thermal comfort while keep source using minimum. Hybrid HVAC systems are defined as combination of radiant panels systems with a convective system [4]. There have been some studies about hybrid HVAC systems that are generally focused on power consumption, low exergy and dehumidification capacities of the systems. Kılkış et al. designed a hybrid system for The Ankara Museum of Ethnography (Turkey) in 1995. They investigated the optimum radiant to convective heat transfer split for hybrid HVAC systems analytically [3] [5]. Coupling of radiant/heating and cooling panels, convective ventilation and dehumidification systems were investigated within the scope of ASHRAE Research Project-1140 [6]. Kılkış developed a composite radiant wall panel (CRWP) that can be heating/cooling both radiant and forced convection simultaneously [2] [7]. In 2013, BubbleZERO laboratory which is included combination of radiant and convective systems was established by researchers from ETH-Zürich [8] [9]. None of the above-mentioned studies not investigated vertical temperature distribution of the combined systems. In this study we focused on the vertical temperature distribution of the hybrid heating system. On the other hand, there have been several studies about vertical temperature distribution of HVAC systems. METHODS Theory Operative temperature, To, is the fundamental metric of thermal comfort temperature, thus we have investigated vertical air temperature difference under different operative temperatures. Operative temperature can be calculated as the average of mean radiant temperature and dry-bulb air temperature. The equation is shown in Equation (1) [10] [11].

2 (1)

Here, dry-bulb air temperature, Ta, can be measured by a proper sensor. Mean radiant temperature can be calculated with several ways. In this study, mean radiant temperature calculated with Equation (2). Here, Tgl is black globe temperature, e is the the black globe surface emittance (0.95), D is diameter of the black globe (0.15m) and Va is air velocity [12] [13].

98

2731,10 10 ,

,

/

273 (3)

Experimental Setup In this study, experiments were conducted in a special test chamber which were established according to the ANSI/ASHRAE Standard 138 - Method of Testing for Rated Ceiling Panels for Sensible Heating and Cooling with some minor differences. These differences effects only standard panel performance rating. Test chamber walls enumeration and schematics of the integrated hydronic heating/cooling pipes of the surfaces are given in Figure (1) [13] [14] [15].

Figure 1. Test chamber walls enumeration All interior surfaces of the test chamber were coated with aluminum sheets that have 0.9 thermal emittance. Thermal resistance of insulation for all walls, except door, is 2.7 m2K/W, thermal resistance of door insulation is 2 m2K/W. Each wall, floor and ceiling designed as a zone to be able to control their interior surface temperatures independently. Working fluid temperature of each zone are controlled with 3 on/off automatic control valves with 10 s positioning time. Chilled water demand supplied by an 8 kW air source heat pump, chilled water stored at, 200 liter, cold water tank. Hot water demand supplied by 3 kW electric boiler which have also 200 liter water capacity. A view of experimental facility is shown in Figure (2).

Figure 2. Experimental facility 80 calibrated K-Type thermo-couple were attached to the interior surfaces of the chamber in order to measure temperature distribution and average surface temperature of each surface. Moreover, in order to obtain vertical dry-bulb air temperature distribution, 11 calibrated K-Type thermo-couple were located to vertical centerline of the chamber. Levels of these 11 sensors are given in Table 1. In the table marked levels show required vertical temperature measurement levels according to ASHRAE Standard 55 [11].

99

Table 1.Levels of vertical air temperature sensors. Thermo-Couple Nu. Height [cm]

1 10* Ankle (seated and standing 2 35 3 60* waist (seated) 4 85 5 110* head (seated), waist (standing) 6 135 7 160 8 170* head (standing) 9 185 10 210 11 235 A black globe sensor and a hot wire anemometer were located in the center of volume of the test chamber. Two different type portable electric fan heaters used for forced convection heating in the test chamber. Type-1 heater has 800 W heating capacity and 110 W fan power; type-2 heater has also 800 W heating capacity and 10W fan power. These sensors and vertical air temperature sensors can be seen in Figure (3-a) and schematics of the sensor and fan positions can be seen in Figure (3-b).

a) b) Figure 3. Sensor positions. a) Interior view of the test chamber, b) Schema of the test chamber Experiments In this study, stand-alone radiant floor heating, stand-alone electric fan heating experiments were conducted. Then radiant floor heating system and electric fan heater were operated together with different capacities in order to obtain different radiant/convective split for hybrid heating system. In order to analyze effects of fan capacity, experiments were repeated with second type of electric fan heater. Experiments which are hybridized first type fan heater and radiant floor heating is named as Type-1 hybrid heating system, combination of second type heater and floor heating is named as Type-2 hybrid heating system. Table 2 shows control system set points for each surface. Table 2.Set-points for experiments.

Set-Points of Interior Surface Temperatures [°C] Radiant Hybrid Convective

1 2 3 4 C F Ta 1 2 3 4 C F Ta 1 2 3 4 C F Ta 22 22 22 19 22 25 - 20 20 20 17 20 23 22 21 21 21 20 21 21 24 24 24 24 21 24 27 - 22 22 22 19 22 25 24 23 23 23 22 23 23 26 26 26 26 23 26 29 - 24 24 24 21 24 27 26 25 25 25 24 25 25 28

100

All data is the average of the measurements for 30 minutes period that were taken after system reached to the steady state conditions. EXPERIMENTAL RESULTS Vertical temperature distribution on different operative temperatures for stand-alone radiant heating system is shown in Fig. 4. As shown in the figure, minimum air temperature is measured at the height of 35 cm.

Figure 4. Vertical Temperature Distribution for Stand-Alone Radiant Heating System. Vertical temperature distribution on different operative temperatures for stand-alone convective heating system type-1, which has 110 W fan capacity, and type-2 which has 10 W fan capacity are shown Fig. 5. As seen in the Figure 5-a, maximum temperature is measured at the height of 35 cm. As seen in the Figure 5-b, for second type heater, over from the 100 cm air temperature is approximately constant and reaches to the maximum value. As can be seen, air temperature distribution trend-lines of the stand-alone convective system type-1 and type-2 is dramatically different from each other. It is important to note that, stand-alone convective heater is positioned on the ground and the sensor which is positioned at 35 cm is directly exposed to the air draft (draught).

0

50

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200

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15,0 20,0 25,0 30,0

Heig

ht [c

m]

Air Temperature [°C]

Vertical Temperature DistributionRadiant

21.9

23.9

25.9

To

101

a) b)

Figure 5. Vertical temperature distribution for stand-alone convective heating systems, a)Type-1 system, b) Type-2 system. Vertical temperature distribution on different operative temperatures for hybrid heating system type-1 and type-2 is shown in Figure 6. It is clear that fan capacity has dominant effect on vertical temperature distribution.

Figure 6. Vertical Temperature Distribution For Hybrid Heating Systems (Type-1 and Type-2) Figure 7. Shows vertical air temperature distribution comparison for type-1 and type-2 systems. In type-1 hybrid system, convective system has dominant effect on vertical temperature distribution because of high fan capacity and thus high blowing velocity. On the other hands, for type-2 hybrid system, dominant sub-system depends on altitude. Convective system has dominant effect to vertical air velocity distribution from floor surface to approximately 50 cm height. Above from 50 cm,

0

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]

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m]

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Vertical Temperature DistributionHybrid - Type 2

22.124.625.9

To

102

convective system has more dominant effect on vertical temperature distribution. It is important to note that, although convective heater was positioned on the floor, it could not effect dominantly because of the low blowing velocity.

Figure 7. Comparison of All Systems DISCUSSION In this study, vertical temperature distribution of the hybrid heating system was investigated experimentally to point out the effects of hybridization of radiant and convective systems and fan capacities of convective systems to vertical temperature distribution. According to the experimental results, for stand-alone radiant heating system vertical air temperature gradient is very low and the distribution is acceptable according to ANSI/ASHRAE 55 Standard. On the other hand, for stand-alone convective system, even if fan capacity is very low, vertical temperature distribution gradient increased. For hybrid system, if fan capacity of convective system will increase, convective system becomes more and more dominant on vertical air distribution. If low capacity fan used in the hybrid system, domination of convective system on the vertical air temperature distribution will significantly decrease. Furthermore, vertical temperature gradient became comply with the limits of ANSI/ASHRAE Standard 55. In further studies, distributed fan systems with lower capacities or distributed nozzles can be used in order to obtain more smooth vertical air distribution and indoor air quality homogeneity. It is clear that, design of hybrid systems requires multi-objective optimization that includes both energy consumption and exergy destruction analyses and detailed thermal comfort and indoor air quality analyses. Thus, it is clear that, further studies require multi-objective, iterative methods that include both simulative and experimental studies. This study aimed to establish a strong background for creation of initial model of iterative simulative studies. ACKNOWLEDGEMENT The experimental setup, which was used in this study, was established by the financial support of The Scientific and Technological Research Council of Turkey (TUBITAK) as a part of project number TEYDEB 2120177. The authors thanks to Mr Gültekin Şahin (Manager of Gentem Engineering Co.) for his valuable contributions to the experimental facility.

0

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Vertical Temperature DistributionComparison for Type 1

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Hybrid

0

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103

REFERENCES

[1] B. Kılkış, "A Dynamic Operative Temperature Sensor for Low-Exergy High Performance Buildings," in ASHRAE Transactions, Atlanta, 2009.

[2] B. I. Kılkış, "From Floor Heating to Hybrid HVAC Panel - A Trail of Exergy-Efficient Innovations," ASHRAE Transactions, pp. 343-349, 2006b.

[3] B. I. Kılkış, S. R. Suntur and M. Sapçı, "Hybrid HVAC Systems," ASHRAE Journal, 1995. [4] ASHRAE, "Panel Heating and Cooling," in ASHRAE HVAC Systems & Equipments, ASHRAE, 2008. [5] I. B. Kilkis, "Utilization of wind energy in space heating and cooling with hybrid HVAC systems and heat pumps," Energy and

Buildings, vol. 30, pp. 147-153, 1999. [6] D. G. Scheatzle, "ASHRAE Research Project Report (RP 1140) - Establishing a Baseline Data Set for the Evaluation of Hybrid

(Radiant/Convective) HVAC Systems," ASHRAE, Atlanta, 2003. [7] B. I. Kılkış, "Cost optimization of a hybrid HVAC system with composite radiant wall panels," Applied Thermal Engineering, pp. 10-

17, 2006a. [8] M. Bruelisauer, K. W. Chen, R. Iyengar, H. Leibundgut, C. Li, M. Li, M. Mast, F. Meggers, C. Miller, D. Rossi, E. M. Saber, A.

Schlueter and K. W. Tham, "BubbleZERO—Design, Construction and Operation of a Transportable Research Laboratory for Low Exergy Building System Evaluation in the Tropics," Energies, vol. 6, pp. 4551-4571, 2013.

[9] F. Meggers, J. Pantelic, L. Baldini, E. M. Saber and M. K. Kim, "Evaluating and adapting low exergy systems with decentralized ventilation for tropical climates," Energy and Buildings, vol. 67, pp. 559-567, 2013.

[10] ANSI/ASHRAE - Standard 138, "Standard 138 - Method of Testing for Rating Ceiling Panels for Sensible Heating and Cooling," ANSI/ASHRAE, Atlanta, 2013.

[11] ANSI/ASHRAE - Standard 55, "Standard 55-Thermal Environmental Conditions for Human Occupancy," ANSI/ASHRAE, Atlanta, 2004.

[12] ASHRAE, "Measurement and Instruments," in ASHRAE HAndbook of Fundamentals, ASHRAE, 2009. [13] M. F. Evren, A. K. Özsunar and B. Kılkış, "A Controlled HVAC Test Chamber Design and Calibration Through Different Operative

Temperature Measurements Techniques," in Turkish National HVAC&R Congress (TESKON) 2015, İzmir, Turkey, 2015. [14] M. F. Evren and B. Kılkış, "Designing, Manufacturing and Calibrating a Prototype of Room Type Operative Temperature Sensor with

an Anthropomorphic Micro-Mannequin," in Turkish National HVAC&R Congress (TESKON) 2015, İzmir, 2015. [15] M. F. Evren, Experimental Investigation of Optimum Radiant to Convective Split Ratio for Hybrid HVAC Systems, Ankara: M.Sc.

Thesis, Gazi University, 2015.

104

[Abstract:0028][Exergy] RESIDENTIAL ENERGY EFFICIENCY ANALYSIS FOR END USE OF NATURAL GAS

DISTRIBUTION MARKET

Nezih Enes Evren Energy Expert, Energy Market Regulatory Authority (EMRA)

nevren@epdk.org.tr

SUMMARY It is considered that the regulations related the efficiency improvement alternatives for building heating systems as combined heat and power technologies have some effects on the electricity and natural gas markets of Turkey where 59% of total amount of natural gas is used in electricity generation and this rate is 21% for end use of residential heating. Since 48% of total electricity generation is based on natural gas, combined heat and power technologies for end use and the possible impacts on the markets which are regulated under authority of Energy Market Regularity Authority (EMRA) should be projected. The main objective of this study is to develop a methodology by using the actual data of residential natural gas consumption in order to compare the efficiency levels of the individual hating (IH), central hating (CH) and district hating (DH) systems commonly used for residential buildings in distribution market. Such comparison is considered crucial to support better policy-making. Results of the comparative analysis are applied to citywide housing stock data so as to find out the potential emission reductions in case widespread use of CH and DH systems with cogeneration are achieved in cities of Turkey. INTRODUCTION The purpose of this research is to develop a methodology to evaluate energy efficiency performance of three different residential heating systems by using the actual data of residential natural gas consumption. Such a methodology is assumed to be an effective support tool for policymaking to mitigate energy use and carbon emissions in cities [1]. This study is focused on three different residential heating systems commonly used in distribution market that Individual Heating System (IH), Central Heating System (CH) and District Heating System (DH). The local natural gas distribution activity began in 1988 in Turkey. Due to the natural gas distribution license tenders made by Energy Market Regulatory Authority (EMRA) since 2003, the coverage of the distribution network has expanded to 76 cities covered by 69 natural gas distribution regions. The 95% of subscribers of companies with license located within the currently available natural gas distribution regions are residential. Therefore, it is clear that residential natural gas consumption efficiency is of the utmost importance for the Natural Gas Market of Turkey [2]. Nearly 80% of the approximately 9 million residential subscribers in natural gas distribution sector are using individual combi boiler heating systems in Turkey. This rate can be seen more dramatically for Bursa city, which is the case study city of this research. According to the data provided by the Bursa Natural Gas Distribution Company (BURSAGAZ), annual total consumptions of residential consumers using individual heating systems, central heating and district heating system are separately considered for their natural gas related CO2 emissions. In this study, based on the annual total residential consumption data of the company, some citywide outcomes are projected for Bursa city. The citywide outcomes will indicate energy saving and greenhouse gas (GHG) mitigation potential of likely policies that regulate the transition from individual heating systems to central and district ones especially with the cogeneration applications based on combined heat and power technologies. Annual residential natural gas consumption data used with the consumption levels and related CO2 emissions calculated for each different heating system. The total citywide residential areas heated by IH, CH and DH systems are found by using the subscription data of the natural gas distribution company in Bursa with the assumption of average number of heated houses for one gas meter of CH and DH systems. Therefore citywide residential natural gas consumption level and related CO2 emissions of Bursa are calculated as the existing conditions of baseline scenario. Furthermore, some other scenarios are projected to be able to understand the citywide effects of alternative residential heating systems and cogeneration. METHODS According to the tariff regulation of EMRA for the natural gas distribution companies, there are natural gas consumption levels are determined by annual consumption rates. These annual residential consumptions data of BURSAGAZ are considered as IH

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for less than 10,000 Sm3/year, CH for between 10,001 – 100,000 Sm3/year and DH (i.e. DH1) for 100,001 – 1,000,000 Sm3/year. There is one residential meter data has bigger than 1.000.001 Sm3/year consumption (DH2). This meter is used for one of the oldest DH system of Bursa City and heats two different neighborhoods have 2873 houses. According to the field research the average number of houses for one gas meter are assumed for CH and DH1 group of consumers as respectively, 40 and 600 houses per one meter. Climatic differences between the different parts of Bursa city are neglected. To be able to understand the citywide cumulative condition, micro-conditions of the buildings are ignored; average values are taken into account. The existing condition are listed and the some scenarios are projected such as what would be the possible effect of half of the existing IH systems changed in to CH systems, what would be total residential natural gas consumption of Bursa city if the all houses connected in to a DH system or what would be the total residential CO2 emission mitigation if all the houses heated by DH with cogeneration system. Marginal cases are also considered for projections to be able to see the marginal mitigation effects of alternative heating systems. Results of the projection scenarios have been compared with in both total amounts and the percentages for mitigations with respect to the alternatives. The citywide outcomes will indicate energy saving and GHG mitigation potential of likely policies that regulate the transition from individual heating systems to central and district ones. CALCULATIONS AND PROJECTIONS According to the 2014 data of BURSAGAZ, residential natural gas consumption levels and related numbers of gas meter on the field are determined. It is estimated that 40 and 600 houses per one meter for CH and DH1 group of consumers as respectively. Therefore number of houses can be calculated for each heating system. Related CO2 emissions are calculated according to the Emission Factors for Greenhouse Gas Inventories of Intergovernmental Panel on Climate Change (IPCC) and Annunciation for Accurate Billing Principles on Determination of Amount of the Natural Gas Consumption of EMRA (Table 1). According to 2014 values noticed in Emission Factors for Greenhouse Gas Inventories of IPCC, emission factor for natural gas is recorded as 56,1 kg CO2 equivalent for 1 Giga-Joule generated energy (kgCO2/GJ). In addition to IPCC value, according to the unit conversion regulation of Annunciation for Accurate Billing Principles on Determination of Amount of the Natural Gas Consumption of EMRA (Article 4), total emitted CO2 amount can be calculated as 2,1493 kgCO2 during the combustion of 1 Sm3 of natural gas with 100% combustion efficiency [3].

Table 1. Existing Conditions as a Baseline Scenario for Bursa City

Residential NG Consumption

(Sm3) Number of Gas Meters

Average Number of

Heated House per One Meter

Total Number of

Heated House

Average NG Consumption

per House (Sm3)

Residential CO2 Emission of Bursa City

(kgCO2)

IH 416.494.841 594.754 1 594.754 700,28 895.172.361,76 CH 41.485.199 1.616 40 64.640 641,79 89.164.138,21 DH1 7.901.278 32 600 19.200 411,52 16.982.216,81 DH2 2.216.951 1 2.873,00 2.873 771,65 4.764.892,78

TOTAL 468.098.269,00 596.403,00 3.514,00 681.467,00 686,90 1.006.083.609,56 According to the baseline scenario some other scenarios projected such as a first projected scenario it is assumed that 50% of the existing IH systems would be converted into CH systems. The second scenario has an assumption as all of the IH converted into CH. The third projected scenario shows the results for what if all of the IH converted into CH and 50% of the CH converted into DH1. For the fourth projection the assumptions is all of the IH and CH users converted into DH1. For the fifth projection DH2 converted into DH1 to be able to understand the differences between old system district heating and new newer ones with proper insulation and efficiency precautions. For the sixth and the last scenario is projected to be able to determine the cogeneration effect on the DH1 system (Table 2).

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Table 2. Projected Scenarios, Calculations and Potential Mitigations

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Figure 1. Basic System Schema for Cogeneration Process For the last projected scenario the basic system of the cogeneration is considered as shown in Figure 1. Here, the amount of electricity production of the cogeneration is notified by E. The amount of useful heat production of cogeneration is denoted by H. For this purpose, according to the Directive 2004/8/EC on the Promotion of Cogeneration Based on Useful Heat Demand in the Internal Energy Market and Amending Directive 92/42/EEC (2004/8/EC), H is calculated as total heat production minus any heat produced in separate boilers or by live steam extraction from the steam generator before the turbine [4]. E H (1) Here C, shown in Equation 1, is the power-to-heat ratio of cogeneration, it can be selected coefficient seen on Table 3. In this study, C is selected as 0.75 according to the Directive 2004/8/EC [5].

Table 3. Default Power to Heat Ratio According to the Type of Electric Generation Unit Type of the unit Default power to heat ratio, C

Combined cycle gas turbine with heat recovery 0.95 Steam back pressure turbine 0.45

Steam condensing extraction turbine 0.45 Gas turbine with heat recovery 0.55

Internal combustion engine 0.75

Figure 2. Power and Heat Supply Schematics for Projected Sceranios 1-5

Figure 3. Power and heat supply schematics for Projected Scenario 6

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Figure 2 and Figure 3 shows power and heat supply schematics for Projected Scenarios 1-5 and Projected Scenario 6 respectively. Total natural gas consumption amounts of first five scenarios can be calculated by using Equation 2. If electricity demand would be supplied by the cogeneration unit, NGE can be considered as zero [6]. This is because, 59% of total amount of natural gas is used in electricity generation in Turkey and this rate is 21% for end use of residential heating. %48 of total electricity generation is based on natural gas on the other hand [2]. Thus, in new scenario, NG’total supply both electricity load and heating load. Natural gas consumption for heating demand calculated with Equation 3.

….. (for Scenarios 1-5) (2)

′ ′ ′ ….. (for Scenario 6) (3)

Natural gas consumption for heating is calculated below with using Eqn. 3:

′ ′ ′

′H E

Here E can be written as type of H by using Eqn. 1:

′H H C 1 C

Here H is the total NG consumption of the Projection 6, C is 0.75 from Table 1, ηH is 0.55 and ηE is 0.52 [5].

′ H1

0.550.750.52

′ 0.38H

Since the total residential natural gas consumption of Bursa is calculated as 280.440.636,19 Sm3 for Projected Scenario 5 there is a new case can be considered with the cogeneration system as Projected Scenario 6. In the light of the given explanation above, for the last projection, total residential natural gas consumption could be calculated as %38 of the total residential natural gas consumption of Bursa for 2014 is 106.567.441, 75 Sm3. RESULTS AND DISCUSSIONS According to the calculations and comparison study the most efficient system used for residential heating is DH that is very uncommonly seen in Bursa. Although the DH2 is an old heating system and considerably less efficient in terms of the first law of thermodynamics, the differences between DH1 and DH2 results shows that if the old system could be restored as much as the average of the DH1, natural gas saving would be around 1.034.639,97 Sm3 per year. According to the basic assumptions considered in the method the second efficient system is CH as it is expected and the less efficient system is IH which is also have the largest sharing among the heating systems in Bursa. It is calculated that if the DH system would be used for all around the Bursa, just for residential natural gas consumption could be 280.440.636,19 Sm3 and related CO2 emissions became 602.751.059,36 kgCO2 which is 60% of the baseline scenario of existing conditions. Furthermore, the projections indicate that the difference between DH1 and DH2 is limited. When this finding is considered with the Rational Exergy Management Model the importance of the cogeneration becomes apparent especially for DH systems as it shown in Projected Scenario 6. Since total amount of natural gas consumed by the DH system is remarkably bigger than other systems and the settling organization of DH system needs a separated center for boiling equipment the cogeneration systems can be considered as an improving argument for DH system implementations. Calculations shows that DH1 with the cogeneration supplies 38% savings for both residential natural gas consumption and related CO2 emissions with respect to all in DH1 case of Projected Scenario 5. This rate is also around 23% of the existing conditions of baseline scenario. On the other hand, to be able to understand the possible effects of micro-cogeneration and cogeneration implementations on the residential heating systems, some new projections can be investigated with similar data considering the perspective of Rational Exergy Management Model. Considering both consumptions and emissions of samples and citywide projections a better policy making argument can be calculated by using the methodology executed for this study.

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Figure 4. Results of the Comparison Study

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ACKNOWLEDGEMENT The author would like to express my appreciation to Prof. Dr. Birol Kılkış, Dr. Şiir Kılkış, Assoc. Prof. Dr. Osman Balaban, BURSAGAZ employees and Mustafa Fatih Evren for their contribution and supports for the evolution of this study. REFERENCES 1. Evren, N E. 2015, A Comparative Analysis of Efficiency and Carbon Dioxide Mitigation Potential of Residential Heating

Systems in Bursa (Turkey), M.Sc. Thesis, METU, Ankara 2. Evren, N E. 2014, Enerji Verimliliği ve Ekserji Kuramı, Enerji Verimliliği Açısından Doğal Gaz ve Konutlarda Enerji

Verimliliği, Expertise Thesis, EMRA, Ankara 3. Evren, N E. 2015, A Comparative Analysis of Efficiency and Carbon Dioxide Mitigation Potential of Residential Heating

Systems in Bursa (Turkey), M.Sc. Thesis, METU, Ankara 4. EU. 2004, Directive 2004/8/EC, on the Promotion of Cogeneration Based on Useful Heat Demand in the Internal Energy

Market and Amending Directive 92/42/EEC. EU Official Journal, L52/50, Vol. 47, pp. 50-60 5. Kılkış, B. 2007-b, Analysis of Cogeneration Systems and Their Environmental Benefits, HVAC, Refrigeration, Fire

Fighting and Sanitary Journal - Fundamentals of HVAC Design & Application Appendix, Vol. 26, TTMD: Ankara, 6. Kılkış, Ş. 2007, A Rational Exergy Management Model for Curbing CO2 Emissions, ASHRAE Transactions, ASHRAE:

Atlanta

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[Abstract:0030][Energy Efficient Buildings] STUDY OF DIFFERENT SCENARIOS OF BUILDING EXTERIOR INSULATION SYSTEM

USING EQUEST

Abdellah Zerroug1, Egils Dzelzitis2. 1. Electrical department, Faculty of technology Sétif1 University

2. Heat Gas and Water technology, Riga Technical University abzerroug@yahou.fr

SUMMARY In this paper, a study of the effect of varying the insulation type and thickness on heat space energy consumption. Different thicknesses of polystyrene, polyurethane and polyisocynurate has been used in three scenarios for each type of insulation material, small, medium and large. The simulation was performed by using the eQUEST software. This software which has been found by previous work to be the most close to the real recorded values for energy consumption. A comparison between the thickness variation with the same insulation and the heat space energy consumption , as well as the comparison between different types of insulation with the same thicknes and heating energy are investigated. Results of the simulation show that when increasing the thickness of insulation results in a decrease of energy consumed,, butut doubling the thickness of insulation does not result in the doubling the savings of energy. INTRODUCTION Jinghua Yu et al [1] use eQUEST software to analyse the effects of envelope factors on energy saving of AC, which included five single strategies of exterior wall thermal insulation absorbance of exterior wall, ratio of window to wall, categories of glazing and kinds of shading system, and two combined strategies. The effect of heat insulating on heating and cooling energy consumption of residential building in hot summer and cold winter zone have been studied by [2]. The influence of residential air conditioning load on the exterior wall heat insulation in hot summer and cold winter zone was studied by [3]. The impact of structure and environment on global energy consumption was developed by [4]. C. Filippin, S. Flores Larsen, [5] analysed the energy consumption patterns in multi-family housing in a moderate cold climate, he found that the winter energy consumption of the multifamily dwellings is lower that of single-family dwelling. Turki and Zaki [6] investigated the effect of insulation and energy storing layers upon the cooling load. Bolatturk [7] calculated the optimum insulation thicknesses, energy savings and payback periods. He used the heating degree-days concept to obtain the annual heating and cooling requirements of building in different climates zones. Durmayaz et al [8] estimated the heating energy requirement in building based on degree hour method on human comfort level. Some researchers used the life cycle cost analysis to optimize the insulation thickness Hasan [9]. The effects of insulation materials on energy saving in Iranian building are studied by Farhanieh and Sattari [10]. Bokos [11] study the comparison in energy savings before and after application of thermal insulation in the exterior envelope. The natural gas consumed by residential heating systems in terms of degree-days is studied by Sarak and Satman [12]. A mathematical model was developed by Sofrata and Salmeen [13] to find the optimal insulation thickness. Mohammed and Khawaja [14] determined the optimum thickness of insulation for some insulating materials used in order to reduce the rate of heat flow to the building in hot countries, and he mentioned that the solar radiation has the most important factor. The effect of climatic zones on the choice of the insulation type and thickness has been studied by Sallal[15] using the life cycle model. The life cycle cost analysis using the degree day was also used by Comakli and Yuksel [16] to investigate the optimum thickness of insulation for coldest cities in Turkey. Daous et all [17], used also life cycle cost analysis in order to determine the optimal insulation thickness under steady periodic conditions. Sisman et al [18] determine the optimum insulation thickness for different degree day region in Turkey for a lifecycle number of years by taking in consideration the thermal conductivity and the price of insulation material, average temperature in the region, fuel price for the heating and the present worth factor PWF. Dombayaci [19] studied the environmental impact of optimum insulation thickness; he used coal as a fuel source, and expanded polystyrene as insulation material. The effect of average electricity tariff on the optimum insulation thickness in building walls by using a dynamic heat transfer model and an economic model based on the present worth method was investigated by Al-Sanea et all [20]. Mahlia et al [21] developed correlation between thermal conductivity and the thickness of selected insulation material s for building wall. Significant economic advantage in energy consumption can be seen by using insulation to achieve high performance building envelope was demonstrated by Lollint [22]. Ozel and Pihtili [23] used an implicit finite difference method for multi-layer wall during winter and summer to obtain the optimum location and distribution of insulation for all wall orientations. S.Ali Hussain [24] Jafri make a review of soon optimum insulation thickness for building envelope, he summarized previous references, the place, the insulation material and thermal conductivity, and components of building envelope. Alexander Gorshkov, et al used the life cycle analysis to assess energy savings delivered by building insulation.

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METHODS. The building envelope construction is composed of three mean parts, roof surfaces, above grade walls, and ground floor, the roof surfaces and above grade walls characteristics are divided into five parts, construction, external finish and colour, exterior insulation, additional insulation and interior insulation. For construction, different scenarios can be fitted from eQUEST library such as wood frame standard, wood frame, wood frame 24 in, wood frame > 24in, metal frame 24 in and, metal frame > 24in, or it can be costumed layer by layer. For external finish we can choose any material from the following: Aluminium, asphalt pavement, clay tile, concrete, felt, bituminous, film, Mylar aluminized, glass spandrel, gravel, marble, roof built up, roofing shingle, steel galvanized bright or weathered, vapour deposited low-e coating, and wood/plywood. For the colour we can choose any colour from dark to white, gloss or flat, or lacquer, we can also choose the colour according to the absorption coefficient. Exterior insulation, polystyrene with standard thickness from 1, 1.5, 2,3,4,5,and 6in may be used in our variables scenarios, then polyurethane with the thicknesses from 1,1.5,2,3,4,5,6in, polyisocyanurate may also be used in our examples with the same standard thickness. Thermal values of polystyrene are ranged from R-4, for a thickness of 1in, to R-30 for a thickness of 6in, values for polyurethane range from R-6 for a thickness of 1in, to a value of 36 for a thickness of 6in, values of thermal resistance of insulation with polyisocyanurate varies from R-9 for a thickness of 1in, to a value of 42 for a thickness of 6in. Additional insulation are expressed in thermal values, and the eQUEST software give standard values of R-7, R-11, R-13, R-15, R-19,, R-21, R-26, R-30, R-38,R-49,and R-60. Ground floor is defined by its exposure, on earth contact, over conditioned space (adiabatic), crawl space, unconditioned space, parking garage, or exposed to ambient conditions, the type of construction of the ground floor can be 2, 4, 6, or 8in concrete, or 1in to 2in plywood underlayment. Exterior cavity insulation can be from ploystyrene , polyurethane, polyisosyanurate, with different thickness ranging from 1,1.5.2,3,4,5in, or with batt insulation with different R values ranging from R-3to R-38. Interior insulation could be from polystyrene, polyurethane, or polyisosyanurate, with different thicknesses and R-values from R-4 to R-2. Light concrete with thickness ranging from 1.25 to 4in can be used as internal finish, different kind of carpet with pad or without pad, fibre or rubber pad, tile from vinyl, ceramic or stones may be used as finish. When the slab penetrates the wall plan, the type of slab insulation could be the same as the insulation materials used before, and the slab edge finish can be aluminium, asphalt, brick, concrete, film, glass, marble, steel, stucco, vapour deposit, or wood/plywood. Building interior construction. Building interior construction is divided in four main parts, the top floor ceiling (above attic), other floor ceiling, vertical walls, and floors. The top floor ceiling is composed of interior finish, framing, Batt insulation, and rigid insulation. Interior finish may be made of lay-In acoustic tile, drywall finish, or plaster finish. Batt insulation, canFramimg is made of wood standard framing, wood advanced framing, or metal stud 24in o.c. The batt insulation that could be added to the top floor ceiling have a standard R values ranging from R-3 to R-60. The rigid insulations that may be fixed on the top floor ceiling are polystyrene, polyurethane, or polyisocyanurate with a thickness of 1in or 1.5in. For other floor ceiling, between each level, interior finish could the same as in the top floor ceiling, lay-In acoustic tile, dry wall finish or plaster finish. The batt insulation R values for other floor ceiling have values of R-11, R-13, R-19, R21, and R-30. Floors are characterized by their internal finish, construction, concrete cab, and their rigid insulation. Internal finish may be made of carpet, carpet with rubber pad, carpet with fibre pad, vinyl tile or ceramic/stone tile. Construction may be made of 2 to 8in concrete, or 1 to 2in plywood underlayment. Concrete cap may be made of 1.25in of light weight concrete to 4in LW concrete. Rigid insulation is made of polystyrene, polyurethane, ployisocyanurate, with thickness ranging from 1in to 3in. Different scenarios. The building chosen to be considered in the following is a multifamily mid rise, and is situated at Ledrugas street, number 7, Riga, Latvia, the weather data file for Latvia is used, the building area is measured in feet (2230.3.76 ft), the number of levels is 5, and as a heating equipment a heating coil, the year 2010 is chosen for the simulation. The roof is not pitched roof, but it is 6in attic above last floor. The roof surfaces construction chosen is metal frame 24 in,o,c. as external finish we choose asphalt pavement weathered, and medium colour of 0.6 abs. For above grade walls, we choose for construction HW concrete of width 4inches, and with no external finish, with a medium absorption of 0.6, and 2 in polystyrene as external insulation. For floor, the exposure over crawl space was chosen, the construction chosen is 4 in concrete, the external insulation 2 in polystyrene R-8, interior insulation used 1 in polystyrene and as a finish 1.25 of LW concrete and vinyl tile, and there was no board insulation or finish selected for slab edge penetring wall plane. The simulation results for electrical energy consumption are given in MWh, but for space heating and domestic hot water, the simulation results are given in gas consumption by one million Btu, or Mbthu.

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Strategies Three mean strategies were adopted for the analysis of the effects of varying insulation type, and thickness on annual heating space energy consumption. The first strategy is to act upon the external building envelope insulation, by changing the thickness and the type, while keeping all the others, interior building insulation and windows type and categories constant. The external building envelope is divided in three parts, exterior insulation for roof, above grade wall (vertical wall), and ground floor. The second strategy is to act upon interior construction, by changing the thickness of batt insulation while keeping all the others as constant. The interior construction is divided in three parts as well, top floor ceiling (under attic), other ceiling, vertical walls, and floors, but vertical walls, other ceiling and floors are kept constant without variation during these scenarios. The third strategy is to act upon exterior windows type and glazing on space heating energy consumption. The external building envelope. For the external building envelope insulation three type of insulation were used, the polystyrene, the polyurethane, and the polisoycynurate with different thicknesses, we choose small, medium and large insulation thicknesses. The following table 1 summarizes the different insulation scenarios. Table 1 sumarizes the differents insulation scenarios.

Scenarios Insulation type Thickness in inches Roof insulation Above grade wall Ground floor 1 Polystyrene thickness 2 2 2 2 Polystyrene thickness 4 3 4 3 Polystyrene thickness 6 3 5 4 Polyurethane thickness 2 2 2 5 Polyurethane thickness 4 3 4 6 Polyurethane thickness 6 3 5 7 polyisocynurate thickness 2 2 2 8 polyisocynurate thickness 4 3 4 9 polyisocynurate thickness 6 3 5

RESULTS. Polystyrene thicknesses effects on space heating energy consumption The results of the comparaison of the space heating annual energy consumption, when acting upon building exterior envelope insolation by changing the thickness of the insolation material while maintaning the other factors constant, to eveluate only the effect of adding more insolation using the same type ( polystyrene in Scenarios 1,2, and 3), In the first scenario, the polystyrene thickness are 2 inches for roof, wall, and floor. In the second scenario, the polystyrene thickness are 4 in for the roof, 3 inshes for the wall, and 4 in for the floor, and in the third scenario 6in insolation for the roof, 3 in for the wall, and 5 in for the floor. The results are given in million Btu and are summarized in Figure 1.

Figure 1. Polystyrene thickness effects on energy consumption Polyurethane thicknesses effects on space heating energy consumption The results of the variation of the thickness of the polyurethane used as insolation material for the building envelope( roof, wall, and floors), on the space heating annual energy consumption when keeping the other factors constant, ,are given in Figure 2. The insolation material thickness used in scenario 4,5, and 6. In the 4 scenario, the polyurethane thickness are 2 inches for roof, 2 inches for the wall, and 2 inches for the floor. In the fifth scenario, the polyurethane thickness are 4 in for the roof, 3 inshes for the wall, and 4 in for the floor, and in the sixth scenario 6in insolation for the roof, 3 in for the wall, and 5 in for the floor. The results are illustrated in Figure 2.

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Figure 2. Polyurethane thickness effects on energy consumption Polyisocyanurate thicknesses effects on space heating energy consumption The results of the variation of the thickness of the polyisocynurate used as insolation material for the building envelope( roof, wall, and floors), on the space heating annual energy consumption when keeping the other factors constant, ,are given in Figure 3. The insolation material thickness used in scenario 7,8, and 9. In the 7th scenario, the polyisocyanurate thickness are 2 inches for roof, 2 inches for the wall, and 2 inches for the floor. In the 8th scenario, the polyisocynurate thickness are 4 in for the roof, 3 inshes for the wall, and 4 in for the floor, and in the 9th scenario 6in insolation for the roof, 3 in for the wall, and 5 in for the floor. The results are illustrated in Figure 3.

Figure 3. Polyisocyanurate thickness effects on energy consumption Comparison of energy consumption for small thickness of different insulations types. The comparison between different insulation materials with the same thickness are shown in the figure 4, the scenario 1 refers to the polystyrene material with 2 inches thickness used for roof, 2 inches for wall, and 2 inches for floor, the scenario 4 refers to the polyurethane insulation material with the same thickness as for the polystyrene. And scenario 7 refers to the insulation material polyisocyanurate of 2 in thickness for roof, wall and floor. The results are illustrated in Figure 4.

Figure 4. Comparison of different insulation type small thickness. Comparison of energy consumption for medium thickness of different insulations types. The comparison between different insulation materials with medium thickness are shown in the figure 5, the scenario 2 refers to the polystyrene material with 4 inches thickness used for roof, 3 inches for wall, and 4 inches for floor, the scenario 5 refers to the polyurethane insulation material with the same thickness as for the polystyrene. And scenario 8 refers to the insulation material (polyisocyanurate) of 4 in thickness for roof, 3 inches for wall, and 4 in for floor. The results are illustrated in Figure 5.

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Figure 5. Comparison of different insulation type medium thickness. Comparison of energy consumption for large thickness of different insulations types. The comparison between different insulation materials with large thickness are shown in the figure 6, the scenario 3 refers to the polystyrene material with 6 inches thickness used for roof, 3 inches for wall, and 5 inches for floor, the scenario 6 refers to the polyurethane insulation material with the same thickness as for the polystyrene. And scenario 9 refers to the insulation material (polyisocyanurate) of 6 in thickness for roof, 3 inches for wall, and 5 in for floor. The results are illustrated in Figure 6.

Figure 6. Comparison of different insulation type medium thickness. DISCUSSION As a conclusion the use of scenario 2 ( 4in for roof, 3 in for wall, and 4 in for floor) as insulation for external envelope is more convinient than scenario 3 (6 in roof, 3in wall, 5 in floor). The mean different percentage of gain is only 2.52%. while the percentage of added thickness of insulation is is 33.33% for the roof insulation and 20% for the floor insulation, the wall insulation remains with the same thickness 3 inches. The use of scenario 5 ( 4in for roof, 3 in for wall, and 4 in for floor) as insulation for external envelope is more convinient than scenario 6 (6 in roof, 3in wall, 5 in floor). The mean different percentage of gain is only 3.66%. while the percentage of added thickness of insulation is is 33.33% for the roof insulation and 20% for the floor insulation, the wall insulation remains with the same thickness 3 inches. The use of scenario 8 ( 4in for roof, 3 in for wall, and 4 in for floor) as insulation for external envelope is more convinient than scenario 9 (6 in roof, 3in wall, 5 in floor). The mean different percentage of gain is only 3.99%. while the percentage of added thickness of insulation remains with the same thickness as in previous scenario. As a conclusion the use of polyurethane and polyisocyanurate as insulation for external envelope seems to be more efficient than using polystyrene, with the same thickness of 2 inches each. The mean percentage of energy saved in space heating when using polyurethane is 32.57%, and when using the polyisocynurate is 37.18% with a difference of 4.62% between polyurethane and polyisocynurate. If the cost difference is large between polyurethane and polyisocynurate, then the most efficient insulation to be used is the polyurethane. The conclusion of the use of polyurethane and polyisocyanurate as insulation for external envelope seems to be more efficient than using polystyrene, with the same thickness. The mean percentage of energy saved in space heating when using polyurethane is 19.45%, and when using the polyisocynurate is 25.02% with a difference of 5.57% between polyurethane and polyisocynurate. If the cost difference is large between polyurethane and poliisocynurate, then the most efficient insulation to be used is the polyurethane. The conclusion of the use of polyurethane and polyisocyanurate as insulation for external envelope seems to be more efficient than using polystyrene, with the same thickness. The mean percentage of energy saved in space heating when using polyurethane

Biçimlendirilmiş: Yazı tipi: 10 nk, Altı çizgisiz, Yazı tipi rengi:Otomatik

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is 19.52%, and when using the polyisocynurate is 25.97% with a difference of 6.45% between polyurethane and polyisocynurate. The most efficient insulation to be used is the polyurethane. REFERENCES

1. H. Yubo, F Xiangzhao, Affection of window-wall ratio on energy consumption in region of hot summer and cold winter, rchitecure Technology 32,(10) 2002,35-36

2. Z. Shihuai, H, Xiadong, W, Xinyun, Analysis the effect of heat insulating on heating and cooling energy consumption of residential buildings in hot summer and cold winter zone, , China Construction Metal Structure 1 ( 2006) 26-29

3. N. Yongfei, L. Zehua, C. Gang, the influence of residential air-conditioning load on the exterior wall heat insulation in hot summer and cold winter zone, Journal of Huaihua University25,5,2006,113-115

4. S.Bilgen, structure and environment impact of global energy consumption, Renewable and Sustainable energy reviews 38(2014) 890-902.

5. Subhash Mishra, JA Usmani, Sanjeev Varshney, Energy saving analysis in building walls through thermal insulation system ,International Journal of Engineering Research and Applications IJERA ISSN -9622, Vol 2, Issue 5 2012, pp128-135

6. A.Al Turki and G.M.Zaki, cooling load response for building wallscomprising heat storing and thermal insulation layers, Journal of Energy Conservation Management 32.1991.235-247.

7. A. Bolatturk, Determination of optimum insulation thickness for building walls with respect to various fuels and climate zones in Turkey, Journal of Applied Thermal Engineering 26,2006,1301-1309.

8. A. Durmayaz, M. Kadioglu, and Z.Sen, an application of the degree hours method to estimate the residential heating energy requirment and fuel consumption in Istanbul, Jornal of Energy 25,2000, 1245-1256.

9. A. Hasan, Optimizing insulation thickness for building using life cycle cost, Journal of Applied Energy, 63,1999,115-124

10. B. farahnieh and S.Sattari, Simulation of energy saving in iranian buildings using integrative modelling for insulation, Journal of Renewable Energy,31,2006,417-425

11. G.C.Bakos, Insulation protection studies for energy savings in residential and tertiary sector, Journal of energy and Buildings,31,2000,251-259.

12. H.Sarak and A. Satman, The degree day method to estimate the residential heating natural gas consumption in Turkey, a case study, Journal of Energy, 28,2003,929-939.

13. H. Sofrata, and B.Salmeen, Selection of thermal insulation thickness, Fourth Saudi Engineering Conference,5,1995,389-399.

14. J.Mohammed, and A.L. Khawaja, determination and selecting the optimum thickness if insulationfor building in hot countries by accounting for solar radiation, Journal of Applied thermal Engineering, 24(17),2004,2601-2610.

15. K.A.Sallal, Comparison between Polystyrene and FiberGlass roof insulation in warm and cold climates, Journal of renewable Energy,28,2003,603-611.

16. K.Komakli and B. Yuksel, optimum insulation thickness of external walls for energy saving, Journal of Applied Thermal Engineering,23,2003,473-479.

17. N. Douas, Z. Hassen, and H.B. Aissia, analytical periodic solution for the study of performance and optimum insulation thickness of building walls in Tunisia, Journal of Applied Thermal engineering,30,2010,319-326.

18. N. sisman, E, kahya, N. Aras, and H.Ara, determination of optimum insulation thickness of external wall and roof for Turkey’s different degree day regions, Journal of Energy Policy, 35,2007,5151-5155.

19. O.A.Dombayci, The environment impact of optimum insulation t5hickness for external walls of buildings, Journal of Building and Environment,42,2007,3855-3859.

20. S.A. al-Sanea, M.F. Zedan, S.A. Al=Ajlan, Effect of electricity tarrif on the optimum insulation thickness in building walls as determined by a dynamic heat transfer model, Journal of Applied Energy, 52, 2005, 313-330.

21. M.I. Mahlia, B.N. Taufi, Ismail, and H.H. Masjuki, Correlation between thermal conductivity and the thickness of the selected insulation materials for building wall, Journal of Energy and Buildings, 39,2007,182-187.

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24. S.Ali Hussain Jefri, P.K. Bharti, M. Jamil Ahmad, Optimum insulation thickness for building envelope A review, IJRET, international journal of research in engineering and technology, volume 4, issue 9, 2015, 214-218.

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[Abstract:0032][Sanitary System Applications] INNOVATIVE ANTIBACTERIAL NANO ADDITIVE SOUND AND NOISE INSULATION

USAGE AT HOSPITAL’S AIR CONDITIONING DUCTS’ INSIDE INSULATION

C. Volkan Dikmen1, Alkan Sancak2

1Mechanical Engineer, DKM Construction and Consultancy, Istanbul 2Chemical Engineer, DKM Construction and Consultancy, Istanbul

SUMMARY A healthy human being’s immune system can resist to microorganisms and viruses that are found in air. Unfortunately people-patients that are present in hospitals have low immune systems. Besides this reaching of bacterias that can happen inside air conditioning ducts to areas which patients are inside will reason different deadly functions. In this article microorganism and bacteria development at sound and heat insulation materials that are used in hygienic ventilating and air conditioning systems distribution elements (air ducts). Bacterias that can happen at air ducts and absorbing surfaces, density of these, spreading of these to hospital areas and effects of this is discussed. As a solution effects of sound and heat insulation materials that are produced from recycled material, antibacterial, nano technologic, hygienic film coated is examined. INTRODUCTION In today’s world insulation becomes more important than ever. Insulation is at the forefront for almost all sectors. Thermal insulation and sound insulation are the leading type of isolation and in these areas, a lot of studies are being done. It is important to say that materials which is providing thermal insulation and sound insulation, must have hygienic properties also. This subject will be investigated in this presantation and the result of the experiments will also being considered Hygiene and Clean Room Hygiene and clean room terms are same in theory but in practice they are diffrent because the hygiene term is related with the micro-organism spices and number, on the other hand clean room term is related with particle number and radius. Clean room or hygienic enviroment term is based on the control of the number of particle and micro-organisms, the minimization of production and access of particle and micro-organisms, control of temperature, humidity, pressure and similar parameters. Some particles can be seen with human eyes, some of them are not. Generally human hair has a 100 micron diameter and also 50 micron diameter particle can be seen with human eyes but a particle with 0.5 micron diameter can not be seen with human eyes. It is shown in figüre 1[1].

Figure 1: Examples of Different Micron Diameters Spread of Particles by Human Beings A person who is only standing, spreads 100000 particle per minute. A person who is only moving head or foot, spreads 500000 particle per minute. When standing up or sitting down, a person spreads 2,5 million particle. A person who is running, spreads 10 million particle per minute. When a person sneezes, spreads 1 million particle, 39.000 of them are microbial. When a person coughs, spreads 5 thousand paritcle, 700 of them are microbial. When making a speech of 100 words, a person spreads 250 particle, 40 of them are microbial[1]. Hospital Infection Hospital infection is a kind of infection which starts at hospital, not before the hospitalisaiton. Hospital infection, generally starts in 48-72 hours after the hospitalization or in 10 days after discharge from hospital. Diffrent studies show that 3%-14% of inmates get hospital infection. It is known that the percent is 4 even in most developed countries. The reduction of micro-organisms from

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200 to 20 in metercube, resulted with reduction of hospital infections from 3,5% to 1,5%. Enviromental factors that triggering hospital infection shown in figure 2. [1].

Figure 2: Enviromental factors that triggering hospital infection Thermal Insulation And Sound Insulation In Clean Rooms Thermal insulation and sound insulation is also important in hospitals and similar places like hygiene. The temperature in clean room is 22 °C (20-24°C) and relative humidity in clean room must be %45 (%40-55). Big air flows are needed in clean rooms and ducts which are transporting the air, have noise problem. Rooms with laminar flows must have the intensity of sound lower than 65dB. Situations in which the intensity of sound is important like in surgeries, owning systems for reduciton the intensity of sound under 40 dB is needed[2]. Noise Criteria For Noise Criteria, related with different frequencies, sound pressure levels have different values. A graph with these values is below[4].

Figure 3: NC Graph[4] Giving a number to identified curve, it is possible to show the criteria with a single number. Sound level criteria is the most used NC curves. Equal sound level changes is used for identifying these curves. In figure 3 NC curves are given[4]. dB Equivalent for Various Sounds Human ear can sense the sounds between 0-140 dB. Some examples of sounds in that area are given below[1];

Whisper : 30 dB

Speech: 40-60 dB

Scream: 80-90 dB

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Airplane Departure: 120-140 dB

Rifle Fire: 130 dB

Sound Level for Hospital Sections In hospitals every different areas have different NC’s. Some examples of that are given below[5].

• Surgery Room, NC25

• Surgery Room in’s and out’s, NC35

• Sterile Material Storage, NC45

• Patient Prepare Room, NC35

• Waking Room, NC25

• Intensive Care, NC25

• Delivery Room, NC30

• Newborn, NC35

• Quarantine Room NC35

• Special Care Room, NC30

Importance Of Hygiene in Air Conditioning Ducts It is already known that, the air which we are breathing contains kinds of bacteria, micro-organisms and pollens. A lot of this paritcles can not be seen by human eye. Air conditioning is very important for the special and hygienic places. Places like surgery room, intensive care room or patient room must have a good air quality and must be controlled regularly. In surgery rooms, patient must be stay away from micro-organisms and bacterias[6]. Examples of Non-Hygienic Air Ducts It is possible to say that the precautions for air ducts are important and when they are not be provided, air ducts are useful places for growth in bacteria and micro-organisms which are harmful for human health. Examples of non-hygienic air ducts are given below.

Figure 4: Example of non-hygienic air duct [7]

Figure 5: Example of non-hygienic air duct [8]

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Figure 6: Example of non-hygienic air duct [8] Factor’s Effecting Micro-organisms and Bacteria Formation Formation of Micro-organisms and bacterias is based on a lot of different factors.

Physical Factors; Heat, radiation, surface tension, osmotic and hydrostatic pressure, humidity, drying, electric

Chemical Factors; Oxygen, Redox Potential, hydrogen ion concentration,

Biological Factors,

Mechanical Factors; Shaking, filtration, centrifuge, force, pressing ve vibration,

[9]. METHOD Tunex (Recycle Rubber Acoustic and Thermal Insulation Hygienic Material) Tunex is an insulation product which is developed with experince and speciality of DKM in sound insulation. It is shown in the figure 7.

Figure 7: Tunex Tunex provides high level sound absorption in structures, mechanical volumes and air conditioning ducts. Tunex has a low thermal conductivity coefficient, as a result of that it also provides thermal isolation. Beeing non-flammable, nature friendly, recyclable and native product (LEED and BREEAM certification high point) makes Tunex attractive.

As a result of measurements, Tunex has high sound absorption coefficient at different frequency.

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Figure 8: 50 mm thickness and 150 kg/m3 density “NBR (Nitrile-Butadiene Rubber) Foam” material sound absorption coefficient change with frequency. Sound Absorption Coefficient of Different Species at Different Frequncies As a result of comparisions of Tunex with other examples based on the change of sound absorption coefficient with frequency one can say that Tunex has similarities in some spaces, on the other hand Tunex has better results in some spaces. The graph is given below.

Figure 9: Sound Absorption Coefficient of Different Species at Different Frequncies Tunex vs Other Acoustic Foam Materials Tunex, has higher sound absorption coefficient than the other acoustic foam materials. Hygienic film coating can make Tunex anti-bacterial and that makes Tunex more special. Tunex is also non-combustible, non-shreded and has no fiber erosion. Anti-Bacterial Properties of Silver Nanoparticle Silver nanoparticles have high anti-microbial activity as a result of passing the cell membrane faster than the other particles. As a result of that, smaller particles have bigger surface area-volume ratio. Bigger ratio makes silver nanoparticles more attractive. Silver nanoparticles can be used very effectively for reducing or destroying the unwanted bacterias and micro-organisims[3,10]. RESULTS Silver nanoparticle added film coated Tunex material, is tested with ISO 22196 standard for antibacterial properties. The report is shown in the figure.

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Figure 10: ISO 22196 Test Report DISCUSSIONS Nanoparticle Reinforced Isolation Material Silver nanoparticle added film coating to Tunex which is already a material for thermal insulation and sound insulation, provides the production of antibacterial, thermal insulation and sound insulation material. Usage of this material in places like hopsital, provides hygienic, thermally insulated and sound insulated areas with very good results. REFERANSLAR 1. Mobedi, M. (2002). Hastane Hijyenik Havalandırma Ve İklimlendirme Sistemleri.

2. Öztürk, H., Atalay, Ö, Yılancı, A., & Dinç, İ. (2005). Tesisat Mühendisliği Dergisi. Hijyenik Havalandırma Sistemleri Ve

Bu Sistemlerin Enerji Bakımından İncelenmesi, 90, 37-45.

3. Thirumurugan, G., & D., M. (2012). Silver Nanoparticles: Real Antibacterial Bullets. Antimicrobial Agents. 4. http://www.engineeringtoolbox.com/nc-noise-criterion-d_725.html 15.01.2016

5. Anıl, O., Arslan, A., Boylu, A., Evren, E., Gacaner, G., Gencer, Ü, . . . Ulutupe, L. (2009). Ix. Ulusal Tesisat Mühendisliği

Kongresi. Hastane Hijyenik Alanlarının Klima Ve Havalandırma Proje Hazırlama Esasları.

6. Gürdallar, M. (2003). VI. Ulusal Tesisat Mühendisliği Kongresi Ve Sergisi. Hijyen Ve İç Hava Kalitesi Bakımından Hvac

Sistemlerinin Temizliği.

7. http://www.ackhvac.com/fiberglass.php 15.01.2016

8. http://inspectapedia.com/Fiberglass/Fiberglass_Insulation_Mold_Causes.php 15.01.2016

9. Arda, M. (2000). Temel Mikrobiyoloji; Genişletilmiş İkinci Baskı. Bakterilerin Üremelerine Etkili Faktörler.

10. Prabhu, S., & Poulose, E. (2012). Prabhu and Poulose International Nano Letters. Silver Nanoparticles: Mechanism of

Antimicrobial Action, Synthesis, Medical Applications, and Toxicity Effects, 2(32).

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[Abstract:0033][Comfort Cooling] INVESTIGATION ON DOUBLE EFFECT DUAL-HEATED ABSORPTION REFRIGERATION

SYSTEM

Kenan Saka1, İbrahim Halil Yılmaz2, Ömer Kanaklı1 1University of Uludağ, Bursa,

2University of Gaziantep, Gaziantep Corresponding email: kenansaka@uludag.edu.tr

SUMMARY The low pressure generator (LPG) of a double effect absorption refrigeration system receives heat energy from the high pressure condenser (HPC) in the literature works. Increasing the capacity of the LPG entails to improve the performance of the double effect absorption refrigeration system however there is a boundary limit where the HPC cannot satisfy the required load by the LPG. In this study, a practical approach to improve the heat capability of the LPG of a series flow double-effect water/lithium bromide absorption refrigeration system has been presented. The proposed approach embraces dual-heating idea which undertakes supplying additional energy source to the LPG from the return of the primary source feeding the HPG. The obtained results reveal that the coefficient of performance (COP) of the system can be raised about 2.5% following this methodology. Keywords: Absorption refrigeration, lithium bromide, low pressure generator, COP INTRODUCTION Absorption refrigeration systems (ARSs) show varieties based on their designs. Researchers have had a chance of investigating different designs by making their own simulation programs. Single and double effect water/lithium bromide ARSs were mostly analyzed in the literature [1−3]. Also, several types of multi-effect absorption cycle have been analyzed such as the triple [4] and quadruple effect absorption cycles [5]. However it must be noted that as the number of effects increases, coefficient of performance (COP) of each effect will not be as high as that of the single effect [6]. Moreover, higher number of effects leads to more complexity in the system. Thus, the double effect cycle takes commonly part in commercial usage [7]. According to ASHRAE (American Society of Heating, Refrigerating and Air-Conditioning Engineers), double effect absorption systems can be classified as series, parallel and reverse parallel flow. The series flow system is simpler relative to the latters. The studies performed on double effect series flow systems for coefficient of performance (COP) were dealt by many researchers. The COP improvement in these systems depends largely on operational arrangement. Increasing either the high pressure generator (HPG) temperature up to a certain level or effectiveness of the solution heat exchangers (SHEs) were shown to improve the COP [8‒11]. Increasing the circulation ratio [9‒11] or solution concentration ratio [10] degrades the COP. Arun et al. [11] revealed the variation in the thermal equilibrium temperature at the low pressure generator (LPG) under varying system parameters including COP based on the HPG temperature. Further, they presented a system operating using dual-heat mode but utilizing a low grade waste heat at the LPG unlike the HPG source. Gomri and Hakimi [12] studied a series flow double effect ARS in their study. They showed the effect of HPG and LPG temperatures on the COP and exergy efficiency. The COP decreased with elevated HPG temperature and increased with rise of the LPG temperature. It was obtained that the maximum COP ranges from 1.2 to 1.3 in their calculations. Kaushik and Arora [13] compared the energy and exergy efficiencies of the single and double effect ARSs. The COP range has obtained 0.6−0.7 and 1.0−1.28 for the single and double effect, respectively. For optimum COP value, while the HPG temperature has been proposed as 91°C in the single effect, it is 150°C for the double effect. In the present study, a series flow double effect ARS using water/lithium bromide fluid pair was studied by concentrating on the high pressure condenser (HPC) and LPG unit. The thermal balance between the HPC and LPG plays an important role in the COP enhancement since the heat rejected from the HPC is utilized as energy source by the LPG. In case the HPC could not provide sufficient thermal load to the LPG, the COP of the ARS would degrade. It is proposed for the condition that backing up the LPG component by the primary source leaving the HPG can furnish the COP. For this reason, a simulation program has been developed to analyze the COP of the series flow double effect ARS thermodynamically. ABSORPTION REFRIGERATION CYCLE The basic components of a double effect ARS are illustrated in Figure 1. The ARSs do not have a compressor as compared to the vapor compression refrigeration systems. The HPG and absorber that provide to benefit from solar energy or industrial waste

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heat are the distinctive components of these systems. The condenser, evaporator, throttling valves and solution pump are the other components. The superheated vapor produced by the HPG is condensed by the HPC, and the rejected heat is used by the LPG as a heat source. The heat rejected by the HPC may not be sufficient for the LPG operating at high temperature in some cases. In this case, the pressurized hot water feeding the HPG provides energy indirectly to the LPG thus the system has double effect heating source. The vapor which is used as a refrigerant within the cycle becomes water after its pressure fell, and later completes the cycle passing from the condenser, evaporator and absorber subsequently. By the way, three pressure levels form in the cycle as high, intermediate and low. The system has three concentration levels, namely weak wX , strong sX and stronger

stX in LiBr solution while circulating in the HPG and absorber. While the ARS is subjected to the heat gain by the HPG, LPG, evaporator and solution pump, this heat gain is rejected in the condenser and absorber by the cooling waters. The evaporator produces chilled water. The solution heat exchangers, SHE I and SHE II, provides the weak solution to enter the HPG at relatively higher enthalpy thus the heat absorbed by the HPG falls. This event improves the COP of the ARS.

Figure 1. Series flow double effect ARS and its components. THERMODYNAMIC ANALYSIS The energy and mass conservation equalities for the HPG are written below

1187 2hmhmhmQ OHswHPG (1)

The equalities for mass balance and concentration,

OHsw mmm2

(2)

ssww XmXm (3)

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The circulation ratios are obtained using equations (2) and (3). The analyses in this study involve two distinct circulation ratios as

ws

w

XXXf

1 (4)

sst

s

XXXf

2 (5)

The circulation ratio provides to define the capacities per unit superheated vapor used. The energy balance between the HPC and LPG has been given merely since these two components were examined in this study.

101152211421 11 hfhfffhffqLPG (6)

1211 hhqHPC (7) The supplementary energy need of the LPG can be calculated by

11

'191919

mhhm

qDH

(8)

The COP of the ARS without using additional heat source is defined as

PHPG

EI wq

q

COP (9)

where PAHPGP fPPw 114 The important thing in double effect ARSs can be considered as the use of heat rejected from the HPG. The additional heat subjected to the LPG is a part of the primary energy source or a kind of waste heat which is an extra energy input to the system can improve the system performance. Thus it should be included in the performance calculations. This additional energy gain can be considered to lower the COP but it leads to increase in the waste heat usage. Both obtained COP values for two cases will be compared in discussion part.

PDHHPG

EII wqq

q

COP (10)

RESULTS AND DISCUSSION The component capacities for an ARS having a 300 kW cooling capacity were compared to a study available in the literature and the results are given in Table 1. However, it is noted that the capacity has been reduced to 100 kW for the latter analyses. The thermodynamic properties of the working fluid were taken from [14−16].

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Table 1. Calculated capacities for the system components. C130HPGT , C80LPGT , C35AT , C35CT , C5ET , 0.7, III , 0.95P

Component Gomri [12] Present study

HPG (kW) 252.41 248.59 LPG (kW) − 156.08 Condenser (kW) 167.20 166.38 Absorber (kW) 385.24 382.66 Evaporator (kW) 300.00 300.00 Pump (kW) 0.00 0.057

SHE I (kW) − 95.76

SHE II (kW) − 154.82

COPI 1.189 1.207 The analysis being made includes some specific assumptions; The system operates under steady-state conditions. Water phase is saturated liquid at the outlet of the condenser. Water phase is saturated vapor at the outlet of the evaporator. Pressure losses in heat exchangers and pipelines are negligible. Reference state for water and air is 25ºC and 1 atm. Capacity is fixed for the evaporator. The HPG is fed by hot saturated water. There is no heat exchange between the system components and surroundings. Figure 2 shows the thermal unbalance between the HPC and LPG as a function of the LPG temperature. Under given conditions, the rejected heat by the HPC provides sufficient energy to the LPG up to 83°C. After this point, the capacity of the LPG increases dramatically, and the energy given by the HPC cannot maintain the thermal balance. This case describes a boundary condition in which the system operates physically normal. The dual-heat concept can provide further system operation widening the operational domain at this point. It is of great importance in terms of the COP rise by operating the LPG at higher temperatures.

Figure 2. Variation in the capacities of the HPC and LPG based on HPC temperature.

40

45

50

55

60

65

70

80 81 82 83 84 85 86

Hea

t cap

acity

(kW

)

LPG temperature (°C)

THPG = 133°C, TA = 35°C, TC = 35°C, TE = 5°C

HPC LPG

127

Figure 3 demonstrates the significance of the LPG temperature in double effect ARS. As the LPG temperature increases, the COPI increases as well. Higher the LPG will increase the COPI much more however the maximum temperature satisfying the thermal balance indicates that further COP rise is not possible shown in Figure 2. The proposed solution for this problem is to support the LPG with an additional thermal load.

Figure 3. COP variation with LPG temperature. In Figure 3, the COPI has been obtained as 1.277 at the uttermost point. When the dual-heating is applied, the COPII has increased to 1.308 which corresponds to 2.5% rise in the COP (see Figure 4). It leads to increase the LPG temperature from 83°C to 84°C even if all the operating conditions are identical. 1°C temperature increment in the LPG forms 107 kj/kg heating load difference between the HPC and LPG. As it is seen from Figure 4, the additional heating does not affect the COPII up to a certain value (100 kj/kg) where the LPG does not show a reaction due to lack of sufficient energy input. It reacts after the threshold value exceeded as clearly seen from Figure 4.

Figure 4. COP variation based on bare and dual-heat mode.

1

1.05

1.1

1.15

1.2

1.25

1.3

1.35

75 77 79 81 83

COP I

LPG temperature (°C)

THPG = 133°C, TA = 35°C, TC = 35°C, TE = 5°C

1.26

1.27

1.28

1.29

1.3

1.31

1.32

25 50 75 100 125 150

COP

Heat supplied (kj/kg)

THPG = 133°C, TA = 35°C, TC = 35°C, TE = 5°C

COPI

COPII

128

It is important to analyze the outlet temperatures of thermal components in the ARSs. The heat exchange between the HPC and LPG components is a significant issue. The outlet temperature of the HPC cannot be lower thermodynamically than the outlet temperature of the LPG. This is generally the accepted agreement in the literature. In this study, it was accepted that the outlet temperature of the hot water used in the HPG exits 140°C and enters at the same temperature to the secondary heater. As it is remarked in Figure 5, the temperature fall is fairly low in the secondary heater since the mass flow rate of the source is high. The fall in the temperature relies on the operating LPG temperature and should be carefully controlled to satisfy the possible conditions. Extracting more heat from the secondary heater allows the COP to rise higher as Figure 5 indicates.

Figure 5. Variation in the outlet temperature of the secondary heater with increasing additional heating. CONCLUSIONS In this study, the internal thermal equilibrium of a series flow double effect ARS was investigated. At higher LPG temperatures, the COP of the ARS becomes higher but the HPC could not meet the energy need of the LPG. This circumstance limits the thermodynamic operational field of the ARS. The problem was partly solved by introducing a practical approach which proposes to heat the LPG using the source leaving the HPG. For this reason, a thermodynamic analysis was performed and simulated by means of a computer program. It has been shown that operating the LPG at higher temperatures can be maintained by the dual-heat concept. Provided that the additional heat supplied is accounted as waste heat and disregarded, the COPI elevates to 1.308 from 1.277 for 1°C temperature increment of the LPG. This means that the COPI can be increased 2.5% following the approach proposed under the given conditions. Nomenclature f circulation ratio h enthalpy, kJ/kg m mass flow rate, kg/s P pressure, kPa q heat capacity, kJ/kg

Q heat transfer rate, kW T temperature, °C

Pw pump work, kW X concentration solution, %

Greek symbol effectiveness efficiency specific volume, m3/kg

139.2

139.4

139.6

139.8

140

100 200 300 400 500

Out

let t

empe

ratu

re (°

C)

Heat supplied (kj/kg)

THPG = 133°C, TA = 35°C, TC = 35°C, TE = 5°C

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Subscript A absorber C condenser DH dual heat E evaporator HPC high pressure condenser HPG high pressure generator

OH 2 water I the first II the second LPG low pressure generator s strong solution st stronger solution w weak solution Abbreviation ARS Absorption Refrigeration System COP Coefficient of Performance HPC High Pressure Condenser HPG High Pressure Generator LPG Low Pressure Generator SHE I Solution Heat Exchanger I SHE II Solution Heat Exchanger II REFERENCES 1. Kaynaklı Ö., Saka K., Kaynaklı F. Absorbsiyonlu soğutma sisteminde farklı eriyiklerin kullanılabilirliği ve performans

değerlerinin incelenmesi. 11. International HVAC+R Technology Symposium, İstanbul, 108−115, 2014. 2. Saka K., Yamankaradeniz N., Kaynaklı F., Kaynaklı Ö. Hava Soğutmalı Çift Kademeli Absorbsiyonlu Soğutma Sisteminin

Enerji Ve Ekserji Analizi. 12. Ulusal Tesisat Mühendisliği Kongresi, İzmir, 1135–1151, 2015. 3. Kaynakli O., Saka K., Kaynakli F. Energy and Exergy Analysis of a Double Effect Absorption Refrigeration System Based

on Different Heat Sources. Energy Conversion and Management, 106: 21–30, 2015. 4. Devault R.C., Marsala J. Ammonia-water triple-effect absorption cycle. ASHRAE Transactions, 96: 676–682, 1990. 5. Grossman G., Zaltash A., Adcock P.W., Devault R.C. Simulating a 4-effect absorption chiller. ASHRAE Journal, 45–53,

1995. 6. Srikhirin P., Aphornratana S., Chungpaibulpatana S. A review of absorption refrigeration technologies. Renewable and

Sustainable Energy Reviews, 5: 343‒372, 2001. 7. Ziegler F., Kahn R., Summerer F., Alefeld G. Multi-effect absorption chillers. International Journal of Refrigeration, 16:

301–310, 1993. 8. Vliet G.C., Lawson M.B., Lithgow R.A. Water–lithium bromide double effect absorption cooling system analysis. ASHRAE

Transactions 5: 811–823, 1982. 9. Kaushik S.C., Chandra S. Computer modeling and parametric study of a double effect generation absorption refrigeration

cycle. Energy Conversion and Management, 25: 9‒14, 1982. 10. Xu G.P., Dai Y.Q., Tou K.W., Tso C.P. Theoretical analysis and optimization of a double-effect series-flow-type absorption

chiller. Applied Thermal Engineering, 16: 975‒987, 1996. 11. Arun M.B., Maiya M.P., Murthy S.S. Equilibrium low pressure generator temperatures for double-effect series flow

absorption refrigeration systems. Applied Thermal Engineering, 20: 227‒242, 2000. 12. Gomri R., Hakimi R. Second law analysis of double effect vapour absorption cooler system. Energy Conversion and

Management, 49: 3343–3348, 2008. 13. Kaushik S.C., Arora A. Energy and exergy analysis of single effect and series flow double effect water–lithium bromide

absorption refrigeration systems. International Journal of Refrigeration, 32: 1247–1258, 2009. 14. Kaita Y. Thermodynamic properties of lithium bromide-water solutions at high temperatures. International Journal of

Refrigeration, 24: 374–390, 2001. 15. Mostafavi M., Agnew B. The impact of ambient temperature on lithium bromide–water absorption machine performance.

Applied Thermal Engineering, 16: 515–522, 1996. 16. Chua H.T., Toh H.K., Malek A., Ng K.C., Srinivasan K. Improved thermodynamic property field of LiBr–H2O solution.

International Journal of Refrigeration, 23: 412–429, 2000.

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[Abstract:0034][Indoor Air Quality and Comfort Conditions] HEAT STRESS, THERMAL EXPOSURE AND COMFORT, AMBIENT NOISE AND LEVEL OF

ILLUMINATION ON WORKPLACES; ANALYSIS OF MEASUREMENT TECHNIQUES WITHIN THE SCOPE OF WORKPLACE SAFETY

Şükrü Onur YİĞİT

TESTO ELEKTRONİK VE TEST ÖLÇÜM CİHAZLARI DIŞ TİC. LTD. ŞTİ Adres: Fulya Mah. Vefa Deresi Sok. Gayrettepe İş Merkezi C Blok No:5/1 D:1-2-3-4-5 34394 Şişli / İSTANBUL Telefon: 0212

217 01 55 E-Posta: oyigit@testo.com.tr

ABSTRACT The law for Work Health and Safety numbered 6331 regulates the duties, responsibilities, rights and obligations of the employers and employees in order to grant the health and security of workers and making the health and security conditions better. According to the law and the related regulations one of the obligations brought to the employer is defining the risks on the working areas. In order to define these risks, thermal exposure and heat stress, thermal comfort, illumination level and sound level are the physical parameters that need to be measured and evaluated. Using different sensor types, measurement and calculation techniques are common on the product design for measuring physical parameter. But for measurements which are directly realted with such an important issue like human health, conformity to the international standards is essential in order to guarantee the traceability. On the perspective of heat stress, thermal exposure and comfort TS EN 27243 (ISO 7243) and TS EN 7730 (ISO 7730), workplace sound level and sound exposure TS ISO 2607 and ISO 1990, illumination level measurements ISO 8995 and DIN EU 5035 norms make the basis of measurement techniques. In this study, the measurement necessity of the mentioned physical parameters, the details for mesurement techniques, measurement site properties, the classification of measurement sites and the norms are examined, some example measurement results are given with the effects of changing ambient and other conditions on the thermal comfort parameters. Key Words: Thermal comfort, Thermal Exposure, Heat Stress, WBGT, PMV/PPD, TS EN 27243, ISO 7730, Illumination, Sound Level INTRODUCTION

Ensuring the workplace safety and protecting the worker’s health are the key elements to protect the most valuable source of the job; the human. The physical structure of the workplace or the workplace conditions according to the properties of the job might affect the worker in short-, middle- or long-run. At this point it must be considered that this effect is sometimes very easily recognized, but sometimes not. For example the dust and toxic gases in a chemical raw material production facility will harm the workers if they are not equipped with convenient and sufficient protection items. Similarly it is also pretty recognizable that in a glass production facility, the high-temperature heat treating ovens create a significant heat stress on the works who are not equipped with appropriate protective clothes or who work more than the allowed safe working time under these conditions. On the other hand if an office of a call center is evaluated, it may not be easily seen that constant exposure to the high noise level will have harmful effects in long term, or the carbon dioxide level due to inadequate ventilation will both decrease the working performance and will cause temporary problems like headache and nauseation. Workplace safety and worker’s health topic should be considered by evaluating multiple factors like ergonomic and structural elements of workplace, psychological factors and so on. Here the effect of workplace conditions is very important because it is possible to evaluate them by measuring and analysis, which results the improvement of conditions in order to protect the worker’s health. In this direction the Law 6331 “İş Sağlığı ve İş Güvenliği Kanunu” indicates that the employer is responsible of making the measurements and analysis of the workplace to be done for the worker’s health and workplace safety concept and accordingly the control, measurement, inspection and research are carried out in order to define the risks that workers are facing [1]. Under the regulations of the law, the physical ambient parameters such as heat exposure, thermal comfort, illumination and noise level should be measured to define the risks of the workplace. This is one of the obligations of the employer. Except that, the measurement of vibration, the ambient dust concentrations, benzene, toluene, xylene, arsenic and similar harmful gas concentrations are obligatory in order to protect the worker’s health in the workplace.

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HEAT EXPOSURE AND THERMAL COMFORT

High ambient temperature on workplace is a common condition in many branches of the industry, especially metal (iron-steel), glass and ceramic industries. Not only these specific ones, in many other facilities of primary industry, factories and other production plants, raw material conditioning facilities, end product quality test institutions, there are various heating systems such as large drying furnaces, constantly working ovens, stoves and incubators, which lead to the hot or very hot ambient conditions. People working in these facilities expose a constantly growing heat stress because of the expansive heat energy. Employees working under high heat stress face not only the well-known problems of high heat stress such as heat stroke, heat exhaustion, cramps caused by temperature, heat spoils, but also under possibly harmful risks because of wet palms, steamy protective glasses, dizziness, hot surfaces and steam because of the high temperature [2]. Another significant problem occurs on human body after the long-term heat exposure is circulatory disorder. Workers under a growing heat stress indoor or outdoor conditions are under a direct risk, at the same time it can be argues that workers above 65 years old, over weight, with heart or blood-pressure diseases and with continuous drug use are under a bigger risk [2]. In order to define the high heat radiation, the heat stress index can be defined by the Wet-Bulb-Globe Temperature (WBGT) measurement. WBGT index is a magnitude of the heat stress somebody has been exposed and should be measured, recorded and averaged along the working hour (generally 8 hours). WBGT index contains the conditions of the ambient air from the heat stress perspective and is defined by measurement of 3 temperature parameters: Radiant Temperature / Globe Temperature (Tg) Heat radiation (expansion) is measured by the globe probe which consists of an empty globe which is capable of absorbing the heat energy at maximum level and a temperature sensor which is located in the center of the globe. The globe temperature (Tg) is determined by the ambient temperature, air flow velocity and thermal radiation. This parameter value generally stands between ambient air temperature and approximately reflected temperature values. Wet Bulb Temperature (Tnw) It is the balance temperature which occurs between the air and a wet surface (here a thread rag) over which the water evaporates. How hot the air, which is not yet saturated by the humidity is, that much water is evaporated from the surface and the cooling will be that much which is caused by the evaporation. Dry Thermometer (Ambient Air) Temperature (Ta) Temperature value of the ambient air. An example digital probe set to make WBGT measurement is given on the Figure-1.

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Figure 1 An example digital probe set to make WBGT measurement

It is obvious that the measurement equipments should comply with international standards in order to take correct, safe and internationally comparable measurement results. The technical specifications of measurement equipments for WBGT measurement are defined on the standard TS EN 27243 [3]. Natural wet-bulb temperature sensor (Tnw):

The shape of the precise part of the sensor : cylindrical The outer diameter of the precise part of the sensor : 6 mm ± 1 mm Length of the sensor : 30 mm ± 5 mm Measuring range : 5 °C - 40 °C Measuring accuracy : ± 0,5 °C The whole precise part of the sensor must be covered with a highly hydrophilic white cord (e.g. cotton). The base of the sensor should have a 6 mm diameter. In order to reduce the heat transfer from the base to the sensor, 20

mm part of it should be covered by the cord The cord should be knitted in arm-shape and cover the sensor carefully. Very tight or very loose coverage is harmful for

the measurement accuracy. The cord should be kept clean. The lover part of the cord must sink in a tank full with distilled water. The length of the free part of the sensor on the air

should be in between 20 – 30 mm. The water tank should be designed in a way that the ambient radiation shall not heat up the water inside.

Globe Temperature Sensor (Tg): Diameter : 150 mm Average Emission Coefficient : 0,95 (darkish black globe) Thickness : As thin as possible Measuring range : 20 °C - 120 °C Measuring accuracy :

− between 20 °C - 50 °C: ± 0,5°C − between 50 °C - 120 °C: ± 1°C

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Ambient Air Temperature (Ta): Ambient temperature as a fundamental parameter can be measured with any convenient method, no matter the shape of

the sensor is. The ambient air temperature sensor should be protected with a structure from the radiation, which especially does not

affect the air flow around the sensor. Measuring range: 10 ºC – 60 ºC Measuring accuracy: ±1 ºC

For the WBGT measurement which is carried out in a mobile way, we face with two important issues; the determination of the measurement spot and the total time of the measurement. The essential thing for determining the measurement spot is the actual working place of the employee. WBGT index is an element of the working place that is mostly defined by the equipments which are needed for this work to be done, and generally it might be not possible to manipulate or change this index. As an example if we consider a foundry, it will mostly not be possible to lower the heat stress around the metal melting pot by reducing the temperature of the pot, because the pot shall reach this temperature value for the melting process. That is why the distance of the measurement spot to the heat source is totally dependent on the borders of the workplace of the employee. Therefore, the workplace should be defined, the spots in which the employee spends maximum time during the work should be specially considered and measurement device should be positioned as close to the heat source as the employee stands. So the measurement device should be positioned on the area of the working place defined by the working procedures of that specific task. The special working spot in which the employee is exposed to the heat at the maximum level should be determined and taken as a must-measurement-spot. Similar to measurement spot, the total time of the measurement is directly related to the working time of the employee. The maximum working time can be defined as the ideal total time for the heat exposure measurement. For the definition of the minimum measurement time, the conditions of the workplace and the measurement equipment should be taken in consideration. Every measuring equipment has an adaptation and reaction time because of its structural design. This time shows how long does the equipment need to be adapted to the conditions of the measurement area after it is brought there. In other words, it shows long does it take that the measuring equipment shows acceptably close values to the real measurement result, which can be taken as a minimum time for the measurement. At the same time if the ambient conditions in the workplace fluctuate frequently and much, the measurement should be carried out in longer time periods. If the conditions are fairly stable during the working hours, some implications can be done via mathematical calculations for the whole working period from shorter measurement results. When the “stable state” is reached, which means that the correct measurement results (not changing results) are evaluated, the measurement can be finished. Measurement specifications related to the heterogeneous ambient If the on the area which is around employee some parameters don’t have a stable value, WBGT index will be defined in 3 different positions which correspond to the height of head, stomach and ankle to the ground level. If the employee is on a standing working position, the measurements will be done respectively 1.7 m, 1.1 m and 0.1 m from the ground level. If the employee is on a sitting working position, the measurements will be done respectively 1.1 m, 0.6 m and 0.1 m from the ground level. The average level of the WBGT value can be calculated with the Formula 1.

2

4 1

If the analysis carried out before the heat stress, on the spot or the area around this spot shows that the place is approximately homogenous (heterogeneous ≤%5), a simplified procedure which points out the use of WBGT value defined by the measurement only at the stomach level can be accepted. No matter the conditions are, if there is a conflict on the interpretation of the analysis, the WBGT index determined by the normal procedure (3 level measurements), shall be taken as reference [3]. The WBGT index is calculated with the formula (2) and (3) after the measurement of the discussed 3 temperature values [3]. In this formula, WBGTS indicates the parameter which should be considered for the measurement places under the sunlight. Without solar radiation (indoor or outdoor): WBGT = 0,7 x Tnw + 0,3 x Tg (2) With solar radiation (outdoor): WBGTS = 0,7 x Tnw + 0,2 x Tg + 0,1 x Ta (3) The evaluation of the WBGT index calculated after the measurements is carried out also in cuh the way that described on the TS EN 27243 (or DIN EN 27243) standard. Here we can divide the employees in 2 groups:

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People who have been acclimatized to heat: Continuous workers on the workplace in which there is heat exposure. People who have not been acclimatized to heat: Temporary workers on the workplace in which there is heat exposure; such as temporary service people on the workplace. The standard also indicates the heat acclimatization method in order to let employees to be adapted to the workplace: “For those who have been acclimatized to heat and who have not, working/resting cycles are defined corresponding to the evaluation of WBGT and reference values on Appendix A according to this standard. The increase on the working hours from one state to the other must be step by step in terms of degrees during a total time period of 7 days” [3]. There should be extra precautions taken to reduce the exposed heat stress, if the limit values defined on the regulations are exceeded. Some examples of these precautions are; extra breaks, the use of appropriate personal protection equipments, reducing the physical metabolic activity. The reference values for the WBGT heat stress index are given on the Table 1 [3].

Table 1 WBGT Heat Stress Index Reference Values

On Table 1, the metabolic rate class is related to the characteristics of the task and indicates the energy which is spent to carry out this task. Some activity examples according to their metabolic rate class are given below: Metabolic rate 0: Inactive resting Metabolic rate 1: Writing, drawing, stitching, driving a car, pricking of small equipments Metabolic rate 2: Driving a truck or similar big construction vehicles, gardening, whitewash, mortar Metabolic rate 3: Constructional work (use of digging tools, bricklaying etc.) Metabolic rate 4: Running, carrying heavy objects, heavily constructional work Let’s interpret the reference table with an example case: on a task which belong to the metabolic rate class 1, the exposed WBGT index of an employee who has been acclimatized to heat should be maximum 30 ºC, the exposed WBGT index of an employee who has not been acclimatized to heat should be maximum 29 ºC (the values corresponds to a measurement time period of maximum 8 hours). There is not a certain indication about the frequency of the measurement. But the measurement should be carried out or replied if there are complaints from employees about the heat stress exposure, the thermal or structural conditions of the workplace are changed. Although some workplaces are under a high heat stress because of the high temperature sources, residential areas such as offices, shopping malls, seminar and congress rooms, supermarkets are considerably easier places for air conditioning compared to the industrial workplaces. One can argue that in such type of working places where the heat stress due to strong heat sources is not significant, other ambient parameters except temperature are also effective. Six parameters should be considered in order to define the thermal comfort of an employee on a workplace. These parameters are; ambient air temperature, ambient air humidity, air velocity independent of the direction (which has effect on the employee), globe/radiant temperature (spread from heat

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sources), the clothing of the employee and the activity carried on the workplace. Four of these parameters are measurable; temperature, humidity, air velocity and radiant temperature, two of them are defined according to the characteristics of the task or the personal choices of the employee; clothing and the activity. Basically it should be considered that the felt ambient conditions vary from one person to the other. The critical question here is such that which should be taken as reference. It may be discussed that in order to minimize the complaints of the employees because of the thermal conditions, four basic principles for ambient air conditioning can be followed;

1.1. The air conditioning system should have been designed according to the circumstances of the workplace and the efficiency of the air conditioning equipments should be measured and controlled periodically.

1.2. The maintenance of air conditioning equipments should be periodically carried out. 1.3. It is not enough to measure only the ambient conditions. Since the main concern is the health and comfort of the

employees, the felt thermal comfort level of the employees should be measured and determined, too. 1.4. All the measurements are carried out in compliance with the international standards.

The national standard about the general thermal comfort is TS EN ISO 7730 standard [4], which was derived by the translation of the international standard ISO 7730 [5]. The standard defines two indices which are evaluated by above mentioned six parameters and therefore make it possible to analytically definition and interpretation of thermal comfort. PMV; Predicted Mean Vote is an index which foresees an average decision value, which bases on the thermal balance of the human body and which was derived by the experiments carried out on a large human test group according to a 7-level thermal sensitiveness scale. This 7-level thermal sensitiveness scale is given on the Table 2.

Tablo 2 7-level Thermal Sensitiveness Scale The thermal balance of the body arises, when the produced internal heat energy of the body is equal to the lost heat energy of the body due to the heat dissemination to the ambient. On average ambient conditions, the thermal balance mechanism of the human body automatically balances the skin temperature with some methods like swelter. The PMV value is calculated with a mathematical formula which has been derived by the empirical data collected from the general experiments on a test human group. So the PMV value makes it possible to determine the completely subjective concept “feeling comfortable” in such the way as objective, measurable, controllable and adjustable. The mathematical formula of the PMV index, the coefficients and parameters in the formula is given on equation (4) [5].

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M: Metabolic rate [W/m2] W: Effective mechanical power [W/m2] Icl: Isolation caused by clothes [m2.K/W] fcl: Surface area factor of clothes ta: Ambient air temperature [ºC] tr: Average radiant temperature [ºC] var: Relative air velocity [m/s] pa: Partial pressure of the water vapor [Pa] hc: Convective heat transfer coefficient [W/(m2.K)] tcl: Temperature caused by the surface of the clothes [ºC] 1 metabolic unit = 1 met = 58,2 W/m2; 1 clothing unit = 1 clo = 0,155 m2.ºC/W Here two artificial units are defined to determine the effect of clothing and the activity, which are respectively “clo” and “met”. The effect of these parameters will be explained more in detail on the following part of the paper. The individual decisions used for deriving the PMV scale lay around the average values and can be used to predict the discomfort level of people and their indication of hot- or cold-feeling in areas they stand. PPD; Predicted Percentage Dissatisfied is an index, which gives the percentage of the people on an environment, who are displeased because of the thermal conditions (people who decide as the environment is too hot or too cold). The formula (5) for the calculation of this index is derived from the decisions of people about the thermal status and explained also in the international standard ISO 7730 [5].

100 95 exp 0,03353 0,2179 5 By using these two indices, they thermal comfort graph to determine the thermal evaluation of working places can be easily drawn. The thermal comfort graph is given in Figure 2.

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Figure 2 Thermal Comfort Graph with Explanations

The activity affects the felt thermal ease of the body, just because it changes the consumed energy on the unit square meter surface of the human body. It is expected that during a seminar, the listeners will probably feel themselves more comfortable than the person who makes the presentation. The consumed energy on the human body according to the activity is given on the Table 3, with their metabolic rates [5].

Table 3 Metabolic Rates, Activity – Energy Relation

Similar to the activity, the clothes worn during the work time affects the thermal comfort, just because it changes the isolation on the unit surface of the human body. It is expected that on the supermarket, if one of the two cashiers wear a vest but the other only t-shirt, the one with the vest will feel more warmly. The heat isolation of the working clothes are given on the Table 4 one by one and as combined [5].

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Table 4 Heat Isolation Coefficients of Clothes

The values which show model design criteria of residential buildings according to the standard ISO 7730 is given on the Table 5 [5].

Tablo 5 The Model Design Criteria For Various Workplaces

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INDOOR AIR QUALITY, TURBULANCE AND CARBON DIOXIDE Indoor air quality and carbon dioxide measurements are not obligatory measurements from the workplace safety and workers health perspective, since these parameters don’t affect the health of workers directly nor create a vital danger for employees. But on the other hand, the feeling comfortable and working efficiently are not only dependent to the thermal conditions. When we consider the air conditioning and ventilation systems, one of the main aim of the systems from home-type split air conditioners to the bigger capacity systems that are used in offices, shopping malls and industrial facilities, is supplying fresh air and therefore creating a healthy and comfortable area for people in there. The air conditioning solution which are equipped with appropriate filters and well-designed exhaust systems, causes removing the bacteria and viruses from the ambient by creating an air turbulence with the help of the objects around , which can be breathed from the ambient air [6]. Appropriate air conditioning is directly related to the comfortably living and efficiently working in offices and residential buildings. Indoor air quality indicates the required quality of the air in residential buildings or around them, in order to sustain the health and comfort conditions of the people in these buildings [7]. In order to ensure the indoor air quality, the gas compounds, particles, microbial pollutants and the thermal comfort conditions of the people must be determined. In indoor environments we can talk about the measurement of four physical parameters to define the air conditioning needs; temperature, humidity, carbon dioxide and air pressure. It has been already discussed in this paper that the thermal comfort feeling is related to the clothes, activity and the ambient conditions, in other words to the body temperature of a person dependent to air temperature, humidity and the air movement. ASHRAE 55(2004) standard points out 30-60 %RH ambient air humidity and 20-26 ºC ambient air temperature for the thermal comfort [8]. Besides international studies show that the survival and spreading of microorganisms on the air is dependent on the relative humidity, temperature, oxygen, wind, air turbulence, water and nutriment content on the air [9]. Which means the measurement of physical parameters of indoor air quality and the degree of air turbulence is necessary not only for a comfortable environment, but also for ensuring the clean air in indoor areas. The turbulence intensity measurement according to EN 13779 standard shows the percentage of air change, in other words the effectiveness and efficiency of the air conditioning. Carbon dioxide is a colorless, unscented gas, sourced by metabolic activities of living creatures and is relatively easy to be measured compared to other pollutant gases. In houses and residential buildings accumulation of carbon dioxide above a certain level causes some symptoms such as state of sleep, headache, and disorder on concentration, which in general is called as Sick Building Syndrome (SBS). In building the main source of the carbon dioxide is human and in order to decrease the complaints due to the fresh air, carbon dioxide concentration should be reduced or it should be completely removed from the environment via an appropriate air conditioning and fresh air supply system. According to the EN 13779 standard the typical CO2 level for indoor areas under IDA4 category is 1200 ppm, for IDA3 category 800 ppm, for IDA2 category 500 ppm, for IDA1 category 350 ppm. Similar to carbon dioxide foreign substances like dust or pollen have a negative effect on indoor air quality. Therefore it is suggested to have filters in air conditioning systems, which hinder the movement of these substances through the indoor area. ILLUMINATION The wave length of the visible light for humans is in between 380 nm and 780 nm. An illumination level measurement device (Lux meter) must evaluate the light exactly at the same way as a human eye does, in order to get meaningful results. Under the workplace safety and workers health point of view, the main concern is the illumination level, because this level is the indicator of the adequate light in workplaces. With the term adequate light, it is mentioned that during the tasks are carried out at the workplace, in the short-run and the long-run, the illumination neither disturbs the worker, nor causes a disease on the eye. After the brief explanation of the measurable parameters on photometry, it will become clearer why it is important to measure the illumination level. The light flux (or luminous flux) is the amount of total light power which a human eye senses and defined with the unit Lumen. The light intensity (or luminous intensity) is a measure of the light flux which spreads from a light source on a solid angle and defined with the unit Candela. But the illumination is a not a subject of the light source, but instead is related to the lightened surface and means the light flux which falls perpendicularly on a unit surface area. It is defined with the unit Lux and the calculated with the mathematical formula (6).

6

Inadequate illumination might cause fatigue and decrease of visual performance on working loner times and on the long-run problems on the health of eye. For workplace safety and workers health measurements, the following regulation and standards

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are taken consideration: “EN 12464-1 Işık ve Aydınlatma – Çalışma yerlerinin aydınlatılması – Bölüm 1: Kapalı Çalışma Alanları”. For some job types, the nominal illumination levels according to the European DIN 5035 norm is given on Table 6.

Table 6 Nominal Illumination Levels For Different Job Fields NOISE Noise level is indicated by the unit decibel, which is calculated as the ratio of pressure caused by sound waves to the reference pressure 20 micro pascal in 20Log basis. The mathematical formula of the unit decibel is given on equation (7).

20 7

Measurement of noise level in workplaces can be evaluated in two concepts. The first one is the noise exposure of workers, which is measured by equipments called dosimeter. These equipments contain a special filter called 1/3 octave band filter, which suppresses all other noises and passes only the noise that a human ear can hear. In order to make the exposure, these devices should be capable of measuring time weighted average value and recording it. The second type of devices are for measuring the noise caused by machines and devices in the workplaces. These measurements are evaluated under the regulation, which is about the protection of workers from the risks caused by the noise level and are detailed with the standard “TS 2607 ISO 1999 Akustik – İş yerinde maruz kalınan gürültünün tayini ve bu gürültünün sebep olduğu işitme kaybının tahmini”. Figure 3 shows both the pressure level and the decibel level of the noise caused by some machines and in different locations/workplaces.

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Figure 3 Schematic Presentation of Noise Levels DISCUSSION Heat stress, thermal comfort, indoor air quality, carbon dioxide, illumination level and noise level are some key physical parameters and concepts to determine the workplace safety and workers health and at the same time which have a significant effect on work quality and efficiency. It is important to measure them in the way described on international standards, control them constantly and periodically, record and document them appropriately. It is argued that making the measurements only for obligatory purposes should never the main approach. In order to make these measurements conveniently, it is required a significant investment, well-qualified and trained staff. As a result of the measurement processes, there should be the aim of creating a quality understanding for the constant improvement of working conditions and the perspective of continuous improvement. REFERENCES [1] RESMİ GAZETE, 30 Haziran 2012 Cumartesi, Sayı: 28339; İş Sağlığı ve Güvenliği Kanunu, Kanun No. 6331 [2] CENTERS FOR DISEASE CONTROL AND PREVENTION, Workplace Safety & Health Topics, Heat Stress, June 2014 [3] TS EN 27243, Sıcak Ortamlar – WBGT (Yaş – Hazne Küre Sıcaklığı) İndeksine Göre Isının Çalışan Üzerindeki Baskısının

Tahmini, Nisan 2002 [4] TS EN ISO 7730 Orta dereceli termal ortamlar- PMV ve PPD indislerinin tayini termal rahatlık için şartların belirlenmesi [5] ISO 7730, Ergonomics of the thermal environment – Analytical determination and interpretation of thermal comfort using

calculation of the PMV and PPD indices and local thermal comfort criteria, 2005 [6] GAMMAGE R.B., Indoor Air and Human Health, Second Edition, Edited by Gammage R.B. and Berven B.A., Lewis

Publishers [7] Indoor Air Quality, Wikipedia; the free encyclopedia, 2014 May [8] ISMAIL et.al.,Indoor Air Quality Issues for Non-Industrial Work Place, December 2010, Universiti Kebangsaan Malaysia [9] PEGAS et.al.,Outdoor/Indoor Air Quality in Primary Schools in Lisbon: A Preliminary Study, September 2009

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[Abstract:0043][Renovation of Existing Buildings (Historical And Other)] ENERGY EFFICIENT RETROFITTING SCENARIOS COMPARISON IN TERMS OF

BUILDING FUNCTION IN TURKEY’S DIFFERENT CLIMATE REGIONS

Inst. K. Ferit Çetintaş1 , Prof. PhD. A. Zerrin Yılmaz2 1 Istanbul Arel University, Istanbul

2 Istanbul Technical University, Istanbul Corresponding e mail: kemalcetintas@arel.edu.tr

SUMMARY As it is known, buildings are responsible from 40% of total energy consumption. Because of that energy efficient in buildings is a key topic in world’s agenda. Energy efficient retrofitting in buildings as important issue such as new building design because of existing building stock. Building function has great effect on building energy consumption, since internal gains change according to building’s function. In this study energy efficient retrofitting scenarios are examined in terms of building function. Aim of the study is to examine energy efficient retrofitting scenarios’ performance in different building function. A reference building is selected which can be used as apartment or office block. According to reference building’s conditions energy efficient retrofitting scenarios have been determined. Scenarios applied in two climatic zones in Turkey. Energy efficient retrofitting scenarios showed different energy performance in office block and apartment. As a result of the study, it can be said that energy efficient retrofitting scenarios should be determined according to building function and climate. 1. Introduction: Energy supply and environmental problems, which are related with energy consumption, are most important topics of world agenda. Buildings are responsible from %40 of world energy consumptions and related carbon emissions [1]. That is why energy consumption in building sector become an important issue for world’s future. Environmental problems such as global warming, which is related with carbon emissions, could be prevent with energy efficiency and using renewable energy sources. Therefore, energy efficiency in building sector is a critical approach for environmental issues. Energy efficiency in building sector has two different sides which they are energy efficient building design and energy efficient retrofitting existing buildings. Improving energy efficiency of existing building as important as energy efficient building design. Countries all over the world have different approach and targets for energy efficiency in building sector. Each country has determined own targets and approach according to the economy and climate. European Union Parliament has established energy performance directive for buildings (EPBD) in 2002 [2]. EPBD has several energy efficiency targets in building sector such as decreasing energy consumption %20 by 2020. Improving energy efficiency of existing buildings encourage in EPBD. In 2010 European Union Parliament recast EPBD and added some new targets such as cost optimal energy efficiency level and zero energy buildings [3]. Cost optimization in energy efficient design strategies become a key point in energy efficiency because of world’ economy. Economic crisis in U.S, Europe and Asia become life cycle cost optimization as important as energy efficiency. Therefore a new methodology on calculation of life cycle cost optimal levels of minimum energy performance requirements for buildings and building element [3]. Energy efficient building retrofitting approach is similar in different building types. But internal gains and system boundary conditions are different for different functions. Especially internal gains from users or equipment change according to building function. That is why aim of the study determine as comparing energy efficient retrofitting scenarios in terms of building function. Office building and dwelling are chosen for comparison the effect of building functions. Energy efficient retrofitting scenarios are effecting energy consumption are studied in same reference building with office or dwelling activities. Energy efficient retrofitting scenarios have been simulated for two different climate regions in Turkey. Thus energy efficient retrofitting strategies can be determined for different climate regions in Turkey. Study also provides that determination of optimum retrofitting scenarios according to building function. Carbon emissions also calculated in this study. Effect of retrofitting scenarios on carbon emissions according to different building function has been studied. Life cycle cost studies of energy efficient retrofitting scenarios are not included in this study. Retrofitting scenarios have been evaluated energy consumption reduction and carbon emissions point of view. 2. Method: In this study a method, which is formed of five different steps, is used for calculation of energy efficient retrofitting alternatives effect on total energy consumption and carbon emissions. Method has individual five steps and details of the steps are given in below. a) Determination of reference building and reference conditions b) Determination of assumptions and boundary conditions which is necessary for energy performance calculations c) Determination of energy efficient retrofitting scenarios d) Selection of energy performance calculation method and calculation tool e) Making calculations and getting results.

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2.1. Determination of reference building an reference conditions: Reference building and reference conditions determination is the key point of energy performance calculations. Reference building provides comparison between design or retrofitting alternatives and reference case. In most of studies reference building determined virtually. In this study reference building formed as virtually. Reference building is formed of according to building stock. Reference building 12 storey building with a basement which is 576 m2 typical floor plan and basement (Figure 1). Same reference building use as completely office or dwelling. Energy efficient retrofitting scenarios performance compare in terms of building function. Reference building physical properties are given in table 1.

Figure 1. Reference building typical floor plan and elevation.

Table 1. Reference building physical properties.

Building element Layers External Wall External paint, Cement plaster (2cm), Brick (19cm), plaster (2 cm.) U: 1,57 W/m2K Roof Gravel (5cm), water insulation (1cm), screed (3cm), thermal insulation (EPS 5 cm), reinforce

concrete slab (20cm), plaster (1cm). U: 0,55 W/m2K Slab (internal) Ceramic tiles (1cm), screed (1cm), reinforce concrete slab (20cm), plaster (1cm). U:

3,44 W/m2K Slab (Fit to ground ) Ceramic tiles (1cm), screed (1cm),thermal insulation (EPS 8cm), reinforce concrete foundation

(20cm), water insulation (1cm), gravel (5cm). U: 0,6 W/m2K İnterior Wall Plaster (1cm), brick (8,5cm), plaster (1cm) U: 2 W/m2K Window Air filled clear double glass pvc window (3+13+3mm) U:2,4 W/m2K Transperency Ratio %25

2.2. Determination of assumptions and boundary conditions which is necessary for energy performance calculations:

Calculation of energy consumption can be complex problem according to chosen calculation methodology. Each calculation methodology has their own boundary conditions and assumptions. Assumptions can be divided into three categories which are climate, building’s physical and users. Climate data which are temperature, humidity, air velocity etc. is one of the necessary data for energy performance calculations. Climate data can be getting from some software such as Meteonorm or from energy performance simulation software database and other database such as ASHRAE [4, 5]. Building physical data includes transparency ratio, building element’s thermal characteristics, thermal zones, lighting system, hvac system etc. These data can be gathered from building’s plans and technical specifications. Lastly assumptions about users have to be determined. According to building function, user density, activity, internal gains from users, internal gain from equipment and building occupancy schedule have to be determinate. Assumptions on users are also complex problem because of lack of data and unpredictable user behaviour. There are some statistical studies and standards on user assumptions such as ASHRAE and BEP [6, 7]. These data tables can be used in calculations or these data can be produced from related statistics. Detailed assumptions, which is necessary for energy performance calculation, are given below in detail (Table 2). Each units in reference building is accepted as an individual thermal zone. Reference buildings has 4 individual unit (use as office or dwelling) and core zone. Primary energy consumption are calculated with national energy conversion factors [7].

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Table 2. Assumption table for dwelling reference dwelling and office building. REFERENCE DWELLING BUILDING

Occupancy Data Assumptions People density / Occupancy schedule 4 person / dwelling unit / Reference [8] Internal gains from users Reference [8]

Lighting Assumptions Lighting instrument and power 20 W fluorescent lamb Automatic control Stepped control

Thermal Comfort Level and Heating System Assumptions Heating and Cooling set point temperature 20 C / 26 C Internal gains from equipment and schedule 19 kWh / m2 Reference [8] Heating system and COP Natural gas central heating with boiler COP: 0,85 Cooling system and COP Chiller (electricity) COP:3

REFERENCE OFFICE BUILDING Occupancy Data Assumptions

People density / Occupancy schedule 12 person for each office unit / BEP- TR [7] Internal gains from users BEP-TR [7]

Lighting Assumptions Lighting instrument and power 20 W fluorescent lamb Automatic control On/Off control Thermal Comfort Level and Heating System Assumptions Heating and Cooling set point temperature 20 C / 26 C Internal gains from equipment and schedule 28 kWh / m2 Schedule from BEP-TR [7] Heating system and COP Natural gas central heating with boiler COP: 0,85 Cooling system and COP Chiller (electricity) COP:3

2.3. Determination of energy efficient retrofitting scenarios:

In this step energy efficient retrofitting scenarios are going to be determined. In determination of retrofitting scenarios used previously studies in literature. Basically retrofitting scenarios are formed of in two groups which they are building shell and lighting system (table 3). Alternatives in lighting system and opaque, transparent element in building shell are combined with each other. All retrofitting alternatives can be seen in table 4.

Table 3. Retrofitting scenarios alternatives. Building Function

Location Lighting System Lighting Control

Insulation Level

Window

Office Dwelling

Istanbul Izmir

20 W Flourescent Lamp 10 W Compact Flourescent

Lamp Led Lamp

None On/Off

Control

None 5 Cm Eps 10 Cm Eps

Double Clear Glass Pvc Frame

Double Low E Glass Pvc Frame

As it is seen in table 3 in this study two different building function in two different climate region in Turkey studied in 108 retrofitting scenarios.

2.4. Selection of energy performance calculation method and tool: Detailed dynamic calculation methodology can give accurate results in energy performance calculations. Therefore in this study detailed dynamic calculation methodology is chosen in energy performance calculations. Design Builder energy performance software, which is used detailed dynamic calculation methodology, in this study. Design builder software is an interface software which is used Energy Plus Software [9].

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Table 4. Energy efficient retrofitting scenarios table. Scenario Lighting System Lighting cont. Insulation level Transparent Element Reference 20 W Fluorescent - - Clear Double Glass PVC Scenario 1 20 W Fluorescent + - Clear Double Glass PVC Scenario 2 10 W Comp Fluorescent - - Clear Double Glass PVC Scenario 3 10 W Comp Fluorescent + - Clear Double Glass PVC Scenario 4 LED Lamp - - Clear Double Glass PVC Scenario 5 LED Lamp + - Clear Double Glass PVC Scenario 6 20 W Fluorescent - 5 cm EPS Clear Double Glass PVC Scenario 7 20 W Fluorescent + 5 cm EPS Clear Double Glass PVC Scenario 8 10 W Comp Fluorescent - 5 cm EPS Clear Double Glass PVC Scenario 9 10 W Comp Fluorescent + 5 cm EPS Clear Double Glass PVC Scenario 10 LED Lamp - 5 cm EPS Clear Double Glass PVC Scenario 11 LED Lamp + 5 cm EPS Clear Double Glass PVC Scenario 12 20 W Fluorescent - 10 cm EPS Clear Double Glass PVC Scenario 13 20 W Fluorescent + 10 cm EPS Clear Double Glass PVC Scenario 14 10 W Comp Fluorescent - 10 cm EPS Clear Double Glass PVC Scenario 15 10 W Comp Fluorescent + 10 cm EPS Clear Double Glass PVC Scenario 16 LED Lamp - 10 cm EPS Clear Double Glass PVC Scenario 17 LED Lamp + 10 cm EPS Clear Double Glass PVC Scenario 18 20 W Fluorescent - - Low E Glass Scenario 19 20 W Fluorescent + - Low E Glass Scenario 20 10 W Comp Fluorescent - - Low E Glass Scenario 21 10 W Comp Fluorescent + - Low E Glass Scenario 22 LED Lamp - - Low E Glass Scenario 23 LED Lamp + - Low E Glass Scenario 24 20 W Fluorescent + 10 cm EPS Low E Glass Scenario 25 10 W Comp Fluorescent + 10 cm EPS Low E Glass Scenario 26 LED Lamp + 10 cm EPS Low E Glass 2.5. Making calculations and getting results:

All retrofitting scenarios energy performance simulate in design builder software. After simulations have been finished all outcomes have organised in a systematic approach for getting results. All results and findings explain in next chapter. 3. Results:

As it is remarked before, retrofitting type effects building primary energy consumption according to its function as it can be seen in results. Each retrofitting scenario has different impact on energy consumption in office and dwelling. As it is seen in figure 2 there is a huge amount of energy consumption difference between reference office building and dwelling. Reference office building energy consumption four times higher than reference dwelling building. Lighting system improvement has great effect on primary energy consumption in retrofitting scenarios. As it is seen figure 2 lighting system improvement, which replaced 20w fluorescent lamb with 10w fluorescent or led lamps, has more efficient in office buildings than dwellings. Office building’s lighting demand higher than dwelling that is why retrofitting alternatives on lighting system more efficient in office buildings. Replacement 20w fluorescent lambs with led lamps and adding lighting control system (scenario 5) has nearly %80 energy saving potential in office building. On the other hand same scenario has nearly %60 energy saving potential in dwelling building. Led lamps has lower internal gain that is why it has decreased building cooling loads but it has increased heating loads. Led lamps provides energy saving in hot climate zones because of decreasing cooling and lighting loads.

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Figure 2. Primary energy consumption of lighting system

Adding insulation layer on external wall is one of major retrofitting approach. Thermal insulation performance on building energy consumption changes according to building internal gains and climate zone. Lamps with high power such as 20w create huge internal gains. Adding insulation layer or increasing insulation layer thickness in high powered lamps alternatives can’t provide energy savings. Heating load is decreased with insulation on the other hand cooling load is increased because of insulation layer. It prevents heat loss from inside to outside so huge amount of internal gains increase cooling load. In figure 3 combination of lighting system improvement and insulation layer retrofitting scenarios effect on primary energy consumption can be seen. As it is seen in figure energy saving potential between reference condition and scenario 16 ( led lamps and 10 cm xps insulation) nearly % 60-70 in office and dwelling building. Energy saving potential in nearly same percentage but amount of energy saving in office building 716 kWh/m2 year and 187 kWh/m2 year in dwelling building in İstanbul. This amount of energy saving is very important for carbon emissions. Office buildings have huge amount of internal gain from people, electrical device and lighting system. Therefore, cooling load is very high in office building. Windows are also important in energy savings in buildings. Changing glass type affect buildings energy consumptions, there is 6 kWh/m2 year energy saving in dwelling building between reference condition and scenario 18 which has just change clear glass with low e glass. On the other hand there is negligible difference in office building between same scenarios. Each retrofitting alternative couldn’t be effective on energy consumption alone. That is why all retrofitting alternatives should be combined with each other to find optimum energy performance improvement. Changing lighting instruments and adding insulation layer has great effect on primary energy consumption in office or dwelling building. Adding lighting control and changing glass type effects on primary energy consumption depends on building function. Energy consumption difference between reference condition and 25th scenario, 161 kWh/m2 year in dwelling and 780 kWh/m2 year in office building. Same retrofitting scenario affect primary energy consumption in nearly five times higher according to building function. As it is remarked before internal gains affect building primary energy consumption greatly. In addition to that, carbon emission has also 4 times saving potential in office building. Different alternatives give different results according to building function. All scenarios primary energy consumption performance can be seen in figure 4.

Figure 3. Combination of lighting system improvement and insulation layer retrofitting scenarios effect on primary energy

consumption kWh/m2 year.

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After comparison of all scenarios according to primary energy consumption in dwelling building 206 kWh/m2 year energy saving potential between reference condition and best scenario (scenario 26). On the other hand in office building 872 kWh/m2 year energy saving potential reference condition and best scenario (scenario 26). Energy saving potential in dwelling building is about %65 and about %80 in office building. Office building has higher energy consumption because of its internal gains that is why, energy saving potential effect its energy consumption broadly.

Figure 4. Primary energy consumption of different retrofitting alternatives kWh/m2 year.

Carbon emissions of all retrofitting alternatives for İstanbul can be seen in figure 5. As it seen in figure especially in office building adding lighting control and changing lighting instruments have great effect on carbon emissions. Electricity consumption is a critical point for energy efficiency. Because it affects primary energy consumption and carbon emission direct and broadly. Same retrofitting scenarios give different results for carbon emissions in office and dwelling building. Office building’s carbon emissions 4 times higher than dwelling building in İstanbul. According to retrofitting scenarios, in dwelling building there is 102 kg CO2 / m2 year and in office building 427 kg CO2 / m2 year carbon emission reduction potential. Same retrofitting alternatives give different energy saving or carbon emission reduction potential. This situation is related with internal gains.

Figure 5. All alternatives carbon emissions ton CO2 / m2 year.

4. Conclusion:

Energy efficient building renovation is one of major target in developing countries such as Turkey. Conventional approach to energy efficient retrofitting couldn’t give expected results. As it is seen in the results of this study, internal gains, affect primary energy consumption and carbon emissions broadly. Same retrofitting scenario gives different results in office and dwelling building because of different internal gain amount. Therefore retrofitting scenarios should be determined according to building function. Changing lighting instruments and adding insulation layers give efficient results according to climate zone. This study shows that different building functions have different potential for energy saving and reduction of carbon emissions. In this study, limited retrofitting scenarios in two different climate zone are examined. Life cycle cost of retrofitting scenarios are not included in this study. Therefore different retrofitting scenarios in different climate zones should be examined from the point of primary energy consumption, carbon emission and life cycle cost in the further studies.

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References: 1- DOE, 2010 Buildings Energy Data Book, Office of Energy and Renewable Energy, US Department of Energy. 2- Directive 2002/91/EC, Directive of the European Parliament and of the Council of 16 December 2002 on the Energy

Performance of Buildings, 2002. 3- Directive 2010/31/EU, Directive of the European Parliament and of the Council of 19 May 2010 on the Energy Performance

of Buildings (recast), 2010. 4- http://www.meteonorm.com/ 5- ASHRAE ‘Standard 169-2013, Climatic Data for Building Design Standards’ 6- https://www.ashrae.org. 7- Ministry of Environment and Urban ‘ Energy Performance Directive of Buildings’ Ankara 2008 8- Yılmaz Z. and others ‘ Reference Building and Calculation Methodology Determination for Tukey Cost Optimal Energy

Efficient Building’ Scientific Research Project 113M596 The Scientific and Technological Research Council of Tukey 2016 9- http://www.designbuilder.co.uk/

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[Abstract:0046][Energy Efficient Buildings] ENERGY MANAGEMENT BY DYNAMIC MONITORING OF A BUILDING OF THE

UNIVERSITY OF VALLADOLID. Rey Martínez, Francisco Javier a; Velasco Gómez, Eloy a; Navas Gracia, Luis Manuel b, González González, Sergio Lorenzo a;

Tejero González, Ana a; Andrés Chicote, Manuel a; Rey Hernández, Javier María; de la Fuente, Samuel c. a G.I.R. de Termotecnia de la Universidad de Valladolid, Valladolid

b Universidad de Valladolid, Valladolid c Tecopysa. Grupo TECOPY, Valladolid

Corresponding email: rey@eii.uva.es

SUMMARY The continuous increase of energy consumption in buildings enhances the importance of implementing energy management systems within the building facilities. These tools allow us to know precisely both energy consumption and use within the building. Monitoring energy consumption provides a clear view not only of the amount, but also of where and when energy is consumed in the building. Besides, a rear analysis of this information allows us to deduce whether there exists an inappropriate consumption, and thus the possibilities of improving building efficiency. A monitoring tool has been implemented within an academic building at the University of Valladolid, applying technological resources of Information Technology and Communication through dynamic monitoring of electrical and thermal parameters. Results obtained are gathered and analysed to directly contribute to improve the use of energy, reduce costs associated with its generation and use, and improve the thermal comfort of the building occupants. INTRODUCTION To ensure an adequate level of indoor conditions for the occupants of the buildings within its different campuses, the University of Valladolid faces an energy cost that entails one of the highest annual budgets of this institution. Consequently, the University of Valladolid is closely engaged with a problem that also affects the environment and the society, beyond the mere economic issue. Some years ago a project called SMART CAMPUS was launched together with the company TECOPYSA, targeting the optimization of the energy efficiency as well as the environmental care of the University Campus. The main interest of this project lies on the high energy consumption and wide range of possibilities of improvement of the case study, through the implementation of a smart and integrated technological tool for the design of energy management networks in sustainable environments. The starting point of the project implementation was a preliminary study on the energy costs and conditions of every building. It was observed that the conditions of one of the buildings where the School of Industrial Engineering is set, together with the information available concerning the involved energy costs, showed that energy efficiency was very low. Comparatively to the remaining buildings of the same campus, this one entrained the highest ratio of electricity consumption and one of the highest ratios in terms of thermal energy. Before the development of this preliminary study, there was no knowledge concerning neither the location of the final energy consumers nor the moment when they consumed such energy. This lack of information also restricted the effectiveness of any measure to be implemented for the improvement of occupant’s comfort and systems efficiency, as well as for the reduction of energy consumption and carbon emissions involved. The consecution of a number of steps, as the gathering of disaggregated consumptions, information analysis, etc. will highlight the adequate practices to optimise the energy use, achieving an efficient exploitation of the available resources. Among the most important functionalities to be achieved through this project, is the possibility of a continuous monitoring of the energy consumption. Being a series of set points targeted precisely, every time they were overpassed the system would send an alarm to the user, informing that a limit of consumption, comfort or cost has been exceeded. In this manner, the energy use can be managed and controlled avoiding superfluous consumptions as well as those occurring beyond operating timetables, which importantly affects the final bills [1]. The solution here proposed can be extrapolated, being possible to stretch the analysis and measures to any building within the campus. The implementation of these systems would result into the acquisition of a high performance level within the international evaluation programs for sustainable certification. This would apply to the whole university campus, comprising all its buildings. It would become a pilot scenario unique in Spain. The main purpose of the project is thus to cover all functionalities and necessities targeted in the university campuses, concerning the knowledge on the final energy consumption, through the implementation of existing technologies directed to the improvement of both efficiency and quality of the offered services. Implementation of this technology will involve the optimization of both human and material resources, contributing to:

a) The improvement of the energy use and occupant’s comfort. b) Reduction of costs derived from energy generation and use. c) Creation of a multi-platform management tool.

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METHODS Target building: use, occupation and energy supply. The selection of the building of the School of Industrial Engineering (EII) was made according to the available information, the technical viability for the setting up of the different measurement devices as well as the opportunities of improvement. The target tertiary building is mainly dedicated to academic and research activities. It was built in the end of the 80s and has a total built area of 20,397 m2 (useful floor area of 16,715 m2), distributed into 4 floors. The main façade faces southwest. With respect to the different facilities, they are characterised by the following points:

Lighting: mainly conventional fluorescents with electromagnetic ballast. Heating: 2 natural gas boilers, 540 kW. Cooling: individual devices (split) distributed throughout the building. DHW: (it does not apply). Other devices: 19 AHUs (Air Handling Units), 3 lifts, computers and laboratory equipment.

Concerning occupation and timetables: Monday to Friday: from 8:00 am to 10:00 pm. Saturday: from 9:00 am to 2:00 pm. The energy supply is provided by:

Electricity: supplied by Iberdrola, with tariffs discriminated in 6 time periods. Natural Gas: supplied by Gas Natural Fenosa. Access tarif: 3.4.

Figure 1: 2014 energy consumption.

System’s implementation and development. A number of steps are developed in order to meet the objectives:

Study of the existing facility. In this step the systems are analysed to know deeply how energy is being generated, distributed and used within the building; both through the thermal and electrical distribution systems. This would enable a global view of the project and the existing possibilities of energy management. Identification of the main points of consumption. Facilities to be monitored are determined in terms of theoretical partial consumptions or power installed, as well as the equipment required for a total energy management of the existing system. Study on the location of sensors for comfort measurement. The most significant spaces are determined for the measurement of the thermal comfort factors, as well as the number of spaces to be monitored. Study of available sensors in the market. A market survey is made to find the most adequate solution for the monitoring and energy management, in economic and technological terms, agreeing with the requirements established in the previous step. Comunication protocols to be used are decided, together with the exact number of devices to be employed, their location, data flow and operation schemes integrated in the existing facility. Installation of the selected sensors. Installation of control and data reception systems into the switchboard. In this step data reception devices and signal concentration systems are installed to receive the information from the sensors installed in the previous step. They are installed in an existing switchboard with enough space or within an extra specific switchboard for energy management. Wiring installation. Interconnection of every element participating into the sensor network, depending of the communication protocol selected for each case.

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Software implementation. Energy management software is provided in one of the local servers or in an additional server with that particular aim. Sensor configuration. Implementation of the sensors installed in the energy management software. Operation checking within a series of tests and validation. The system is then subject to an exhaustive set of tests to check and validate its operation according to the requirements specified in the project. Study of the results obtained and generation of proposals. After the validation, operation checking and data reception during a significant period of time, a study of the obtained data is developed. Then, improvement measures are proposed, combining existing generation systems with renewable sources in order to efficiently supply the building energy demand and reduce costs. In the same way, further measures will be proposed aiming to conceal timetables and energy efficiency policies within the most consuming facilities, or within those consuming more energy than the strictly needed.

Consumption monitoring: parameters and equipment used. It is always intended to measure as many parameters as possible according to the technical and economic restrictions. For the present case, the monitored variables are specified next: Concerning the electric facility, the measured parameters have been: active power, reactive power, power factor, cumulated

consumptions, period excess and THD (global). The points of analysis at different levels are: - Whole-building level (global): electric global consumption; air conditioning consumption in the Left Wing (Heat

Pumps); air conditioning consumption in the Right Wing (Heat Pumps); roof AHUs consumption; outdoor lighting. - Second floor: lighting consumption; power consumption; power consumption at the laboratory QO. - First floor: lighting consumption; power consumption; power consumption at laboratory E and at classroom 1.7. - Ground floor: lighting consumption and power consumption at classrooms B1 and B6. - Basement: lighting consumption, power consumption at the pumps room; power consumption at laboratories CF, MR

and T. Thermal facilities:

- Natural gas consumption in boilers (flow rate). - Water flow rate in the heating primary loop. - Supply and return temperatures in the primary loop.

Water supply facility: - Whole-building level (global): water consumption.

Thermal comfort: - Outdoor air: Dry bulb Temperature (T) and Relative Humidity (HR). - First floor: classroom 1.7 (T, RH, CO2). - Ground floor: classrooms B1 and B7 (T, RH, CO2). - Second floor: Library (T, RH), Laboratory QO (T, RH), Corridor (T, RH). - First floor: meeting room (T, RH), corridor (T, RH), Computers room (T, RH). - Ground floor: entrance hall (T, RH), Classroom B3 (T, RH), Offices 1, 2, 3 and 4 (T, RH). - Basement: Laboratories CF, F and M (T, RH), north corridor (T), Hall (T), South corridor (T, RH).

The figure 2 show devices and sensors installed for the signal gathering, distribution and treatment: Figure 2: Installed equipment

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The partner company TECOPYSA developed a web application that allows the remote access to the monitored variables, also enabling: Visualization, representation and evaluation of every monitored variable. Consequently, consumptions and operation within

the building can be followed in real time. Development of an analysis with the historical data corresponding to any period from its implementation. Generation of predefined and configurable reports. Information exportation in different formats for a detailed study in further applications (spreadsheets, simulation, etc.). Comparison among the various University buildings in terms of the values derived from the measured variables RESULTS The analysis of the monitored variables gives information about the energy consumption. Hence, it has been possible to determine the load profiles of the building, knowing when and where it is consumed, as well as the conditions of the target spaces. An important restriction lies in the gathering of appropriate and useful information about the conditions of the facilities, both thermal and electric. It is possible to link consumptions to particular places within a building, provided that electric networks and switchboards are proper and adequately connected. In some places it has been particularly difficult to differentiate the origin of the different consumptions, due to the number of possible sources. The higher the level of consumer’s differentiation is, the more efficient will be the monitoring. Given that the number of monitoring variables is quite high, the evolution and trends of only some of the most interesting (in terms of electric and thermal consumptions and environmental conditions) are presented next. Concerning electric parameters, the global consumption of the building is shown in Figure 3 through the representation of the power curve. This highlights the important electric consumption out of the occupancy periods, which falls beyond 100 kW during nighttime and weekends. A subsequent analysis from this profile, together with the ones from secondary switchboards of power and lighting, showed that most of the consumption was due to: air conditioning systems (left operating even when the building was closed), and those derived from data centers, experimental devices running in laboratories, circulation pumps of the heating primary loop and security lighting. Moreover, the trend showed that electric energy demand evolved accordingly to the occupation level, as expected.

Figure 3. Electric power consumption of the bulding (global).

Concerning lighting, figure 4 shows the power consumed in the first floor during a typical week. It can be observed that from 05.00 a.m. to 08.00 a.m., when the building starts to be occupied, lighting devices are already at a 50% of their power. The only occupants during this period are the staff of the cleaning service (6 to 8 people), who switch on the lights at a time instead of doing it progressively. This is caused by a lack of zonification in some cases, and by an inadequate use in others.

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Figure 4. Power consumption due to lighting in the first floor (typical week).

With respect to the electric consumptions highlighted, a set of saving and energy efficiency measures were proposed, some of which are already implemented and their efficiency being checked at present [2]. These are:

Substitution of the existing lighting in corridors, halls and toilets by fluorescents T5 with electronic ballast, also reducing and optimizing their number and disposition. Presence detectors were installed in toilets.

Substitution of the lighting outdoors by LED lamps, obtaining a 70% decrease in the power installed at the same time that lightening levels are improved.

Information to the building users with the aim of enhancing their engagement with the efficient use of every facility: switching on the lighting progressively and accordingly to the actual needs, switching off the air conditioning and other systems at the end of the workday, etc.

Provide timetables programmed to connect auxiliary HVAC systems, even enabling the restriction of their use.

Figure 5 shows the monitoring of thermal comfort and air quality parameters (dry bulb temperature and CO2 concentration in this case). It demonstrated that a system to regulate ventilation was required in order to maintain the IAQ within recommendable values. The increase in temperature and CO2 levels in occupation peaks demonstrates that an adequate ventilation system is missed.

Figure 5. Temperature and CO2 concentration in a classroom.

Concerning the monitoring of the heating system’s thermal parameters, a study has been developed to modify the program of boilers availability, which aims to reduce the energy consumption and improve the indoor conditions. A reduction of the operating timetable was thus proposed, adjusted to the occupation period (Figure 6). Boilers could then operate at higher loads with better performances. Based on the monitoring tool, it was possible to develop an analysis of the impact of this measure. This is based on the comparison of results after and before its implementation during weeks with similar outdoor climate conditions. Figure 6 shows that boilers are completely off during weekends and night periods according to the new programmed schedule.

Figure 6. Boilers thermal power during 3 weeks: (object week: new timetable).

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Considering the new data of the boilers energy consumption, an average decrease of 20% during the target week was observed, compared to the previous week of reference. This would result into a decrease of 340,000 kWh in the annual gas consumption CONCLUSIONS The monitoring of the energy consumption and indoor conditions at different levels has been developed within the building of the School of Industrial Engineering of the University of Valladolid, through the implementation of a monitoring tool developed together with the company TECOPYSA. Data gathered has permitted the identification of key energy consumers, enabling the implementation of measures that result into a most efficient operation of existing facilities. Some of these measures have regarded the substitution of existing lighting indoors (including zonification and presence detectors) and outdoors. Besides, boilers operation has been reprogrammed to fit to the actual occupant period, incurring into a annual decrease in the gas consumption of up to a 20%. Furthermore, measures proposed not only have resulted into lower consumptions, but also have demonstrated to meet the thermal comfort targets of the building occupants. The main interest of the installed devices and the consequent study is the extrapolation of the experience to the analysis of further buildings within the University Campus. REFERENCES

[1]. Francisco Javier Rey Martínez y Eloy Velasco Gómez. “Eficiencia energética en edificios. Certificación y auditorías energéticas”. Editorial Thomson. Madrid 2006. (In Spanish)

[2]. Ministerio de Economía y Hacienda. “Estrategia de Ahorro y Eficiencia Energética en España 2004-2012. Sector Edificación”. 2003. (In Spanish).

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[Abstract:0047][Modeling and Software] INVESTIGATION OF RESISTANT WIRES’ RUPTURE PHENOMENON USED IN

FUEL-OILBURNER PRE-HEATERS BY ANALYTICAL AND NUMERICAL RESULTS

Barış Elbüken Alarko-Carrier Research and Development Center, Heating Division, Gebze-Kocaeli/Turkey

baris.elbuken@alarko-carrier.com.tr SUMMARY In this paper, the reason of the rupture event occuring in Fuel-Oil burner pre-heaters’ resistant (heat generating) wires is tried to be inspected analytically and numerically. The dual stage fuel-oil burner (min. capacity 58 kg/h, max. capacity 180 kg/h) and its pre-heater is modelled with heat conduction equation with heat generation plus the continuity equation for single direction alignment (from inlet to the outlet of the heater) plus the Kistiakowsky Equation also taking fuel evaporation in the pipeline as a possibility for both of the calculations. The multi parameter involving heater’s inner temperature distribution behavour has been tried to be examined by a MATLAB script developed by the writer in order to understand the resistant wire exposure to temperature. It is seen that the effect of fuel flowrate is not a major effect on the wires’ fate because of thermostat co-working. The main difference between the calculations is the effect of thermostat cut off function which is not included in the analytical solution but the numerical one. The numerical simulations enlightened the dominant effect of thermostat sensing delay, so the overheating event occuring on the resistant wires. Intolerable delay in thermostat response results with a quick drop in thermal efficiency and an increased possibility on wire rupture. INTRODUCTION The pre-heaters used in the fuel-oil burners are devices responsible for decreasing the viscosity of the fuel to be injected towards the injectors to appropriate values by transferring heat energy to the flowing fuel. The inspected device in this paper consists of 3 sandwich layers assembled adjacently from the fuel inlet side towards the heater exit. Every layer has a steel sheet commonly shared between the pairs. The sandwich layer is made up of fuel coils and a heater plate assembly between them. Two joining central bolts + nuts with wide washers and four tightening bolts + nuts are responsible for holding all the components still with 2 additional high insulator (adiabatic wall) plates at both ends of the assembly. All the three heater plates has an electrical heat production of 3,5 kW which gives a total load of 10,5 kW.

Figure 1.Heater model assembly and sub-components with burner. Any reason for overheating the ribbon shaped resistance wires – whose thickness is 0,3 mm – above the maximum temperature of operation (1300 oC) will probably result with a tear or damage on the ribbons [1]. An example for rupture on the ribbons is seen below.

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(a) (b)

Figure 2.a) Sample image for a ruptured heater plate and wire, b) Plate assembly. Major variations (poor quality of the fuels in the market) in the physical properties of the No:4 Fuel-Oil in Turkey may be a primer cause for the resistant wires’ rupture phenomenon in the heavy oil pre-heaters. Regarding to the MSDS documents of Fuel-Oil No:4, the fuel has a boiling temperature minimum of 177oC and a maximum of 371oC [2]. Various documentations about Fuel-Oils and heavy Fuel-Oils tend to be between nearly 170oC and 650oC [3]. The burner’s fuel pump which is responsible for conveying the fluid fuel-oil through the pre-heater to the injectors depends functionally on the fuel’s viscosity which is also a function of the fuel temperature passing by the pump’s gears. Thus the boiling point can also be taken into account as a corrolative property with the fuel viscosity. The operating thermostat is thought to be responsible for the second main reason for the wire rupture. The thermostat is in a linear dry-contact with 3 objects: a heater plate, a fuel carrying coil and an acordeon spring which is a poor heat sensing construction with low sensitivity. Damage or malfunction of the operating thermostats in the field applications was in common with resistant wires’ rupture phenomenon with high coherence.

(a) (b) (c)

Figure 3.a) Physical model for calculations, b) Coil piping, c) Bottom view and thermostat contact environment (denoted in red square). The heat generated may poorly being transferred to the fuel so that the wires being subjected to excess heat resulting with a mass decrease of the convective heat transfer coefficient resulting the rupture. The generated heat flux is 83,6 kW/m2 for the resistant ribbon wires in operation for the upper formation. It is nearly a guarantee for damage if the core temperature of the wires reach the maximum operating limit, 1300oC [1]. The link for the wire specs can be found at the references section. 347 g’s of molar weight for No:4 fuel has been assumed regarding a weighted distribution of massive hydrocarbons [4, 5]. By the aid of Kistiakowsky Equation [6] the latent heat of vaporisation for the fuel has been tried to be calculated. Different problem scenarios depending on different boiling point variations has been analytically calculated and a boiling start position versus fuel boiling point diagram has been generated. This calculation has been made on a linear stepwise temperature increase basis towards the exit coil with the assumption about the ability of boiling. Depending on the impossibility of determining the triple

157

point and the critical point for the so called fuel because of high deviations in its content, boiling has been assumed to be a possible event in the analytical calculations. Altering the analytical solution assumption, starting from the inlet towards the outlet of the heater, consisting a one dimensional, transient, segregated and single phase flow heat transfer code has been written in MATLAB for examining the system thermostat behavour and the time dependent correct heat distribution omitting the possibility of boiling at 20 bar. Analytical results support the subject of the main hypothesis for the problem: fuel hydrocarbon content and its effect on fuel viscosity and boiling point. In the contrary, the numerical simulations have enlightened the destructive effect of the delay in thermostat sensing. The m file of the code can be downloaded via the link [15] in the references section. Boiling point, kinematic viscosity and density may be counted as the main physical properties of the fuel to be taken into account which could easily affect the heat distribution. It is also noted in spectacular documents about Fuel-Oil that, a vaporized mass percentage is also an experimental expectance so that this percentage has also been adopted to fuel flowrate tuning for the numerical simulations. METHODS Analytical heat transfer model equations with boiling assumption + linear heating The physical structure of the heater has been assumed as an adjacent stepwise heater with every stage having its own temperature mean. The fuel entering the heater from the first stage travels and carries heat towards the following stage also with stealing heat from the resistant wires; thus the fuel flow + heater wires’ energy generation is in a direct competition defining the fuels fate, so the wires’.

Figure 4.One dimensional heat transfer mathematical model for the examined system.

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Here ψ is the pipe free curvilinear path length from the inlet to the starting of the boiling event as a percentage of total. The Kistiakowsky equation (eq. 6) has been a tool for producing a sight for the evaporation energy in the analytical model. The subscripts (liquid,gas) states whether the equation’s region of use (from pre-heater inlet to the boiling point or from the boiling point to the outlet of the device). The subscript LH is for latent heat (evaporation) and res for resultant. The temperature notation between the brackets <T> is denoting the central temperature value which belongs to the plane of symmetry for the model. w is the average molar mass of the fuel (347 g/mol), η is the thermal efficiency of the heater depending on former experiments

(~%74), .n is the molar flowrate and NA is the Avogadro’s number. Tin, Tout and Tboil are the fuel temperatures at the heater inlet,

outlet and boiling point in order. The coefficient of convective heat transfer is a highly important parameter which has a big possibility of determining the occurence of the so called rupture phenomenon. Thus the calculation of the convective heat transfer coefficient has been made depending on the Nusselt thermal boundary layer theory with the below equations [7].

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kc p

Pr , mH GraNuk

hD Pr. ,(10)

DH is the hydraulic diameter of the piping coils, β is the volumetric thermal expansion coefficient, µ is the kinematic viscosity, a and b are model constants which is chosen in order to best define the thermal boundary flow and its orientation with the event geometry. The products are the well known Grashof Number (Gr), Prandtl Number (Pr) and the Nusselt Number (Nu), with k denoting the coefficient of thermal conduction for the fuel and h denoting the convective heat transfer coefficient needed for the numerical simulations. (∆T)δ is the temperature difference between the wall and the boundary layer. The heat transfer model is based on the following heat transfer coefficients for conduction and the specific heat for the fuel-oil. Table 1.Heat transfer parameters for the model materials.

In order to use the Kistiakowsky equation, it was neccesary to determine the average molecular weight for the heat absorbed while phase change. The below table is the content for atmospheric residue heavy Fuel-Oils from the report of the American Petroleum Institute (API) with report no: 201-15368B (2004) which also includes the vapour pressures of each type of species.

159

Table 2.Chemical composition and the vapour pressures for different type of hydrocarbon groups in Fuel-Oil. [4, 5]

Referring to the upper table, it is obvious to mention that, predicting the molecular weight to be constant for Fuel-Oil distillates of different origins is not healthy but also can be stated that, weighted average of hydrocarbon compounds in the content gives a clear reference on molecular weight. Table 3. Examples about possible weighted formations of Fuel-Oil No:4 content. [8]

The molecular average weight of Fuel-Oil No:4 has been taken as 347 g/mol with reference to the same report of API. The analytical model gave an important sight about the wire overheating possibility in case of fuel evaporation. It should be noted that the analytical model does not consist the thermostat cut-off function so that, the results posed in the following pages is a trial for an analytical modelling which takes the multiphase flow as a possibility. Depending on the critical or the triple points of the fuel content, a possible occurence of vaporised fuel flow would also may be noted as a reason for wire rupture. Transient numerical implicit finite volume heat distribution modelling with heat generation in one dimension The vapour pressures of each species type is much more small compared with the pump operating pressure of 20 bar (Table.2). Thus for the numerical simulations, boiling of the flowing fuel was an omitted possibility and has been excluded from the model. This postulate can also be supported by using the Gay-Lussac’s principle for comparing the predicted vapour pressures of each type of species depending on temperature. The maximum vapour pressure @ 25 oC belongs to 7 carboned iso-alkene (C7H16 – Heptane) which is 9000 Pa in Table 3. Even reaching the maximum operational limits of the resistant wire (1300 oC) the pressure multiplier would be no more than 5.3 thus, 20 bar of pumping pressure would be very high enough to keep the molecules in their liquid phase. The heat generated by the resistant wires and the heat transfer between the nodes of the linear model of the pre-heater device, also combined with the convective heat transfer through/from the fuel has been modelled by an implicit (unconditionally stable) one dimensional transient and segregated code which takes fluid mass transfer (incompressible continuity equation) also into account. Thus the momentum equations are excluded from the solution for the ease of adaptation, increased codewriting speed and the structure of the single dimension spatial domain. The code was developed at the Alarko-Carrier R&D Center, Heating Division by the writer.

160

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Above are the equations of continuity for incompressible flow [9], one dimensional and transient heat conduction equation with heat generation [10, 11] and the predicted mass percentage for the naturally evaporated fuel content regarding to the EPA/600/R-03/072 technical report of the United States Environmental Protection Agency, National Exposure Research Laboratory [3]. Here v is for the one dimensional conservative velocity field which is normal to the span of the linear domain (through outwards or inwards) representing the fluid flow from a pipe coil stage to the following. T is for nodal temperature, e’ is for the heat generation power per unit length and t for time in seconds. The equation of continuity is adapted to the simulation by means of information transfer towards the following pipe coil stage at the appropriate timestep corresponding the fuel flowrate explicitly by forward time marching with centering in space. Assuming adiabatic and incompressible flow, the modified Bernoulli equation by the first law of thermodynamics can be used to show the temperature prediction of the n’th coil stage during the segregated mass transfer step of the algorithm.

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B is a magnification constant depending on the timestep size which is responsible for regulating the flow transfer to the adjacent pipe layer with respect to the fuel flowrate. The effect of B can better be understood in the results section on the saw tooth profile of the displayed values. The discretization techniques for the heat equation can be found in numerous textbooks about numerical methods so that we directly give the equations for Euler implicit finite volume discretization which is solved both central in time and space.

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Depending on the fuel hydrocarbon content and the fuel pumping pressure, the ability of conveying the fuel through the circuit towards the injectors may seriously be affected by the fuel temperature passing by the pump gears. Whether depending on a malfunction of fuel depot thermostat or any kind of reason to increase the fuel inlet temperature may have a strong effect on the possibility of wire rupture. Regarding to this information, likewise the mass percentage to be evaporated, the percentage of pumping ability has also been adapted to the numerical model by polynomial curvefitting. By using the viscosity relation to fuel

161

temperature, the fuel pump’s capacity curve [12, 13] values has been transformed into a mass flowrate percentage versus fuel temperature graph for different pump pressures.

(a) (b)

Figure 5.a) Burner fuel pump’s capacity curves [12], b) % Pumping ability depending on fuel temperature. RESULTS Results for analytic solution (58 kg/h fuel-oil) As a result of the analytical model solution, a resistant wire temperature exposure v.s Fuel-Oil boiling point dependance chart has been produced. The possibility of boiling in the fuel carrying pipes can be taken into account for evaluating the gas + liquid fuel-oil entrance to the heater which is also an unwanted situation.

Figure 6.Heater resistant wires’ temperature exposure in case of boiling / multiphase flow. For generating the numerical scenario results, the fuel viscosity – so the pumping ability, thermostat reliability (a measure of correct value reading from the contacting and surrounding media), the thermostat reaction delay, fuel flowrate and the fuel inlet temperature at initial conditions have been chosen as the main simulation parameters. The numerical results have shown that, the temperature overgain is not majorly affected by the fuel boiling point nor the fuel viscosity at the pump’s gears. The effect of these two parameters are not worth plotting again with the fuel inlet temperature. It can also be stated that, fuel flowrate was also not a major parameter affecting the heating behavour (mean temperature at the outlet of the heater) but the effective heater power so the heater efficiency. In the contrary, the effect of the thermostat reliability and the reaction time to read the correct value can barely be underlined as the main safety control parameters for the resistant wires.

162

The below results consist of,

time dependent temperature distribution curves with the last time step result in bold black line, (top left) the sum of total power used for increasing the fuel internal energy from the inlet to the outlet and the power used for

heating the inner space of the heater assembly, (top right) thermostat sensed temperature with the mean outlet temperature of the fluid fuel (bottom left)

the power used for only increasing the fuel internal energy from inlet to the outlet (effective power), (bottom right)

The heater efficiency has been measured depending on previous quality control tests and has been stated earlier. The thermal efficiency of the heater is %74. The effective power is a time series data so that, the mean value has been written by taking the rms value of the series below each graph set.

Results for minimum flowrate (58 kg/h fuel-oil)

Figure 7.1.t_delay=1s, fuel @ inlet=42oC, flow=58kg/h, thermostat set tem.=150oC, rms(U)=1,7kW

Figure 7.2.t_delay=5s, fuel @ inlet=42oC, flow=58kg/h, thermostat set tem.=150oC, rms(U)=1,9kW

163

Figure 7.3.t_delay=20s, fuel @ inlet=42oC, flow=58kg/h, thermostat set tem.=150oC, rms(U)=1,1kW

Figure 7.4.t_delay=40s, fuel @ inlet=42oC, flow=58kg/h, thermostat set tem.=150oC, rms(U)=0,68kW

Results for maximum flowrate (180 kg/h fuel-oil)

Figure 8.1.t_delay=1s, fuel @ inlet=42oC, flow=180kg/h, thermostat set tem.=150oC, rms(U)=7,35kW

164

Figure 8.2.t_delay=5s, fuel @ inlet=42oC, flow=180kg/h, thermostat set tem.=150oC, rms(U)=7,56kW

Figure 8.3.t_delay=20s, fuel @ inlet=42oC, flow=180kg/h, thermostat set tem.=150oC, rms(U)=4,13kW

Figure 8.4.t_delay=40s, fuel @ inlet=42oC, flow=180kg/h, thermostat set tem.=150oC, rms(U)=4,71kW, mid-plates’ temperature exceeds the max. operational limit – 1300oC.

165

DISCUSSION The heat transferred to the flowing Fuel-Oil in a unit time has been plotted vs time and by taking the rms values of the time series data, the resultant calculated values are plotted on the below graphs depending on the thermostat delay behavour. The heater thermal efficiency is calculated as follows and efficiencies at various important operating points are denoted on Figure 9.a.

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(a) (b) Figure 9.Heater effective power (a) and resistant wires’ maximum temperature of exposure (b) due to thermostat delay in sensing.

It is obviously seen that the heater operates in its most efficient state (%72) in case of maximum fuel flowrate with a quick responding thermostat.

High speed response of the thermostat will keep the heater in its highest possible efficient operating range for any constant mass flowrate.

Independent of the fuel flowrate, both the effective power and the heater wires’ exposed temperature values show that, the thermostat can be late to respond no more than 28 seconds. Thus the worst and more efficient operation occurs at this delay value.

Thermostat delay which is more than 28 seconds will also force the system to operate in the unsafe region (T ≥ 1300 oC).

The overdelay in thermostat operation for sensing the correct temperature can be counted as the main reason for the heater resistant wires’ rupture phenomenon.

As a futurework, it is planned to develop the code to a three dimensional and coupled multiphase solver capable of importing universal cad data formats for the wide spectrum of heat transfer and fluid flow problems.

166

ACKNOWLEDGEMENT The problems, literary survey, analytical and numerical modellings and possible solutions to different scenarios with their resuls have been produced in Alarko Carrier Research and Development Center and R&D Heating Equipments Laboratory, Gebze Factory, GOSB, Kocaeli. The so called wire rupture phenomenon has been a research issue from the beginning of a customer complaint at 2014 which was faced at a dual stage Fuel-Oil burner of our brand operating in the field. I would like to thank Mr. Alper ATA (Heating Equipment R&D Manager) and Mr. Mehmet ÖNSEL (R&D Chief) for the valuable informations and experiences they have supplied and shared for this paperwork about fuels, heaters and their sub-systems with their important properties. REFERENCES 1. http://kanthal.com/en/products/materials-in-wire-and-strip-form/strip/resistance-heating-strip/list-of-alloys/ 2. HESS. 2006. No. 4 Fuel Oil Material Safety Data Sheet. MSDS No 15054 (rev 7.1. 2006). 3. Wang, Z, Hollebone, B P, Fingas, M, et al. 2003. Characteristics of Spilled Oils, Fuels, and Petroleum Products: 1.

Composition and Properties of Selected Oils. National Exposure Research Laboratory, Office of Research and Development, United States Environmental Protection Agency, EPA/600/R-03/072.

4. American Petroleum Institute. 2003. Robust Summary of Information on Heavy Fuel Oils. 201-15368B (updated 2004). 5. American Petroleum Institute, Petroleum HPV Testing Group. 2011. Heavy Fuel Oils Category Analysis and Hazard

Characterization. Submitted to the US EPA. Consortium Registration, Interim Final Document. 201-16867A (received 2011 Nov 29).

6. Gilyazetdinov, L P. 1991. Calculation of the Latent Heat of Vaporisation of Petroleum Fractions. Translated from Khimiya i Tekhnologiya Topliv i Masel. No: 12. December 1990. UDC 665.71.035. Plenum Publishing Corporation.

7. Perry, R H, Green, D W, Maloney, J O, et al. 1997. Perry’s Chemical Engineer’s Handbook. Seventh Edition. McGraw-Hill. ISBN 0-07-084941-5.

8. VDI-Verlag GmbH. 1974. VDI-WARMETLAS, Berechnungsblatter für den Warmeübergang. Verlag des Vereins Deutscher Ingenieure – Düsseldorf. ISBN 3-18-400211-X.

9. White, F M. 1991. Viscous Fluid Flow. 2nd Edition. McGraw-Hill. ISBN 0-07-069712-4. 10. Cengel, Y A, Ghajar, A J. 2011. Heat and Mass Transfer: Fundamentals & Applications, Fourth Edition, McGraw-Hill. 11. Patankar, S V. 1980. Numerical Heat Transfer and Fluid Flow. Taylor & Francis. ISBN 0-89116-522-3. 12. http://www.suntec.fr/document/docgenerale/PDFanglais/E1001gb.pdf 13. Curl, H J, O’Donnell, K. 1977. Chemical and Physical Properties of Refined Petroleum Products. NOAA Technical Memorandum ERL MESA-17. Marine Ecosystems Analysis Program. NOAA Environmental Research Laboratories, Boulder, Colorado. Us Dept. Of Commerce. 14. Petrol Ofisi, Madeniyağ Şube Müdürlüğü. 1980. Yakıtlar ve Yağlar. 15. http://researchcompendia.org/compendia/486/

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[Abstract:0054][Comfort Cooling] INVESTIGATION OF THE EFFECT OF AIR FLOW MALDISTRIBUTION ON

EVAPORATOR THERMAL PERFORMANCE

Ergin BAYRAK1,2, Alp Er Şevki KONUKMAN2

1Friterm A.Ş Research and Development Department, Tuzla 34957, Istanbul, Turkey 2Gebze Technical University, Department of Mechanical Engineering, Energy Systems Division,

Gebze 41400, Kocaeli, Turkey Corresponding email:erginbayrak@friterm.com

SUMMARY

Evaporators have some problems such as air velocity, temperature and humidity maldistribution depending on fan characteristics, limitations of heat exchanger and fouling. When designing an evaporator, the parameters taken from manufacturer is used and admitted that air flow where passes on evaporator surface is equal at each point of flow surface. Since this is not possible in practice, this issue directly affects the evaporation rate of internal two-phase fluid. In the literature, the study on airflow maldistribution (AMD) has not been documented very well due to difficulties of elaborate measurements. In the studies performed, this difficulty was handled by changing the path of internal fluid. In this study, effect of inlet air velocity maldistribution on evaporator thermal performance was investigated experimentally. Accordingly, an evaporator, which is consistent with the dimensional limits of laboratory air channel, was designed and identified the distribution and AMD of air velocity passing on each module by using an anemometer. Each circuit of evaporator was considered as an individual module and thermal capacities corresponding to each circuit was calculated by using FrtCoils software considering measurement outcomes. Eventually, non-uniform total capacity was determined and the difference with uniform case was detected. In order to confirm as experimental, the product was tested at calorimetric room at first and then tested at conditioning room ensured same airflow rate but more uniform flow. The capacity difference between experimental and analytical results was seemed to be very close, thus validated the effect of airflow maldistribution on performance. INTRODUCTION The finned tube heat exchangers are used as evaporator and condenser at most of the HVAC (Heating Ventilating Air Conditioning) systems. These devices have some problems due to the incorrect design and the incorrect usage. The most important problems are the non-uniform distribution of both internal and external fluids in the evaporator. In the literature, most of the researchers investigated the effects of internal fluid distribution instead of external flow media distribution. And, practitioners also tried to solve the problem of maldistribution of internal flow. Thereagainst, this study focus on maldistribution of external flow. The effect of maldistribution of external flow on the capacity of the evaporator first investigated computationally via commercial software FrtCoils and then compared with the results of the experiments performed in the Friterm Research and Development Laboratory. The large number of materials may cause in blockage in front of evaporator, which is taken inlet air flow, thereby this issue may induce decreasing air flow rate or non-uniform air flow maldistribution. Datta’s study [1] investigates blockage effect on overall performance of system in automotive cooling systems by creating various type blockage effects. As a result of performing 52 experiments, depending on severity of air maldistribution, cooling capacity decreased by 8.16% for an area blockage of %50 compared with the normal operating condition. Chen [2] has investigated the effect of airflow maldistribution on heat exchanger performance depending on oblique angel of inlet air velocity. Eventually, he has deducted an equation defining the decreasing of evaporator capacity related with oblique angle tightly. Jianying [3] has claimed that some important parameters may be affected some extend depending on AMD. In the course of the experimental analysis, the air velocities were measured at 56 points at inlet of evaporator and determined three different AMD and these degrees according to relative standard deviation formula and then each AMD were entitled as uniform, non-uniform and seriously non-uniform at %18, %49 and %93, respectively. Domanski [4] has considered the effects of air side and refrigerant side maldistribution on the coil capacity. Experimental results have showed that maldistributed air also affects refrigerant distribution, which caused further coil capacity degradation. Another study belonging to the same author [5], it has been detected that the maximum capacity degradation depending on refrigerant and air side are as much as %30 and %8,7, respectively. Besides, [6] and [7] showed that nearly all of the capacity reduction due to non-uniformities in the velocity profile can be recuperated by simply redesigning the tube-to-tube connection sequence. Lee [8]’s numerical and experimental study performed by taking into account of airflow measurement on air cooled condenser by dividing to different segments of air cooled condenser has investigated the effects of different included angles between the air cooled condenser (V type) performance. Consequently, they have detected that changing the angles of coils has considerable effects on airflow distribution and therefore heat transfer performance can be increased %5,29 respectively. Kim [9] has

168

investigated the using of flow balance valve at inlet and outlet of evaporator with respect to AMD in order to recover of loss cooling capacity and COP resulting from non-uniform air flow distribution. For 40% air flow maldistribution, reducing capacity ratio was about %6. Within the scope of this study, an evaporator is designed by considering the test operation as unit cooler used with its fan and DX evaporator used by detaching the axial fan of unit cooler as it can show as image and specification at Figure 1 and Table 1, respectively. Then this product will test at calorimetric room and conditioning room, which have different air maldistribution in front of the evaporators. Thanks to these tests, we would learn not only if air maldistribution identified have an impact on heat transfer rate or not, but also FrtCoils program would verify because this program doesn’t take into account of changing of internal flow characteristic as depend on airflow maldistribution.

Figure 4. General view and dimensional characteristics of tested unit

Table 1. Geometric parameters of tested unit Geometric parameters Values Number of rows 4 Number of tubes per row 22 Transverse tube pitch 35 mm Longitudinal tube pitch 35 mm Tube length 745 mm Tube diameter (inner/outer) 11.86 mm/ 12.5 mm Fin thickness 0.15 mm Fin spacing 7 mm Fin type Flat Fin height 770 mm Tube arrangement inline

METHODS

Two different experimental setups for the capacity measurement and a hot bulb anemometer to identify the air flow distribution formed at the entire face are used in the scope of this study in order to investigate whether the air flow maldistribution has significant effect on heat transfer rate or not. On the other hand, a commercial program entitled as FrtCoils is utilized in order to understand whether flow maldistribution on air side has got any effect on the characteristic of internal fluid. Figure 3 and 4 show the experimental setup and its general schematic perspective including each room together respectively used in this study, including calorimetric and conditioning room. It must be noted that experimental setup illustrated in Figure has been shown for one of the rooms but this diagram is valid for other one too. The refrigerant used in these experiments is R404A. These test setups of two rooms consist of a test unit, air handling unit for conditioning of air, which have humidifier, heaters and centrifugal fans, refrigeration line for regulating of temperature, pressure and flow rate of the refrigerant. The only difference of these room, as the air flow rate can be adjusted with the aid of wind tunnel of conditioning room and calculated of air capacity according to outcomes of sensors. It doesn’t matter for calorimetric room because unit coolers having constant air flow rate have already been used for test operations and the air capacity for this room is calculated from electrical loads.

169

Figure 5. P&I diagram of experimental set-up

Figure 6. Schematic view of experimental set-up

The test processes have been carried out according to Eurovent Standard [10] meticulously. The test results have been taken when the capacity difference between air side capacity and refrigerant side capacity is less then %4. The test duration is about 5 hours. Air flow measurement was carried out at 80 points of each circuit via a hot bulb anemometer, which is only 3 mm diameter and placed at a manual traverse system as shown at Figure 5. The airflow distribution map occurring in front of evaporator surface at each circuit was created according to measurement outcomes. Moreover, the uncertainty value of this measurement was taken into account in the following section of study and specified at Table 2.

170

Figure 7. Hot bulb anemometer used for air flow measurements

Figure 6. Each circuit performed measurement

A mathematical method was adopted in order to describe the air flow maldistribution (AMD) similar as Jianying’s study [3]. This formula is the following:

∑ (1)

Where is the air velocity measured, is the average value of the air velocities measured and is the number of test points. The uncertainty values for the air velocity measurements, temperature and calculated total capacity according to Stephanie Bell’s study [11] is presented below.

Table 2. Uncertainties in experimental measurements

Variable Max. Uncertainty Capacity [kW]

∓2.47%

Air velocity [m/s]

∓(Reading value 0.5+0.03)

Thermocouples (T type)

∓0,3°C

171

RESULTS

Table 3 shows that the values of AMD for both each circuit and entire coil. The first set of results belonging to conditioning room were obtained from the wind tunnel measurements which has no contraction part at inlet of air. In order to ensure more uniform flow, a contraction part for wind tunnel was designed and installed in conditioning room, therefore an improvement trend at outlet temperatures of refrigerant detected. The results without contraction; 6.6°C, 7.8°C, 7.7°C, 6.9 °C and with contraction; 6.6°C, 7.5°C, 7.3°C, 6.5°C. Depending on these results, it was supposed somewhat AMD improvement about by between %2-%5, that is close to %10. Determining of this improvement clearly wasn’t possible due to difficulties of measurements in contraction part. With reference of this detection, test operations and its outcomes were evaluated with respect to this situation. It was seen a deviation of measured velocities between two tests due to accuracy of probe but these values are in the limits of uncertainty range.

Table 3. The value of AMD occurring at each circuits and entire coil

Air Maldistribution Degrees Calorimetric Room Conditioning Room

Circuits Mean Velocity (m/s) AMD Mean Velocity (m/s) AMD

Circuit 1 4,6035 14,75% 3,234854167 16,93%

Circuit 2 4,285 18,43% 3,818677637 10,35% Circuit 3 3,79875 17,17% 4,105001488 12,20% Circuit 4 4,464375 15,95% 3,866578571 15,02% The entire coil 4,28790625 21,6% ∓1,87% 3,756277966 15,7% (∓1,60%

a) b)

Image 1. Airflow distribution maps a) calorimetric room b) conditioning room

During the test process, the most important parameters such as evaporation pressure, air inlet temperature and relative humidity were ensured quite stable as presented in Table 4. Also, the obtained heat transfer rates are very close and in the limits of uncertainty range, that are 7.567 kW and 7.439 kW.

Table 4. The average value of test parameters

Test Room Coil AMD (%) mr (kg/h) Pevap. (bar) Tinlet (°C) RH (%) Q (kW)

Calorimetric Room with fan 17,9 212,1 0,600 10 36,6 7,567

Conditioning Room no fan about 10-13 208,35 0,600 9,99 22,88 7,439

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Figure 8. The test results for each room

The effect of air maldistribution on refrigerant side is seen clearly in Table 5. Owing to more uniform flow media at conditioning room, the outlet temperatures have approached to desired value, which is 6.5°C.

Table 5. The refrigerant outlet temperature values measured via thermocouple

Table 6 and 7 show the capacity outcomes calculated by means of FrtCoils software considering measured air velocities and uniform situation.

Table 6. FrtCoils results according to measurement of conditioning room

Circuits (from up to bottom) AMD (%)

Measured velocity (m/s)

Uncertainty range (m/s)

FRT Capacity considering outlet temp. (kW)

FRT Capacity for uniform cond. (for 3,94

m/s)

1 16,93 3,235 3,42675 1,781 1,9 3,04325 1,689

2 10,35 3,82 4,041 1,856 1,9 3,599 1,767

3 12,20 4,104 4,3392 1,931 1,9 3,8688 1,836

4 15,06 3,86 4,083 1,925 1,9 3,637 1,832 Max. capacity 7,493 7,6 Min. Capacity 7,124

7,38

7,733

7,255

7,622

77,27,47,67,8

Min Max Min Max

Calorimetric Room Conditioning Room

Test Result (kW)

5

5,5

6

6,5

7

7,5

8

8,5

circuit 1 circuit 2 circuit 3 circuit 4

tem

pera

ture

s (°C

)

Desired Outlet Temperatures Calorimetric Room Conditioning Room

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Table 7. Frt Coils results according to measurement of calorimetric room

Circuits (from up to bottom)

AMD (%)

Measured velocity (m/s)

Uncertainty range (m/s)

FRT Capacity considering outlet temp. (kW)

FRT Capacity for uniform cond. (for 3,94 m/s)

1 14,75 4,6035 4,863675 2,059 1,9 4,343325 1,957

2 18,43 4,285 4,52925 1,901 1,9 4,04075 1,833

3 17,17 3,798 4,0179 1,829 1,9 3,5781 1,732

4 15,95 4,464 4,7172 1,955

1,9 4,2108 2,002

Max. capacity 7,744 7,6 Min. Capacity 7,524 Table 8 and Figure 9 demonstrate the percentage of capacity deterioration depending on FrtCoils and experiments separately.

Table 8. The comparison of FrtCoils and experimental results

Max. capacity deterioration (%)

Test Room

Test Result (kW)

FrtCoils result for uniform flow (kW)

FrtCoils result for nonuniform flow

(kW) Frt Nonuniform result-

Frt design capacity Test Nonuniform result-

Frt design capacity Calorimetric Room

Min 7,38 6

7,524 -1,00% -2,89%

Max 7,733 7,744 Conditioning Room

Min 7,255 7,6

7,124 -6,26% -4,54%

Max 7,622 7,493

Figure 9. The percentage of capacity deterioration

-7,00%

-6,00%

-5,00%

-4,00%

-3,00%

-2,00%

-1,00%

0,00%

1,00%

Min Max Min Max

Calorimetric Room Conditioning Room

Capacity Deterioration %

Max. capacity deterioration (%) Frt Nonuniform result-Frt design capacityMax. capacity deterioration (%) Test Nonuniform result-Frt design capacity

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DISCUSSION

In this study, the impact of air flow distribution on evaporator with specific design has been investigated. FrtCoils software and experimental results have verified that although it has been ensured a little improvement at outlet of refrigerant circuits by more uniform air flow, no significant change at heat transfer rate was observed. Hence, the AMD up to about 19% doesn’t have significant impact on the heat transfer rate, which is consistent with Jianying’s study [3]. Therefore, unit cooler investigating in this case study can be used easily without considering air flow effects. Furthermore, the heat transfer rate outcomes taken with FrtCoils and experiments are quite close because of the fact that the AMD identified doesn’t affect the thermal characteristic of internal flow. Further research effort should be considered not only higher AMD but also the blockage types at same AMD, which may occur at practical application. Moreover, the air velocity measurement requires extra experimental effort, so that it should be studied to develop new air flow simulation methods such as CFD tools for typical configurations so as to facilitate the air velocity determination.

ACKNOWLEDGEMENT

Special thanks to Friterm A.Ş. and Dr. Hüseyin Onbaşıoğlu for financial support and assistance to this study.

REFERENCES

1. Datta, S.P., Das, P.K., Mukhopadhyay, S. 2014. Obstructed airflow through the condenser of an automotive air conditioner, effects on the condenser and the overall performance of the system, Applied Thermal Engineering, Vol. 70, pp 925-934.

2. Chen, N., Xu L., Feng, H. D., Yang C. G. 2005. Performance investigation of a finned tube evaporator under the oblique

frontal air velocity distribution, Applied Thermal Engineering, Vol. 25, pp 113–125.

3. Jianying, G., Tieyu G., Xiuling, Y., Dong, H. 2008. Effects of air flow maldistribution on refrigeration system dynamics of an air source heat pump chiller under frosting conditions, Energy Conversion and Management, Vol 49, pp 1645–1651.

4. Lee, J., Domanski, P.A. 1997. Impact of air and refrigerant maldistributions on the performance of finned-tube evaporators

with R-22 and R-407C, Building Environment Division National Institute of Standards of Technology Final Report, U.S. Department of Commerce, Gaithersburg, Maryland.

5. Choi, J.M., Payne, W.V., and Domanski, P.A. 2003. Effects of Non-Uniform Refrigerant and Air Flow Distributions on

Finned- Tube Evaporator Performance, International Congress of Refrigeration 2003, Washington, D.C.

6. Domanski, P.A., Yashar D., Kaufman, K.A., Michalski, R.S. 2004. Optimized design of finned-tube evaporators using learnable evolution methods, International Journal HVAC&R Research, Vol. 10 (2), pp 201-212.

7. Yashar D., Lee, S., Domanski, P.A. 2015. Rooftop air-conditioning unit performance improvement using refrigerant

circuitry optimization, Applied Thermal Engineering, Vol. 83, pp 81-87.

8. Lee, T.S., Wu, W.C., Y.K. Chuah, Y.K. 2010. An improvement of airflow and heat transfer performance of multi-coil condensers by different coil configurations, Int. J. Refrigeration, Vol. 33 (7), pp 1370-1376.

9. Kim, J., Braun, J.E., Groll, E.A. 2008. Analysis of Refrigerant Flow Distribution in Evaporators, International Refrigeration

and Air Conditioning Conference, Paper 966, Purdue University, USA. 10. EUROVENT. 2010. Eurovent Rating Standard for Direct Expansion Forced Convection Unit Air Cooler for Refrigeration.

Eurovent Certification Company, Paris, France.

Bell, S. 1999. A Beginner's Guide to Uncertainty of Measurement, Measurement Good Practice Guide, No.11, Thermal and Length Metrology National Physical Laboratory, Middlesex, United Kingdom.

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[Abstract:0055][Energy Efficient Buildings] GREEN BUILDING APPLICATIONS IN INSULATION INDUSTRY

Ayhan Gökbağ1, Özlem Yıldız1, İrem Teker1

1ISIDEM Yalıtım Sanayi ve Ticaret A.Ş, Türkiye info@isideminsulation.com

SUMMARY Green buildings are based on energy efficiency, reduced environmental impact, economic and healthy buildings and their importance increase day by day for the insulation sector. Today, the energy consumed in buildings constitutes about one-third of the world's energy. This situation dramatically increases the importance of sustainable green building which aims to save energy and nature protection. Most known green building certification systems that have been developed to assess the appropriateness of the concept of green building structure are BREAM from England and LEED from the United States. By such systems scoring is made in different categories, building that are rated above a certain point are awarded green building certification. Reducing the environmental impact of green buildings is occurred by promoting recycling and making arrangements for pollution resulting from waste. These buildings are also healthy that would enable the sunlight and fresh air into the building at the highest level. The most important priority of this kind of buildings is to reduce the energy consumption and this feature is achieved through the contribution of proper insulation. INTRODUCTION Structures that are built in order to meet the one of the most basic needs of human beings, need for housing, are changing our world in significant and irreversible manner. Global climate change and the fact that continuous decrease of basic natural resources confront us every day in many different ways; rising sea levels, increased extreme weather events, worsening of the disease and drought are only a few of them. Larger sizes of natural disasters and too loaded unaffordable energy costs may also be part of our future. Energy resources are consumed to live and about one of third of this energy is spent in buildings [1]. Although the existing successful policies implemented to reduce energy consumption in buildings took place, it just seems to bring local solutions until the green building movement. METHODS In today's world green building movement is born from two questions: What is green building? How can we determine whether a building meets the necessary criteria to comply with this definition? The answer to the first question is simple in theory, in practice it is more complicated though. In fact, the buildings which we call sustainable or green are useful ways to leave a positive impact on the world or at least to reduce a bit of the amount of damage that humankind to nature. Thus, the concept of green building is not just limited to changes made in the construction phase of the building, it is also related with the environment of building and its inhabitants. This choice provides us the opportunity to live in economical, last longer and comfortable buildings. As a concequence of these reasons the concept of green buildings is not a temporary mainstream, it emerges as a radical change in the construction of buildings. In the simplest term, green buildings can be explained by three related objectives:

1. Energy efficiency is the keystone of any green building project. Each building which is well-designed and built according to the concept of green building consumes energy at the lowest possible level and uses renewable energy resources whenever necessary. To built buildings using low energy is just not only beneficial economically, it also brings wider social benefits such as reducing the impact of construction on climate changes, increasing the air quality and decreasing the usage of the energy resources.

2. Conservation of natural resources is the second goal of green buildings. In our day, during construction large amounts of wood, water, metal and fossil fuels are used. Using more durable products and focusing on recycling are structure strategies developed to overcome problems in order to reduce the amount of waste.

3. Increasing the indoor air quality is directly related with the structure of the building. Mold that is usually caused by discharge in pipes or improperly designed heating and cooling systems, lowers the indoor air quality. Another reason for this pollution is some of the chemical gases emitted from building materials

Green building is a systemmatic approach which contains all phases of design and construction that are planning and usage of an area, the selection of materials, energy efficiency and indoor air quality [2]. The latter question that lead the green building act was how it could be understood that a building comply with the definition of green building. The answer to this question was going to put into practice a construction certification system that determines detailed criteria about improved buildings and scores the criteria. In this notice, these systems have been studied to understand green building criteria. The first application of giving certificates was done by BREEAM (Building Research Establishment Environmental Assessment Method) that is launched in England in 1989. This system, that is designed to measure the environmental performance of a

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building, has succeeded in such a short time by bringing a standard definition to the concept of green building and by setting the necessary criterions for the ratings. As being the oldest evaluation system of green buildings, BREEAM has formed a basis for the other certification systems that include LEED. By the year of 2015 there are more than 250000 buildings that are certificated by BREEAM and 1 million more wait on the line to be one. This amount is 20 times more than the ones that have the LEED certificates. BREEAM certificates can be used for any kind of construction and there are various systems that are specially designed for some types of constructions. BREEAM also has the special forms of its applications for the countries like Holland, Norway, Sweden, and Spain. "BREEAM for New Constructions" is one of the building specific certificate systems that includes the consideration of 49 independent subjects. These subjects contains the category of "innovation" as the 10th category, along with other 9 environmental ones. Each subject mentions an environmental influence about the construction and gives some points. These BREEAMS points are given when the construction reaches the highest determinated level of performance of implemantation, for instance when the degree of an harmful effect gets decrased or when the health or welfare level of the people makes progress. The best system for the insulation sector is, no doubt, the one that focuses on the new constructions because the application of the insulation is a lot harder once the construction phase is done and thereby more costly. BBREEAM Rating Benchmarks and BREEAM Environmental Section Weighting and Sample Rating Calculation can be seen in the tables below. Table 1. BREEAM Rating Benchmarks [3]

BREEAM Rating Percentage Score Outstanding 85

Excellent 70 Very Good 55

Good 45 Pass 30

Unclassified 30< Table 2. BREEAM Environmental Section Weighting and Sample Rating Calculation [3]

BREEAM Section Percentage of Weighting Management 12.0%

Health and Well-Being 15.0% Energy 19.0%

Transport 8.0% Water 6.0%

Materials 12.5% Waste 7.5%

Land Use and Ecology 10.0% Pollution 10.0%

Innovation 10.0% The first of the certification systems BREEAM has led to the formation of many new systems and LEED has become one of the most important of these systems originating in the United States and has known worldwide as soon as possible. LEED was first used in 1998 in construction evaluation by USGBC (United States Green Building Council). USGBC can adopt a more active marketing strategy for LEED compared to BREEAM due to donations from member organizations, fees and certification fees and this advantage has also managed to make a name in almost all countries where the work of American origin companies. Since 2008 the USGBC has only supported the LEED at the marketing stage and project register and evaluation have started to be implemented by GBCI (Green Building Certification Institue). According to USGBC's statement in 2015, there are 72,500 LEED-certified building projects in 150 countries [4].

As for the BREEAM, there are also structure specific certification systems are also available for LEED system. The basis of the scoring system is New Building / Great Renovations that is LEED-NC which addresses all scratch builds and large-scale repair projects. The other special systems have been developed for Building Core and Shell, Corporate Interiors, Schools and Existing Buildings. Although, the categories examined are the same for all private systems, score distribution varies within categories [4].

As for BREEAM, we will discussed the new structures which is the area that most interest to the insulation industry during reviewing LEED system categories. The name of this area for the LEED certification system is "New construction and major renovations". The LEED construction rating system includes seven scoring categories. These categories are booked as Sustainable Sites (SS), Water Efficiency (WE), Energy and Atmosphere (EA), Materials and Resources (MR), Indoor Environmental (IEQ), Innovation in Design (ID) and Regional Priority (RP). Distribution points of theese categories and points required to obtain LEED-NC certification is given in the table below [3].

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Table 3. Points Required for LEED-NC 3.0 Ratings [3] LEED-NC Rating Points Required

Platinum 80-110 Gold 60-79 Silver 50-59

Certified 40-49 No Rating 39<

Table 4. LEED Category Allocation for Both LEED-NC 3.0 and LEED-NC v4 [3]

LEED-NC Categories Max Points Sustainable Sites (SS) 26 Water Efficiency (WE) 10

Energy and Atmosphere (EA) 35 Materials and Resources (MR) 14 Indoor Environmental (IEQ) 15

Innovation in Design (ID) 6 Regional Priority (RP) 4 Total Possible Points 110

Today, sustainable building movement has reached an international dimension. There are more than 60 green building council and certification systems which they are implemented that have achieved important goals in their own country. Although Turkey does not have its own certification system, steps are being taken in our country in the name of green building movement and usually benefited from BREEAM and LEED certification systems. According to a statement made by the USGBC in 2014, with 237 LEED-certified buildings, and many evaluation pending projects, Turkey has already taken 9th place among countries in the world [5]. Returning to the importance of green building for insulation sector, the issue of energy efficiency should be examined. In this regard, because it is more prevalent in Turkey, we will focus on LEED criteria. The distribution of LEED-NC rating categories can be seen in Table 4, the highest score with 35 points Energy and Atmosphere (EA) category and category constitutes 32% of the total score. The requirements of this category are provided with insulation as the base. In the LEED-NC system, there are three prerequisites and six sub-categories under the Energy and Atmosphere (EA) category headings. These prerequisites and category names and their corresponding scores are shown in the following table [3]. Table 5. Energy and Atmosphere (EA) Credits and Points under LEED-NC 3.0 [3]

Prerequisite/Credit Name of Prerequisite/Creedit Maximum Points EA Prerequisite 1 Fundamental Commissioning of Bulding Energy Systems - EA Prerequisite 2 Minimum Energy Performance - EA Prerequisite 3 Fundamental Refrigerant Management -

EA Credit 1 Optimize Energy Performance 19 EA Credit 2 On-Site Renewable Energy 7 EA Credit 3 Enhanced Commissioning 2 EA Credit 4 Enhanced Refrigerant Management 2 EA Credit 5 Measurement and Verification 3 EA Credit 6 Green Power 2

Total EA Points Available 35 The first of these prerequisites is taking the building energy systems to fundamental company. Its purpose is ensuring that the conditions required by the building's design team. For this to happen, during the design team fulfilling its mission, must make sure that the possibility of having operated as specified plans and specifications of the structure. In the second prerequisite, maximum efficiency with using minimal energy policy should be adopted. Building designers should show common sense in this regard and must comply with the Energy Cost Budget Method standard requirements. Fundamental Refrigerant Management is the final prerequisite is related to chlorofluorocarbons. Chlorofluorocarbons (CFCs) are ozone disrupts structures in the long-term use of a building's ventilation system. This prerequisite is aimed at protecting the ozone layer by removing chlorofluorocarbons from the structure and it requires no chlorofluorocarbons used in the new HVAC & R systems. In accordance with these prerequisites, designers must submit a plan as a priority relevant to remove these substances from use stepwise.

The optimization of energy performances have the highest score in all LEED systems. The parts of building which related to energy are its outer shell (walls, roof, windows, doors); ventilation equipment; power distribution system; lighting and cooling tools. There are three ways to get credit with these criteria. The first of these options is "Energy Simulation All Buildings". In the LEED Performance Assessment Model, a construction which is recorded as improve by 12%, wins 1 points and each 2% growth

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brings 1 extra point. For the 48% growth, the maximum point that is gained 19. The second option is valid for buildings have under the area of 1860 m2. If the structure is appropriate to the extent prescribed in Advanced Design Guide, can earn 1 point. The third choice is if a building is not a warehouse, laboratory and healthcare structure and has the area under 9290m2, can earn 3 points by planning according to Advanced Structures Key Performance guide which developed by The New Structure Institute.

The second major criteria under the Energy and Atmosphere category is related to the energy sources. LEED, instead of the use of non-renewable energy sources for buildings, encourages the use of renewable energy sources and provides on-site renewable energy systems. In renewable energy used projects in order to get points from this criteria the project performance should be calculated. This performance is determined by the percentage of the annual energy cost of the building to the energy produced. Up to 7 points are available, from 1 point for providing 1 percent of the total building energy requirements to 7 points if 13 percent is provided.

The third criterion, Enhanced Commissioning adds several additional requirements to the Fundamental Building Commisioning category. These requirements include reviewing the energy system design, reviewing contractor submittals, creating a systems manual for building operators, verifying training of operators, and rechecking the building operation within 10 months of occupancy to verify performance.

Forth criterion of Energy and Atmosphere, provides options of decreasing coolers in the construction which are damaging the ozone layer and also indicates how climate changes can affect these substances. Two options to apply are non-use of coolers or choosing a cooler which has the minimum potential effect on ozone layer.

Apart from the aforementioned energy saving criteria, as the 5th credit zone, 3 points are credited to the establishment of a sensor system that provides feedback on the operations in the buildings. The last criterion, "Green Power", which is credited with 2 points, has the feature of being the continuation of the renewable energy criterion. If the energy and resources in the building can not be renewed, another option is to utilize electrical energy from a producer that uses renewable sources. In order to gain points in this criterion, it is obligatory to guarantee a 2 year contract with the green power supplier and to obtain at least 35% of the electricity that will be used in the building from this supplier.

As it could be seen from the Energy and Atmosphere category explanations, energy efficiency credits are crucial in green building rating systems. Insulation is the key to energy efficiency and foundation to a green environment. Buildings that are well insulated also decrease operating costs while increasing energy efficiency. The most important point about choosing the right type and amount of insulation is to implement it during construction process and not after. DISCUSSION

Although BREEAM and LEED systems share the same aim, there are some points in which they differ and these differences are analyzed by the people who would apply for these certificates. Differences between two certification systems are generally not the categories but the way these categories are evaluated. The most basic difference is that BREEAM is more application-oriented and LEED is more result-oriented. For BREEAM, the way for getting better scores is to implement specific applications into the building. For LEED, the application chosen is less important when compared to the result obtained. As a result of this difference in focus, we might say that LEED designers are more flexible. In addition to dissimilarities in application, there are also differences between categories. In general, it is seen that BREEAM categories are more extensive compared to LEED. Only 66% of BREEAM categories are included in LEED, whereas this percentage is 80% for LEED categories. BREEAM experts carry on scientific research to determine the categories. On the other hand, LEED categories are determined through voting where members of USGBC and thousands of people from construction sector have the right to vote. The difference between category selections enables LEED to have no barrier between the system and construction sector, which makes it more preferable especially for international projects. As a matter of fact, it is still a question mark in the construction sector that though it is stated that there are thousands of buildings which are BREEAM certified, there are no detailed explanations about the certification levels of these buildings [6]. When ratings are compared, it is seen that there are specific points for each criterion in LEED system and the total points for the project is the sum of all criteria points. For BREEAM system, equal points could be achieved for each criterion but criterion percentages differ. In this case, the total project points are calculated via the multiplication of criterion percentage and the points taken for that specific criterion. The institutions auditing the certification systems are another major difference. LEED audits are from GBCI (Green Business Certification Inc.) whereas BREEAM audits are people independent of any institution who are educated about the subject and have authorization. The budget given by LEED includes audit costs whereas for BREEAM certification and audit costs are different budget items. For this reason, it is some times more expensive to get BREEAM certification compared to LEED. For both certification systems, there is a point range spared for innovation category and innovation points are given when extensive success is shown in other categories. What separates LEED from BREEAM is that, LEED also have regional priority

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order. As a result, LEED enables to obtain more points from some criteria according to the environment in which the project is assembled. Unfortunately this system is only applicable to projects within US for now because regional priority topics for other countries are yet not determined by USGBC. This situation is a disadvantage for LEED considering the international projects. The focus of insulation sector is definitely the energy efficiency credits and BREEAM has more criteria at this point compared to LEED. To get full score from LEED, 50% efficiency according to given standards is enough. For BREEAM however, only net-zero energy buildings get full credits. Since even the 50% required for LEED saves a lot of energy, this criterion by BREEAM is quite compelling. Another similarity between systems is that both give importance to energy modeling. It is an obligation to perform energy modeling of large buildings for LEED and BREEAM gives more points for energy modeling when there are also obtain energy efficiency points with alternative applications. The most advantageous practice of BREEAM is to use European standards. LEED criteria is defined according to American standards but if standards in your own country are more strict, you are given the right to use national standards. The fact that both certifications have no expiration date is a similarity. Only for buildings in “Extraordinary” BREEAM status, there is a need to get “BREEAM In-use” certificate every three years. USGBC declared that they would ask for energy and water consumption information just for the records. Another similarity is that both certification systems gives great importance to documentation and projects with improper documentation are not taken into account [3]. Considering all points compared, it is not possible to say one system is better than the other. According to building type, the stage at which application is considered, which certification level is targeted and budget, both systems have advantages and disadvantages. There also exist other alternatives however LEED and BREEAM systems are still more favorable since their competition improves both systems each day.

REFERANSLAR 1.United Nations Environment Programme – Sustainable Buildings and Climate Initiative. 2009. Buildings and Climate Change. Summary for Decision Makers. Web. 2.Johnston, David, and Scott Gibson. 2008. Green from the Ground Up: The Guide to Healthy, Sustainable, and Energy-efficient Construction. Newtown, CT: Taunton. 3.Kibert, Charles J. 2008. Sustainable Construction: Green Building Design and Delivery. Hoboken, NJ: John Wiley & Sons. 4.US Green Building Council. 2015. Green Building Facts. 5.Star Gazetesi. 2014. Türkiye Çevreci Binalarda İlk 10’da. 6.Inbuilt. 2010. BREEAM versus LEED.

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[Abstract:0056][Heating, Climatization and Air-conditioning Applications in Buildings] HEAT TRANSFER MEASUREMENTS FOR NON-DARCY FLOW IN 10-PPI METAL FOAM

NIHAD DUKHAN1, ÖZER BAĞCI2, ALTAY ARBAK2, MICHEL DE PAEPE3

1Department of Mechanical Engineering University of Detroit Mercy, Detroit, MI 48221, USA

2Makina Fakültesi Istanbul Technical University

Gümüşsuyu, 34437, Istanbul, Turkey 3Department of Flow, Heat and Combustion Mechanics

Ghent University Sint-Pietersnieuwstraat 41, B-9000 Ghent, Belgium Corresponding email: nihad.dukhan@udmercy.edu

SUMMARY Metal foam is a class of porous media with very high porosity (around 90%) and a large surface area density. The foam internal structure is web-like of thin ligaments surrounding cells that are open to flow. This structure promotes thermal dispersion because it offers a lot of mixing of a flowing fluid. Break up and inception of the boundary layers are phenomena adding to the convective heat transfer. The thermal conductivity of the solid phase of metal foam is also high. Because of all these attributed, metal foam is an excellent heat exchanger technology. There is a need for more experimental data regarding heat transfer in metal form. In this paper, experimental heat transfer data for water flow in commercial open-cell aluminum foam cylinder heated at the wall by a constant heat flux (14,998 W/m2 and 26,347 W/m2), is presented. The foam had 10 pores per inch (ppi) and a porosity around 87%. The measurements included wall temperature along flow direction as well as average inlet and outlet temperatures of the water. Flow speeds were in the non-Darcy regimes: transitional and Forchheimer flow regimes. The behavior of the wall temperature clearly shows thermal development conditions. The experimental Nusselt number is presented as a function of axial distance in flow direction, and showed what seemed to be periodic thermal development. The experimental data can be used for validation of other analytical solutions. The results can also be used to verify numerical models and metal-foam heat exchangers used in air-conditioning for example.

INTRODUCTION Open-cell metal foams are excellent candidates for heat exchanger designs. The foams have relatively high thermal conductivities and very large surface area per unit volume. The internal structure of the foam interferes with fluid flow and causes disturbance by disrupting the boundary layers repetitively, which enhances convection heat transfer between the internal solid surface of the foam and the fluid. Various aspects of manufacturing the foam and its applications, as well as heat transfer in the foam, can be found in [1].

The literature contains many experimental studies for heat transfer inside the foam. Most of these studies used air as the working fluid [e.g. 2-6], while few others used water [7-9]. However, many sets of experimental data for heat transfer in metal foam employed small foam sample sizes (or at least small dimension in one direction) or a short length in the flow direction relative to flow area hydraulic diameter, which makes their results specific to the samples tested, as the data may contain unassessed size and/or entry and exit effects.

There is a need for more experimental data for convection heat transfer in metal foam. The current research team has been involved in researching convection heat transfer due to water flow in various commercial open-cell metal foam [10,11]. In previous works, aluminum foam having 20 pores per inch (ppi) was tested. In this paper results for 10-ppi aluminum foam are presented. The foam sample here was made long enough in order to capture anticipated thermal development and exit region. The issue of thermal development in porous media, in general, has been theoretically discussed, e.g. [12-16]. These articles employed various geometries, boundary conditions and assumptions. The case employed in the current experiment was a cylinder filled with aluminum foam subjected to constant wall heat flux. Nield et al. [12] analytically investigated thermal entry length for the case of a circular-tube porous media subjected to constant heat flux assuming local thermal equilibrium. They ignored hydro-dynamic development in the analysis.

The thermal development and thermal entry length are most often ignored in metal foam heat transfer studies. Also ignored is the exit region. In the current study, direct measurements of wall and inlet and outlet temperatures for water flow inside heated, commercial, open-cell aluminum foam are presented. The foam cylinder tested is sufficiently long to ensure that thermal development and full thermal development are clearly captured, as well as the exit region. Flow velocities covered the practically-important non-Darcy flow regimes. The experimentally-obtained data have intrinsic value, and to the best knowledge of the authors, the data set is novel. It can be used for heat exchanger designs using metal foam.

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EXPERIMENT The experimental heat transfer model was an aluminum tube filled and brazed to aluminum foam core, Fig. 1(a). The tube was 32.50 cm long in the flow direction, and with an inside diameter 5.08 cm. The foam was available commercially [17], and had 10 pores per linear inch (ppi). The porosity of the foam was 88.5 %.

For measuring the wall temperature, several holes were drilled. The diameter of each hole was 1 mm and the depth was 4 mm. The holes were placed 10 mm from each other along the length of the tube. In each hole, a type-K thermocouple was inserted, Fig. 1(b). Thermal epoxy then filled the remaining volume of each hole, while ensuring that the bead of the thermocouple was touching the bottom of the hole.

A surface band heater was wrapped around the outer surface of the tube. The heater had resistance ribbons of 1.66 Ohm/m. Two more mica layers gave electrical insulation and steel sheet layers covered the inner elements and provided structure to the heater. The heater had an electrical power of 1,780 Watts at 60 Volts; it was powered by two 40-VDC power supplies connected in series. Thermal grease was placed between the heater surface and the pipe surface to minimize contact resistance, Fig. 1(c). The whole assembly was then covered with layers of ceramic fiber paper, with a thermal conductivity of 0.058 W/mK in order to insulate the test section from ambient air, Fig. 1(d).

Figure 1. Experimental test section: (a) aluminum tube with foam, (b) thermocouple inserted into wall, (c) heater wrapped around pipe and (d) test section insulated The complete experimental set up is shown in Fig. 2. The test section was connected to two 50.8-mm-diameter 200-mm-long Polyethylene tubes using specially-designed flanges. Temperature probes were inserted in these tubes to measure the inlet and outlet temperatures of the water. Each probe had five thermocouples spanning the cross section in order to obtain a good average value for the inlet and outlet temperatures. The outlets of the Polyethylene tubes were connected to stainless steel pipes 32 mm in diameter and 110 cm in length. A hose and a valve were used to connect the outlet of one steel pipe to a 50-liter tank for collecting water at the outlet over a known timespan for measuring mass flow rates. Higher flow rates were measured using a flow meter.

An elevated (3.5 m above ground) tank with a network of hoses and valves guaranteed a constant water height of 33.2 cm in the tank at all times. This tank supplied heavily-filtered water to the test section. For a given valve setting, the flow rate provided by the tank was practically constant (less than 4 % variation).

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Figure 2. Schematic of the complete set-up: 1. Filtered water inlet, 2. Flow meter, 3. 40 DC power supply, 4. Heater, 5. Test section, 6. Thermocouple wires, 7. Data logger, 8. PC, 9. Polyethylene tubes, 10. Stainless steel tubes, 11. Water outlet, 12. Collecting tank, 13. Mass scale, 14. Second DC power supply.

For a given run, control valves were adjusted and water was allowed to flow through the foam in the test section. A valve at the inlet provided fine control over the mass flow rate. Heat was supplied by the heater to produce and maintain two heat fluxes of 14,998 W/m2 and 26,347 W/m2. These fluxes are based on the inside surface area of the pipe, and they exclude an estimated heat loss of about 5%. For any given run, it took about 30 minutes for steady-state conditions to be reached. The free leads of all thermocouples were connected to the data logger, which sent the steady-state temperature readings to a computer. Water exiting the test section was captured in the collecting tank over a known period of time. Knowing the time and the mass of the collected water, the mass flow rate and the average flow velocity were determined. Alternatively, the flow meter was used to measure higher flow rates.

Uncertainty Analysis The uncertainties in length and diameter of the foam were 0.18% and 1.0%, respectively. These were based on errors associated with the accuracy of their measuring devices. As for the temperature, the error in the reading of a thermocouple was 0.4%, as given by the manufacturer.

As an example of uncertainty in derived quantities, the average flow velocity, U was obtained by dividing the mass flow rate, m by density ρ and the cross-sectional area of the test section A. Hence the percent uncertainty in the average velocity δU/U is given by Figliola and Beasley [18]:

, (1)

which results in 0.37%. The uncertainty in the mass flow rate had been obtained in the same manner, which resulted in identical values as those obtained for the mass flow rate due to the very small uncertainties in the density and the cross-sectional area.

The uncertainty in μ was estimated as 1×10-7 N.s/m2, taken as the accuracy of the reported values in property tables. This value is small enough to cause negligible impact on the overall uncertainty in Reynolds number, therefore it was ignored. Using similar calculation as shown by Eq. (1), the uncertainty in Reynolds number was obtained as 2.69%.

The effective thermal conductivity of the solid aluminum ligaments of the foam was obtained from an analytical one-dimensional model given by Calmidi and Mahajan [19]. The fluid phase effective thermal conductivity was obtained as 0.58 W/m.K. The uncertainty in the effective conductivity was conservatively assumed to be 10%. The uncertainty in the heat flux was assumed to be 10%. The uncertainty in Nusselt number was obtained as ±14.12%.

RESULTS The transitional and Forchheimer regimes have been identified previously by flow experiments using the same set-up and water at room temperature [20]. According to preliminary experiments, the heat at the wall caused little changes in the thermophysical properties of the water, and did not affect flow-regime demarcations.

11

1213

1

7

8

4

3

6

5 910

214

183

Wall Temperature The wall temperature along the test section is shown in Fig. 2. Generally, it increases in the flow direction. The wall temperature is lower for higher flow velocities indicating better heat transfer rates at higher velocities. The wall temperature is also higher for higher heat fluxes at the same velocity. There are three sudden increases in the wall temperature; the most obvious among them occurs at about 142 mm from the entrance, while the other two occur at 212 and 292 mm from the entrance approximately. These jumps are due to manufacturing issues of the foam. The foam was produces in short cylinders and was then placed in the aluminium tube and brazed. Because of some mismatches in the internal cell structure, local non-homogeneities resulted which influenced the flow filed and hence the heat transfer at these locations.

Figure 2. Variation of the wall temperature with the distance from the entrance. For all velocities and heat fluxes, the slope of wall temperature initially changes and then becomes constant. This is not easy to see due to some scatter ranging up to 7.8% and more importantly due to the presence of the three jumps mentioned above. The constant slope of the wall temperature is considered a sign of thermally fully-developed heat transfer [21, 22]:

d

dconstant, (2)

Another observation is that the wall temperature drops suddenly close to the exit of the test section for call cases. This is most likely due to the presence of an exit region which begins at about 293 mm from the entrance. The length of the exit region is then 32 mm. Local Nusselt Number

The bulk fluid temperature is defined by

1

d , (3)

where the flow velocity U is averaged over the cross-sectional area. The actual bulk temperature of water along the tube is extremely difficult to obtain experimentally. A reasonable estimation of the bulk temperature can be obtained as follows. It is well established from analysis, e.g. [21], that for thermally fully-developed conditions that d

ddd

constant, (4)

In other words, the slope of the bulk temperature is the same as that of the wall temperature for each flow velocity, which is readily available form plots of the wall temperature (Fig. 3). Having this slope and the average outlet temperature is sufficient to obtain a straight line representing the variation of the bulk temperature in the fully-developed region. This carries a little error

20

30

40

50

60

0 100 200 300 400

Wal

l tem

pera

ture

(°C)

Distance from entrance (mm)

0.0160 m/s - Forchheimer - Low power0.0213 m/s - Forchheimer - Low power0.0313 m/s - Forchheimer - Low power0.0426 m/s - Forchheimer - High power0.0206 m/s - Forchheimer - High power0.0318 m/s - Forchheimer - High power0.0414 m/s - Forchheimer - High power

184

due to the existence of the exit region. Nonetheless, the use of bulk temperature estimated in this manner is better for estimating Nu as oppose to using a fixed temperature, i.e., the inlet temperature, which is constant and does not capture any thermal physics inside the foam. Similar approach is applied to estimate the bulk temperature in the entry region.

The local Nusselt number is defined as

Nu

", (5)

where the effective thermal conductivity of the fluid, kfe was obtained from a model given in Calmidi and Mahajan [19]. The Nusselt number obtained by Eq. (5) is shown in Fig. 4. Nu is seen to be a function of velocity in the thermal entry region, and to a lesser degree, after the end of the developing region. In the fully-developed region, the Nusselt number is seen to be practically independent or very weakly dependent of the axial distance, especially if one ignores the sudden decreases (which correspond to the three jumps in the wall temperature). The independence of the Nu from the distance (at about four diameters from the entrance) is a confirmation of thermal full development condition observed in the behavior of the wall temperature.

Figure 4. Nusselt number as a function of axial distance. CONCLUSION For water flow in a cylinder internally brazed to metal foam and subjected to constant heat flux, the wall temperature and average inlet and outlet temperatures were measured. Flow rates were in the Forchheimer regime. A thermal entrance region and a thermally fully-developed region were identified, as well as an exit region. The thermal entry length was seen to be significant. As such, the thermal entry region in metal foam should not be causally ignored, nor should the exit region. ACKNOWLEDGMENT This work was supported by the Scientific & Technological Research Council of Turkey (TUBİTAK) under program 1002: 214M267, for which the authors are very thankful.

0

500

1000

1500

0,00 2,00 4,00 6,00 8,00

Nuss

elt n

umbe

r

z/D

0.0160 m/s - Forchheimer - Low power0.0213 m/s - Forchheimer - Low power0.0313 m/s - Forchheimer - Low power0.0426 m/s - Forchheimer - High power0.0206 m/s - Forchheimer - High power0.0318 m/s - Forchheimer - High power0.0414 m/s - Forchheimer - High power

185

REFERENCES 1. Dukhan, N., Editor: Metal Foam: Fundamentals and Applications. DESTech, Lancaster, PA 2013. 2. V. V. Calmidi, R. L. Mahajan, Forced convection in high porosity metal foams, J. Heat Transfer, 122 (2000) 557-565. 3. J. J. Hwang, G. J. Hwang, R. H. Yeh, C. H. Chao, Measurement of the interstitial convection heat transfer and frictional

drag for flow across metal foam, J. Heat Transfer, 124 (2002) 120-129. 4. A. Bhattacharya, V.V. Calmidi, R. L. Mahajan, Thermophysical properties of high porosity metal foams, Int. J. Heat Mass

Tran., 45 (2002) 1017-1031 5. I. Kurbas, N. Celik, Experimental investigation on forced and mixed convection heat transfer in a foam-filled horizontal

rectangular channel, Int. J. Heat Mass Tran., 52, (2009) 1313-1325 6. N. Dukhan, M. A. Al-Rammahi and A. S. Suleiman, "Fluid Temperature Measurements inside Metal Foam and

Comparison to Brinkman-Darcy Flow Convection Analysis," International Journal of Heat and Mass, Vol. 67, December 2013, pp. 877-884.

7. K. Boomsma, D. Poulikakos, F. Zwick, Metal foams as compact high performance heat exchangers, Mechanics of Materials, 35 (2003) 1161-176.

8. G. Hetsroni, M. Gurevich, R. Rozenblit, Metal foam heat sink for transmission window, Int. J. Heat Mass Tran., 48 (2005) 3793-3803

9. S.Y. Kim, J.W. Paek, B.H. Kang, Flow and heat transfer correlations for porous fin in a plate-fin heat exchanger, J. Heat Transfer, 122 (2000) 572-578

10. Ö. Bağcı, N. Dukhan and L. A. Kavurmacıoğlu, “Forced-Convection Measurements in the Fully-Developed and Exit Regions of Open-Cell Metal Foam,” Transport in Porous Media, Vol. 109, No. 2, 2015, pp. 513-526.

11. N. Dukhan, Ö. Bağcı, M. Özdemir, “Thermal Development in Open-Cell Metal Foam: An Experiment with Constant Wall Heat Flux,” Int. Journal of Heat Mass Transfer, Vol. 85, 2015, pp. 852-859

12. D.A. Nield, A.V. Kuznetsov and M. Xiong, Thermally developing forced convection in a porous medium: Parallel-plate channel or circular tube with walls at constant heat flux, J. Porous Media, 6(3), 203-212 (2003).

13. Nield, D.A., Kuznetsov, A.V., Xiong, M.: Thermally developing forced convection in a porous medium: parallel plate channel or circular tube with isothermal walls. J. Porous Media 7(1), 19-27 (2004)

14. Kuznetsov, A.V., Nield, D.A., Xiong, M.: Thermally developing forced convection in a porous medium: circular duct with walls at constant temperature, with longitudinal conduction and viscous dissipation effects. Trans. Porous Media 53, 331-345 (2003)

15. Nield, D.A., Kuznetsov, A.V., Xiong, M.: Effect of local thermal non-equilibrium on thermally developed forced convection in a porous medium. Int. J. Heat Mass Trans. 45, 4949-4955 (2002)).

16. N. Dukhan, “Developing Non-Thermal-Equilibrium Convection in Porous Media with Negligible Fluid Conduction,” ASME Journal of Heat Transfer, Vol. 131, No. 1, January 2009, pp. 014501-1--014501-3.

17. ERG Materials and Aerospace: http://www.ergaerospace.com, accessed October, 2014. 18. R. Figliola, D. Beasley, Theory and Design for Mechanical Measurements, (John Wiley and Sons, New York, NY, 2000),

pp. 149-163. 19. V. V. Calmidi, R. L. Mahajan, The effective thermal conductivity of high porosity fibrous metal foams, J. Heat Transfer,

121 (1999) 466-471. 20. A. Arbak, Ö. Bağcı and N. Dukhan, “Flow Regimes in Commercial Metal Foam Having 10 Pores Per Inch,” International

Conference on Energy Systems, 23-25 December 2015, Istanbul, Turkey. 21. Z. G. Qu, H. J. Xu, W. Q. Tao, Fully developed forced convective heat transfer in an annulus partially filled with metallic

foams: An analytical solution, Int. J. of Heat Mass Tran., 55 (2012), 7508-7519. 22. H. Xu, L. Gong, S. Huang, M. Xu, Int. J. Heat Mass Trans., 76, 357-265 , (2014)

186

[Abstract:0057][Energy Efficient Buildings] IMPORTANCE OF CLIMATE FOR DESIGNING OF A ZERO ENERGY BUILDING

Ahmet ARISOY1, Burcu SAĞLAM1, Burhan YORUK1,

1Istanbul Technical University, Istanbul Corresponding email: arisoyah@itu.edu.tr

SUMMARY

It is known that the EPBD recast has introduced in Europe the near-zero energy building concept based on the “cost-optimal level" criterion, that is the energy performance level leading to the lowest cost during the estimated economic lifecycle. Following the Directive, the Guidelines for establishing a comparative methodology framework for calculating buildings and building elements have been issued.

In these guidelines the definition of energy efficiency measures and packages are strictly related to the climate, considering both temperature and humidity. Therefore, for Mediterranean region specific solutions are certainly required that are significantly different with respect to the ones typical of cold climate.

In this case study a sample building and its HVAC system has been designed as nearly zero energy building “nZEB”. However this design has been developed for two different climates and the differences have been analyzed. The resulting building outer skin and the HVAC systems are quite different from each other.

INTRODUCTION

Energy consumption in residential and commercial buildings corresponds 40% of the total energy consumption in 28 EU countries [1]. So, reduction of this value is an important goal for European Union to reduce its energy dependency. Energy Performance Buildings Directive (EPBD) is the most important action that has taken against this issue. According to this directive, by the end of 2018 every public buildings have to be compatible with nearly zero energy standards and by the end of 2020 all new buildings have to be nearly zero energy [2]. The definition of energy efficiency measures and packages are strictly related to the climate, considering both temperature and humidity [3,4]. In particular, Mediterranean climate is characterized by a dominant cooling demand that requires an analysis of the dynamic behaviour of the building systems.

Behaviour of outer walls play an important role on heat gains and heat losses of the building. Thermal insulation of walls is known as an important measure to reduce static heat loss of buildings for cold and mild climates. However increasing thermal insulation thickness plays a reverse effect on heat gains in dynamically changing hot climates [5]. Thermal solar energy as the source of heat is valuable in cold climates, however PV panels are more effective for nZEB solutions in hot climates. Shading elements and air source cooling equipment are advantageous in hot climates. Shortly the definition of energy efficiency measures and packages are strictly related to the climate, considering both temperature and humidity. Therefore, for Mediterranean region specific solutions are certainly required that are significantly different with respect to the ones typical of cold climate.

There are four groups of solutions to design an energy efficient building [6,7,8]. Depending on the conditions, different combinations of solutions can be selected. These solution groups are: 1. Solutions related with the building outer skin 2. High efficiency HVAC equipment and system selections 3. Heat recovery, free cooling and natural ventilation and lightning. 4. Employing renewable energy sources.

In this case study a sample building and its HVAC system has been designed as nearly zero energy building “nZEB”. However this design has been developed for two different climates independently and the differences have been analyzed. The resulting building outer skin and the selected HVAC systems are quite different from each other.

Berlin has been selected to represent the cold climate and İzmir has been selected to represent the Mediterranean climate. Annual primary energy consumption has been limited to 60 kWh/m2a value as definition of the “nZEB”. The numerical analysis of the thermal system has been carried out by “EnergyPlus” and “OpenStudio”. Renewable energy source of solar PV panels have been designed by using “PVsyst” software. By the help of simulation tools, building outer skin has been optimized for two different climate and the most suitable system has been selected for each climate. Lastly necessary renewable energy systems have been designed.

187

METHODOLOGY

A standard building (compatible with TS 825 [9]. Turkish standard for thermal insulation requirements for buildings) has been considered as the reference case (Fig.1) in this study. İzmir has been selected as the representative of the Mediterranean climate and Berlin has been selected as the representative of the cold climate.

Figure 1. 3D model and dimensions of sample building.

Building is considered as 2 story residential house with a total of 512 m2 floor area. Windows are PVC framed and have a typical dimension of (1.60 x 1.80 m). Total window area to external wall ratio is 24%. All dimensions are modeled as symmetrical to eliminate the dependency of orientation. Each floors have 4 external zones without an inner zone.

Density of 5 cm thick external walls is 800 kg/m3. Thermal mass of walls is an important parameter effecting thermal performance of the building skin. Thermophysical properties of materials are obtained from TS 825 [9]. Walls and windows U values remain within the limits specified in this standard. Overall heat transfer coefficient of outer walls is calculated as 0.67 W/m2K for reference case.

Roof, ceiling and floor heat gains/losses have been taken into consideration. Internal loads are same for these two different applications. 16 People have been assumed and the lighting and equipment generated internal heat loads are compatible with ASHRAE standards [10]. Lightning load is 4.5 W/m2 and equipment load is 2.98 W/m2. Same schedules have been applied for two cases. Domestic hot water energy consumptions are also considered in the calculations. Mechanical ventilation system provides 0.0236 m3/s fresh air per person and additional 0.0003 m3/s fresh air per m2 of the floor area.

Heating set point temperature is 21oC during winter day time and 15.6oC during winter night time. Cooling set point temperature is 24oC during summer day time and 29.4oC during summer night time. PASSIVE MEASURES ON BUILDING SKIN Only walls and windows are considered as passive measures in this case study. These are the most important elements of the building that are effecting energy consumption [7]. Optimization of these elements need to be evaluated. Walls Wall materials, wall thickness, thermal conductivity and the mass of the wall are all effecting parameters. Heat gain or loss through the wall should be optimized depending on the climate. In cold climates increasing thermal insulation thicknesses is the most appropriate solution in terms of nZEB targets. However this is not true for Mediterranean climate.

Mediterranean climate can be considered as cooling load weighted climate. Daily temperature changes are important during summer and intermediate seasons. Heat gain during day may turn out to be a heat loss during night. Thermal mass play an important role in this circumstances. In Fig.2 year round cooling and heating loads from only outer skin of the sample building for two different wall constructions in İzmir are seen. Total thermal conductivity of the walls are same but second wall has more thermal mass comparing the first one.

188

Figure 2. Effect of thermal mass on energy demands in İzmir. S-1 wall is constructed with foam concrete; S-2 wall is

constructed with heavy weight concrete and thermal insulation.

Heating load is reduced 4.4% and the cooling load is reduced 2.3% and total yearly energy demand is reduced 2.8% for Case-2 comparing Case-1. However this thermal mass is not effective in such a way in cold climates. Because of this fact a heavier wall construction has been selected for İzmir and considering the cost, light construction has been preferred for Berlin.

Thermal insulation thickness Thermal insulation of the walls is the key element reducing both heating and cooling loads and designing an nZEB. Insulation materials have been investigated first and considering cost, thermal conductivity and being environmental friendly EPS has been selected as most suitable insulation material for both climate.

Again only taking into consideration outer skin of the building, yearly cooling and heating loads and the total energy demands have been calculated with changing insulation thickness for both cities. İzmir is given in Fig.3 and Berlin is given in Fig.4.

Figure 3. Yearly energy demand of building from outer skin in Mediterranean region (İzmir) with changing thermal insulation

thickness.

0

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25.000

30.000

35.000

Gazbeton (Örnek-1) Betonarme+Yalıtım (Örnek-2)

Ener

gy d

eman

d [k

Wh]

Isıtma Enerjisi Soğutma EnerjisiCooling Energy

Foam Concrete Samp-1 Concrete+insulation Samp-2

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0 25 50 75 100 125 150 175 200 225 250 275 300

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Wh]

Insulation thickness [cm]

Isıtma Soğutma ToplamHeating Cooling Total

Heating

189

Figure 4. Yearly energy demand of building from outer skin in cold region (Berlin) with changing thermal insulation thickness.

Cost effectiveness is another important parameter selecting insulation thickness. Considering above simulation results and the previous studies, thermal insulation thicknesses on outer walls of the building have been defined as 8 cm for İzmir and 15 cm for Berlin. As the result, selected wall constructions for two cities are given in Table 1 and 2.

Table 1. Wall construction for İzmir

Material Thickness Density Conductivity [m] [kg/m3] [W/mK] Exterior plaster 0,02 900 0,35 EPS insulation 0,08 30 0,035 Concrete 0,20 800 0,24 Interior plaster 0,02 1800 0,87

Table 2. Wall construction for Berlin Material Thickness Density Conductivity

[m] [kg/m3] [W/mK] Exterior plaster 0,02 900 0,35 EPS insulation 0,15 30 0,035 Foam concrete 0,20 400 0,13 Interior plaster 0,02 1800 0,87

Insulation of roof has also been studied and similar conclusions have been reached. The thermal insulation thicknesses on roof of the building have been defined as 10 cm for İzmir and 17 cm for Berlin.

Windows The second very effective element of the outer skin is window. Both heat gain and loss from windows has a big share in the total building load. Area of windows, their construction and the shading elements play important roles on window heat loads.

Optimization of window to wall area ratio is a very important task designing the building. This is related with both energy consumption and also illumination of the building. Energy consumption increases linearly with this ratio. In Turkish standard (TS 825) for this ratio under 12% value it is not required any thermal insulation. This ratio has been investigated for both cities parametrically and it has been ended up with minimum window/wall ratio value of 12% for Berlin and higher value of 24% for İzmir.

Direction of windows is another parameter. South facing windows for cold climates and North facing windows for hot climates are preferred.

Type of windows Type of glass, number of glazing, type of frame are effecting parameters. On the other hand required specifications are different for winter and summer conditions. Four different type of fenestration performances have been investigated for both cities. In Fig. 5 only results of İzmir is given. For this city, coated double glazing window reaches the best performance. Also considering cost effectiveness, selected windows and their specifications are given in Table 3 and 4 for each city.

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0 25 50 75 100 125 150 175 200 225 250 275 300

Annu

al E

nerg

y Co

ns. [

kWh]

Insulation thickness [cm]

Isıtma Soğutma ToplamHeating Cooling Total

190

Table 3. Selected window for İzmir Table 4. Selected window for Berlin

Double glazing clear glass window specifications

Dimension [mm] 6-13-6 Space gas Argon Solar absorbance 0,283 Light transmittance 0,408 Coating present U value [W/m2-K] 1,333

Shading elements Shading elements, especially external elements are usually very effective in reducing the heat gains during summer time. There are different type of shading elements developed for this purpose. Some examples of them are given in Fig.6. The shading element used in this study is given in Fig.7.

Figure 5. Yearly energy demand of the sample building from outer skin in İzmir with different type of fenestrations.

Figure 6. Different external shading elements Figure 7. External shading element used in this study.

ACTIVE MEASURES Next step toward the nZEB design is selecting high efficiency and suitable HVAC equipment and systems and using heat recovery measures as much as possible. Meanwhile if it is applicable, free cooling and natural ventilations are powerful tools reducing primary energy consumption.

Efficiencies of thermal systems are very much depend on temperatures of air, water and ground as heat source or sink. Due to this fact certain air conditioning systems are more suitable for a specific application and climate than the others. In this case study only two systems have been compared to each other for year round air conditioning of considered building in two different cities İzmir and Berlin. These systems are Fan-coil system and air source VRV system. Both system also provides mechanical

0

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Tek Camlı Üç Camlı Dört Camlı Film Kaplı,Çift CamlıAn

nual

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rgy

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umpt

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[kW

h]

Isıtma SoğutmaHeating Cooling

Single glazed

Double glazed

Four glazed Double glazed+coated

Triple glazing clear glass window specifications

Dimension [mm] 3-13-3-13-3 Space gas Argon Solar absorbance 0,579 Light transmittance 0,698 Coating non U value [W/m2-K] 1,062

Win

dow

Hei

ght Depth

191

ventilation by integrated fresh air units. Both system has been designed for the sample building in İzmir and Berlin. With the help of the simulation tool EnergyPlus, yearly energy consumptions have been predicted. Ventilation rate, domestic hot water demand, lightning load and equipment load are same for all the cases. Building design has been optimized for each city and outer skin of the building is different in each case and suitable to the climate as mentioned above. Energy consumption of fan-coil system in İzmir is given in Table 5.

Table 5. Energy consumption of fan-coil system in İzmir.

System Heating [kWhheat]

Cooling [kWhelec]

Fan [kWhelec]

Pump [kWhelec]

Fan-coil 1.312 9.470 866 1.830 Other energy consumptions should be added on the fan-coil system to get the total energy demand. Conversion factors for electricity and natural gas are 3.167 and 1.084 respectively. Total annual energy demand is given in Table 6. Energy consumption of VRV system in İzmir is given in Table 7.

Table 6. Total annual energy demand in İzmir with fan-coil system.

System Total heat demand [kWhheat/year]

Total electricty demand [kWhelec/year]

Building primary energy demand [kWhheatı/yıl]

Build. specific prim. energy demand [kWhheat/m2.a]

Fan coil 2.428 14.505 48.569 94,86

Table 7. Energy consumption of VRV system in İzmir.

System Heating [kWhheat]

Cooling [kWhelec]

Fan [kWhelec]

Pump [kWhelec]

Fan-coil 4.882 1.139 762 0

Total annual energy demand is given in Table 8 for VRV system. Domestic hot water has been produced by using electric heaters in this alternative.

Table 8. Total annual energy demand in İzmir with VRV system.

System Total heat demand [kWhheat/year]

Total electricty demand

[kWhelec/year]

Building primary energy demand

[kWhheat/yıl]

Build. Ann. Spec. Prim. Ener. demand

[kWhheat/m2.a]

VRF 5.998 4.240 32.424 63,33

Comparison of energy demands of fan-coil system and VRV system in İzmir is given in Fig.9. In according to the building energy simulation, VRV system is more suitable for İzmir or Mediterranean climate. Without using any renewable energy source in site, specific primary energy demand can be reduced to the value of 63.33 kWh/m2a. The target value is 60 kWh/m2a. So it is needed to supply small amount of renewable energy for this case. This renewable source can be a power generating PV panel.

0

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6000

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Isıtma Soğutma Fan Pompa

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al e

nerg

y dem

and

[kW

h]

İzmir - System components

VRF Fan coil

Heating Cooling Pump

192

Figure 9. Comparison of energy demands of fan-coil system and VRV system in İzmir

Energy consumption of fan-coil system in Berlin is given in Table 9.

Table 9. Energy consumption of fan-coil system in Berlin.

System Heating [kWhheat]

Cooling [kWhelec]

Fan [kWhelec]

Pump [kWhelec]

Fan-coil 14.766 5.890 4.321 1.855 Total annual energy demand in this case is given in Table 10. In this case a condensing type natural gas boiler has been used for heating and domestic hot water production purposes.

Table 10. Total annual energy demand in Berlin with fan-coil and condensing boiler.

System Total heat demand [kWhheat/year]

Total electricty demand

[kWhelec/year]

Building primary energy demand [kWhheatı/yıl]

Build. specific pri. energy demand [kWhheat/m2.a]

Fan-coil 12.550 13.707 57.014 111,36 Table 11. Energy consumption of VRV system in Berlin.

System Heating [kWhheat]

Cooling [kWhelec]

Fan [kWhelec]

Pump [kWhelec]

Fan-coil 17.483 319 758 0 Energy consumption of VRV system in Berlin is given in Table 11. Total annual energy demand is given in Table 12 for VRV system. Domestic hot water has been produced by using electric heaters in this alternative again.

Table 12. Total annual energy demand in Berlin with VRV system.

System Total heat demand [kWhheat/year]

Total electricty demand

[kWhelec/year]

Building primary energy demand

[kWhheat/yıl]

Buil. annual sp. Pri. energy demand [kWhheat/m2.a]

VRF 18.599 3.416 69.722 136,17

Comparison of energy demands of fan-coil system and VRV system in Berlin is given in Fig.10. In according to the building energy simulation, Fan-coil system is more suitable for Berlin or cold climate. Without using any renewable energy source in site, specific primary energy demand can be reduced to the value of 111.36 kWh/m2a. The target value is 60 kWh/m2a and it is needed to supply renewable energy. It could be thermal solar energy.

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and

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Berlin - System components

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Heating Cooling Pump

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Figure 10. Comparison of energy demands of fan-coil system and VRV system in Berlin

It is needed to reduce heating loads further in according to this analysis results and it has been decided to increase thermal insulation thicknesses on the walls and the roof to the value of 25 cm. By this way with the help of solar energy supply, specific energy demand value has been reduced below the target limit value of 60 kWh/m2a. RENEWABLE ENERGY PRODUCTION IN SITE One of the main elements reaching nZEB target is using renewable energy [11]. Without using renewable energy it is not possible to reach the ZEB target. Solar energy has been considered in this case study. Actually thermal solar panels are preferred for cold climate solutions. This is a cost effective solution for such climates. However solar PV panels have been selected as renewable source in this case study for both cities for comparison purposes. Only the roof of the building has been used to place the panels. Considering available space, design of the system has been carried out by using the special computer program PVsyst. The total roof area is 256 m2. Optimum angle for İzmir is 33° and the solar production for this angle is 1650 kWh/a. Same values for Berlin are 40° and 1114 kWh/a respectively. A certain type of PV panel system and invertor has been selected from market and the specification of this panel entered to the program. Number of panels can be placed on the roof are 75 for İzmir and 80 For Berlin. Dimensions of the panels are 1.675x1.00 m2. Produced energy monthly and yearly has been calculated and tabulated. Total produced energy in Berlin is 23.624 MWh and 32.725 MWh in İzmir. When the different losses in the PV system considered, net total electrical energy can be produced from the panels on the roof of the building are 46.14 kWh/m2a for Berlin and 64.03 kWh/m2a for İzmir. Total costs of solar system are 60,460 TL in Berlin and 54,490 TL in İzmir. RESULTS Target primary energy consumption value for the nZEB design was (60 kWh/m2a). Reached value after all the measures in Berlin is 59.57 kWh/m2a and the calculated cost of the systems is 310,940 TL.

Applying all the measures end up with almost a ZEB which means no external primary energy supply to the building in İzmir. This costs 322,950 TL. Considering to compromise the cost and primary energy consumption, this cost can be reduced by reducing the number of solar PV panels. nZEB target can be reached in İzmir if only 4 PV panels to be installed and the cost of the system drops to the value of 290,000 TL.

CONCLUSIONS

Same sample building has been designed as the nZEB for two different cities to see the effect of climate on design strategies. Berlin selected as representative of cold climates and İzmir selected as representative of Mediterranean climate. It has been ended up with two totally different building and its HVAC system. This proves that when considering Mediterranean region, different strategies and methodologies are needed to develop comparing the North Europe where most of the design strategies present in literature today.

Design has been develop in four steps, as building skin, HVAC system, heat recovery and renewables. Applying different measures energy consumption of the building has been reduced and the rest of the primary energy demand has been met by in site renewable energy production.

Thermal insulation, window areas, window types and shadowing elements have been optimized as passive measures. VRF and Fan-coil systems compared to each other for two cities and most efficient active system has been selected for these cities. Solar PV panels have been designed for two cases and they are integrated to the designed systems.

The target value to define nZEB has been selected as 60 (kWh/m2.a) for this case study. Without renewable energy contribution in site, this specific primary energy demand value has been decreased to the 111.36 (kWh/m2.a) value in Berlin. Installing 80x1.675= 134 m2 PV panels on the roof of the building decreases primary energy demand further to 59.57 (kWh/m2a) value which is under the target value.

Without renewable energy contribution in site, this specific primary energy demand value has been decreased to the 63.33 (kWh/m2.a) value in İzmir. This is almost the target value. Installing 75x1.675= 125 m2 PV panels on the roof of the building decreases primary energy demand to -0.71 (kWh/m2a) value. It means it is reached Zero Energy Building. To fulfill the 60 (kWh/m2.a) target for nZEB, only installation of 4 panels can be satisfactory.

Main elements designing an nZEB in Mediterranean region are small and double glazing coated windows, shading elements and reasonable insulation. VRF systems or air to air heat pumps can be used for air conditioning purposes. Applying heat recovery measures, considering free cooling and natural ventilation are to be considered. Electricity or thermal solar energy can be used to produce domestic hot water. Lastly PV panels seem best solution as in site renewable energy production.

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However in a cold region main elements designing an nZEB are small and triple glazing clear windows and very thick thermal insulation. Condensing boilers or ground source heat pumps and conventional chillers can be used for air conditioning purposes. Heat recovery should be considered. Natural gas or thermal solar energy can be used to produce domestic hot water. Lastly thermal solar panel can be used as in site renewable energy source.

REFERENCES

1. Eurostat. 2014. Energy Balance Sheets 2011 – 2012. Publications Office of the European Union. doi: 10.2785/52802. pp 8/79

2. Directive 2010/31/EU of the European Parliament and of the Council of 19 May 2010 on the energy performance of buildings, Official Journal of the European Union, L.153, 2010.

3. Kurnitski, J., et al, (2011). Cost optimal and nearly zero (nZEB) energy performance calculations for residential buildings with REHVA definition for nZEB national implementation, Energy and Buildings, 43(2011), 3279–3288.

4. Panão O., Rebelo, M. P., Camelo, S. M. L., (2013). How low should be the energy required by a nearly Zero-Energy Building? The load/generation energy balance of Mediterranean housing, Energy and Buildings, 61, 161–171.

5. Rodriguez-Ubinas, E., et al, 2014. Passive design strategies and performance of Net Energy plus Houses. Energy and Buildings, 83, 10-22.

6. Kapsalaki, M., Leal, V., Santamouris, M. (2012). A methodology for economic efficient design of Net Zero Energy Buildings. Energy and Buildings, 55(2012), 765–778.

7. Thalfeldt M., Pikas E., Kurnitski J., Voll H. (2013). Facade design principles for nearly zero energy buildings in a cold climate. Energy and Buildings, 67 (2013), 309–321.

8. Barthelmes, V.M., et al, (2014). The Influence of Energy Targets and Economic Concerns in Design Strategies for a Residential Nearly-Zero Energy Building. Buildings 2014, 4, 937-962.

9. TS825. 2009. Thermal Insulation Requirements for Buildings. Turkish Standards Institution. ICS 91.120.10. 10. ASHRAE 90.1, Standard 90.1-2013 -- Energy Standard for Buildings Except Low-Rise Residential Buildings. 11. Gallo A., et al, (2013). Analysis of net Zero-Energy Building in Spain. Integration of PV, solar domestic hot water and

air-conditioning systems. Energy Procedia, 48 (2014) 828 – 836.

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[Abstract:0059][Indoor Air Quality and Comfort Conditions] INVESTIGATION OF THERMAL COMFORT IN A CLASSROOM IMPROVED WITH HEAT

RECOVERY VENTILATION SYSTEM

İbrahim Atmaca1, Orhan Ekren2, Sait C. Sofuoğlu3, Z. Haktan Karadeniz4, Macit Toksoy5

1Department of Mechanical Engineering, Akdeniz University, Antalya, Turkey 2Solar Energy Institute, Ege University, Izmir, Turkey

3Department of Chemical Engineering, Izmir Institute of Technology, Izmir, Turkey 4Department of Mechanical Engineering, Katip Celebi University, Izmir, Turkey

5Eneko, Izmir, Turkey SUMMARY It is known that Indoor Environmental Quality (IEQ) effects human health. Similar with houses and offices, schools are important for kids’ health and success since they spent most of their time in those buildings. For this reason, Chamber of Mechanical Engineers and Local Education Division of Izmir have started a project namely “Indoor Air Quality (IAQ) Education in Elementary Schools: A case Study”. For the case study, a school has selected and a new ventilation system with heat recovery installed to improve IAQ in that school. In this study, thermal comfort conditions of the sample classroom is evaluated before and after the improvement with heat recovery ventilation system. In the sample classroom, the temperature and relative humidity is measured and evaluated as suggested in ASHRAE Standard 55-2004. In addition, the survey is applied to each class in the school to make a subjective assessment. INTRODUCTION As well as houses and workplaces, it has been shown in several studies that the mechanical ventilation of classrooms positively affects the health of students and teachers and learning and teaching performance. Students in well ventilated classrooms perform 14 - 18% better in standardized tests compared to others [1]. Today, there are many respiratory diseases such as asthma that occur due to breathing dirty air. Also, several health problems such as red and watering eyes, headache, dizziness, fatigue, malaise and lethargy occur due to staying in dirty environments. Adequate ventilation of classrooms decrease the concentration of volatile chemical and biological compounds from different sources, which affect students’ health more than adults, thus decrease the risk of students getting negatively affected by these compounds. This study was part of the Indoor Environment Quality in Elementary Schools Training project carried out by the Izmir Chamber of Mechanical Engineers in collaboration with the Izmir Directorate of National Education. As part of the study, indoor air quality and thermal comfort measurements were performed in a classroom with no mechanical ventilation system. Then, a mechanical ventilation system was installed in the classroom with a heat recovery device and the measurements were repeated to observe the change in internal air quality and thermal comfort. According to the measurements, after the ventilation system average CO2 value in the classroom is 1170 ppm (authorized value in classroom is 1500 ppm) while it was 2658 ppm before the ventilation system. So, average CO2 is decreased about 29% by the ventilation system with heat recovery and these results were previously reported in detail in various scientific fields. In this study, thermal comfort conditions of the sample classroom is evaluated before and after the improvement with heat recovery ventilation system. METHODS The study was conducted in Nihat Gunduz Middle School in Isikkent, Izmir. The two-story school building had 16 classrooms and hosted 25 teachers and about 350 students. Thermal comfort measurements were performed in a pilot classroom located on the ground floor, which could represent the whole school building and was thought to have a high concentration of contaminants. The ventilation flow rate required for the selection of the heat recovery ventilation device was determined based on the British Building Bulletin 101 Standard. Accordingly, 1 unit of heat recovery exchanger, 2 fans for fresh air and exhausting and the device with filters to clean the air were installed in the pilot classroom complete with air channels. The image and the project of the device can be seen in Figure 1. The device could work at an air flow rate of minimum 400 m³/h and maximum 1200 m³/h with the existing channel system, which provided the desired minimum and maximum air flow rate. A counter flow type heat recovery exchanger with aluminum plate was used on the device. The exchanger was certified by Eurovent and had minimum 75% efficiency in accordance with the EN 308 standard. The speed of the fan was controlled with a CO2 sensor connected to return air. A two-stage filter was used to clean the ambient air and protect the device. G4 rough and M5 medium quality filters were attached to both suction inlets of the device. In order to increase the blow temperature and warm the environment during the winter when the outside air temperature is too low, the electric heater was attached on the blow channel. Thanks to the automation system, the device could work based on the CO2 sensor during class hours and at maximum capacity during breaks. Blow and suction ventilation channels were circular and insulated with polyurethane based fireproof acoustic foam.

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Figure 1. The image of the heat recovery ventilation system in the classroom and the project The temperature and relative humidity were measured in the pilot classroom before and after the improvement with heat recovery ventilation device. Classroom measurement system is presented in Figure 2. As indicated in ASHRAE Standard 55 [2], temperature and relative humidity values corresponding to ankle, waist and head levels for a seated person, which is 0.1 m, 0.6m and 1.1 m relatively, were measured and recorded every 10 minutes with Hobo U12 devices. Thermal comfort assessment was conducted in the classroom for ten one-week periods in November 2014 - January 2015 before intervention and for two weeks in February 2015 after intervention with heat recovery device. It was observed that students wore warm enough clothes, therefore assessments were made by accepting the 20 oC – 24 oC temperature range as the comfort range as recommended in ASHRAE Standard 55. For relative humidity assessments, the 30-70% range was considered to be the comfort range as recommended by ISO 7730 [3]. In addition, subjective assessment was made by applying the thermal comfort survey to the pilot class and also all school. 7-point scale Predicted Mean Vote (PMV) index were used in for the survey and results were evaluated statistically. In this scale -3 is cold, -2 is cool, -1 is slightly cool, 0 is neutral, +1 is slightly warm, +2 is warm and +3 is hot. Subjective survey was performed for a total of 14 separate classes from December till January, whereas the survey was applied to 2 classes for 3 days in February after the heat recovery device was placed in the pilot classroom. Days with very low outside temperatures were carefully selected for the application of the survey in order to assess days with the most negative thermal comfort conditions for the school. The same questions were asked in different ways to achieve cross examination and thus obtain the most accurate results. The students were asked about their clothing to see whether unusual answers were due to clothing. Answers given during cross examination and to clothing-related question were compared and inconsistent answers were eliminated [4].

Figure 2. Pilot classroom and measurement system.

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RESULTS Average temperature and relative humidity values according to months are presented in Figure 3 and Figure 4 in order to show the effect of the improvement with heat recovery ventilation device. It is clearly seen that temperature and relative humidity values were in the recommended range in November and December. Detailed data analysis show that even minimum and maximum temperature values remained within recommended standards and thermal comfort conditions were met in these months and it can also be seen that the difference between 0.1 m ankle level value and 1.1 m head level value was not more than 3 oC, which is the recommended value for thermal comfort in relevant standards. In January and February on the other hand, the temperature values were found to be much lower than the indicated thermal comfort limit and thermal comfort conditions could not be met.

Figure 3. The change of interclass mean temperature for October 2014 – February 2015 and comparison with recommended values in ASHRAE Standard 55 – 2004.

Figure 4. The change of interclass mean relative humidity for October 2014 – February 2015 and comparison with recommended values in ASHRAE Standard 55 – 2004. Considering they had similar outside conditions, when we examine temperature measurements for January, during which working conditions were usual, and February, during which the heat recovery device was active, a significant difference is not observed in Figure 3, but it can be seen that the temperature in the classroom in February did not drop during breaks when the device run at maximum load. On the other hand, a significant difference is seen for relative humidity values. Although the mean relative humidity seems to be within recommended comfort range for all months in Figure 4, as can be seen from Figure 5 and Figure 6, the maximum relative humidity was over 70% in January, which is the upper recommended relative humidity value in classrooms, whereas it remained within the recommended range in February, when the heat recovery device was active.

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Figure 5. Maximum, minimum and mean interclass relative humidity values for January.

Figure 6. Maximum, minimum and mean interclass relative humidity values for February. The thermal comfort questionnaire was applied in December, when the outside temperature was 9.1 oC. The average temperature and relative humidity values in the school were found to be 22.39 oC and 41.21% respectively. Under these conditions, the average PMV in the school was +0.18. The distribution of students, who voted for different PMV values, can be seen in Figure 7. As can be concluded from the Figure, the students tended to prefer cool or warm. In classrooms with temperatures below 21 oC, the PMV preference of the students tended to be cool, whereas in classrooms with temperatures above 21 oC, the tendency was towards very good or warm. The measurements performed in the pilot classroom for December showed that the temperature values were between the limits required for thermal comfort, which also support subjective assessments. The questionnaire was repeated in January, when the outside temperature is quite low. The average temperature and relative humidity values in the school were found to be 17.61 oC and 45.86% respectively. Under these conditions, the average PMV in the school was -1.66. The distribution of students, who voted for different PMV values, can be seen in Figure 8. The measurements performed in the pilot classroom for January showed that the temperature and relative humidity values were not between the limits required for thermal comfort, which support subjective assessments since most of the students stated that the classroom was cold.

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Figure 7. PMV preferences of the students in December survey

Figure 8. PMV preferences of the students in January survey The questionnaire was only repeated in the classroom where the heat recovery ventilation system was installed in February when the outside temperature was quite low. The average temperature and relative humidity values were found to be 15.98 oC and 41.67% respectively. Under these conditions, the average PMV was -0.95. The distribution of students, who voted for different PMV values, can be seen in Figure 9. The measurements performed in the pilot classroom for February showed that the temperature and relative humidity values were not between the limits required for thermal comfort, which support subjective assessments since most of the students stated that the classroom was cool, somewhat cool, and cold. The thermal comfort in the classroom was voted to be very good when the classroom temperature was 18 oC and 19 oC and this preference was reflected in Figure 9. Another finding obtained from the questionnaire was related to answers given to cross questions by students who found the classroom to be warm and somewhat warm. A total of 24 students stated that the classroom was warm or somewhat warm and 79.17% of these students replied the following question with “I wish the classroom was warmer” or “I wish the classroom was much warmer”. 31.58% of the students who replied the question in this way replied the question “Do you feel thermally comfortable at the moment?” with “No”. Similar results were obtained in previous questionnaires as well. This result indicates that students feel the environment to be cold and have not digested the thermal comfort PMV scale properly. Nevertheless, 80 out of 251 (31.87%) questionnaires were eliminated in the first application in December, whereas 47 out of 285 (16.49%) questionnaires were eliminated in the application in January and 23 out of 168 (13.69%) questionnaires were eliminated from the last application in February. The decrease in the rate of eliminated questionnaires indicate that students started to give more logical answers to cross questions and assess the thermal environment in a more consistent manner. A chart was created using all questionnaires for the average PMV values according to classroom temperatures and an attempt was made to determine between what temperature values students felt the most comfortable. This chart is given in Figure 10. As can be seen from the chart, in the comfort zone is accepted to be ±0.5 range, the students felt comfortable between about 20 oC and 24 oC and the temperature corresponding to the PMV=0 value was found to be 22 oC (R2=0.664). This range is consistent with thermal comfort ranges specified by ASHRAE Standard 55 and ISO 7730.

Figure 9. PMV preferences of the students in February survey

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Figure 10. PMV change in response to indoor temperature CONCLUSION Results obtained at the end of the thermal comfort assessment for the school are summarized below: 1- While the temperature and relative humidity values for October, November and December in the classroom were within the

range recommended in international standards, it was found that classroom temperatures were below thermal comfort limits during January and February, in which outside temperatures are low.

2- While the mean relative humidity values were within limits for both January and February, it was observed that the maximum relative humidity exceeded 70% in January, which is the upper limit for thermal comfort. The relative humidity was balanced thanks to the heat recovery device in February and the relative humidity values remained within limits.

3- Although it was not possible to determine whether the heat recovery device was effective or not since temperatures remained below the limit required for thermal comfort during January and February, it was seen that the temperature in the classroom in February did not drop during breaks when the device run at maximum load. Although partially, this suggests that the heat recovery device will not adversely affect thermal comfort.

4- The subjective results obtained from questionnaires and the objective results obtained from measurements support each other. The students found the environment to be acceptable in classroom temperatures over 20 oC, while they voted it to be cool or cold below this temperature.

5- The questionnaire results show that the students participated in the study felt comfortable within the thermal comfort limits specified in international standards (20 oC – 24 oC).

6- The decrease in the rate of eliminated questionnaires indicates that students started to give more logical answers to cross questions as the study progressed.

REFERENCES 1. Shaughnessy, R.J., et al. A Preliminary Study on the Association between Ventilation Rates in Classrooms and Student Performance. Indoor Air 16(6): 465-468. 2006. 2. ANSI / ASHRAE Standard 55 – 2004, “Thermal environmental conditions for human occupancy”, 2004. 3. ISO 7730, “Moderate thermal environments – Determination of the PMV and PPD indices and specification of the conditions for thermal comfort” International Organization for Standardization, 1994. 4. Teli, D., Jentsch, M.F., James, P.A.B., Bahaj, A.S., Field study on thermal comfort in a UK primary school”, Proceedings of 7th Windsor Conference: The chancing context of comfort in an unpredictable world, Cumberland Lodgei Windsor, UK, 12 – 15 April 2012, London.

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[Abstract:0064][Heating, Climatization and Air-conditioning Applications in Buildings] MULTIPLE CRITERIA DECISION FOR INTEGRATED BUILDINGS

M. Selçuk ERCAN - Alarko Carrier

1._ Abstract “Integrated building” design is a process that treats the building as a series of interacting systems, rather than individual components that function in isolation.. Until recent years, there had been little or no collaboration among different teams such as HVAC, Lighting, Power, Security, Commissioning, TAB, Building Operations etc., even Architect, Contractor and Owner.. The design team, brought together by the main contractor, was selected based on past working relationships, trust, experience and mainly cost. Under these conditions, without a new approach to decision making process, design teams tend to arrive at traditional solutions that is based on the position and authority of the respective decision maker. It is clear that this approach does not meet the needs of an integrated building. In this paper I will propose a new mathematical approach that is useful where teams of people are working on complex problems which require the proper mix of the knowledge of all different teams with different stakes and even different specializations, terminologies, or perspectives. Rather than calculating or simulating using complex algorithms to reach the correct decision, this approach helps decision makers find a rational solution that best suits their goal on their Multiple criteria Decision Making problem. Microsoft Excel is the only tool we will need. This Multi Criteria problem can be any problem such as system selection, vendor selection, the intelligence level of the building or Fault Detection and Diagnosis algorithms in BMS.. 2._ Introduction In 1982, since the first vending machine connected ever to the internet, a new world began to emerge in which even almost everything can be connected via the Internet. Even connecting different systems within a building is a serious problem alone, it is not difficult to imagine how the integration of a city or even a country would lead to difficulties. Yet, there are many technical problems in the integration inside of building. The resulting 'intelligent building' is far from meeting the real needs of the end user. The way to curb such problems is reflecting user requirements to the designs of all subsystems, the designs to the selection of appropriate devices and finally reflecting all of them to the building itself. Not only today of the building but also its future should also be taken into consideration in the design phase. From the start of discussions of User Requirements, until the last day that building completes its life time numerous meetings are held. These series of meetings, where requirements are collected from all stakeholders including the owner or the contractor can transform thoughts into measurable results by means of a simple tool which is called AHP (Analytic Hierarchy Process). 2._ Why AHP One of the first social scientists to notice the negative effects of information overload was the sociologist Georg Simmel (1858–1918), who hypothesized that the overload of sensations in the modern urban world caused city dwellers to become jaded and interfered with their ability to react to new situations. The social psychologist Stanley Milgram (1933–1984) later used the concept of information overload to explain bystander behavior. Many decisions in the design process of our buildings, involve an understanding of complex interactions between different subsystems, from its structure, its operators to its tenants and environment. No parties have enough knowledge or power to take decisions in all building subsystems but easily can determine the criteria for responsible parties take decisions. What kind of decisions can we use AHP? 2.1._ Different people from different levels of management are required to attend to a decision. General Manager in financial matters, design engineers on technical features, a field technician for installation process may be required to participate to a decision on some equipment purchase, for example an Air Handling Unit selection.. How this decision can be taken independently of the positions of these three people? 2.2. Dealing with more than one subsystem When multiple projects are running concurrently, which is the norm in contracting environments, the intersystem interaction becomes problematic. As a result, we may need to focus on interdependence among subsystem projects in order to maximize the objective of group success as opposed to individual subsystem success, such as BMS, HVAC, Fire etc., 2.3._ Projects with high technical risk Technical risk is based on the known technical limits of the technology that is applied to the project. Every intelligent building project is unique, uses newly developed or unproven components such as new firmware, new software or new hardware because of the pace of the technological development. 2.4._ When deciding on equally suitable alternatives. For example, when choosing an automation system, we may have to evaluate numerous equal bids.

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2.5._ Competing Criteria IT department put forward operating system, electronics engineer insists on networking and controllers, HVAC engineer gives the best score to valves.. Each party may claim their criteria have precedence over others. While different experts have divergent opinions, we need a successful method for developing consensus among different experts. 4._ Technical Evaluation of bids for a building subsystem Suppose we evaluate proposals for building automation system of a project. The parameters we will use in the evaluation are First Cost, Operating Cost, Energy Saving, Maintenance, HVAC Control Algorithms, 3rd party Integration, Reliability, Complexity of Installation and Aesthetics which are composed of 9 parameters. Naturally many more parameters may be recommended in the evaluation of such a system. Once the assessment is completed, insignificant parameters should be eliminated by applying the 80-20 rule. In Analytic Hierarchy Procedure, each individual parameter should be compared with each other. Comparing each parameter with each other allows us to test existence of any inconsistency in comparisons. The following scale should be used in comparison. Questions are answered in verbal terms, evaluator converts them to numbers to constitute the pair wise comparison matrix. Extremely Less Important 1/9 Very Strongly Less Important 1/7 Strongly Less Important 1/5 Moderately Less Important 1/3 Equal Importance 1 Moderately More Important 3 Strongly More Important 5 Very Strongly More Important 7 Extremely More Important 9

Pair wise comparison scales

The above scale shows that “Energy Saving” is 5 times more important than “HVAC Control algorithms”.

Pairwise comparisons Firs

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First Cost 1 5 3 9 3 3 5 7 9 Operating Cost 1/5 1 1 5 3 3 7 3 7 Energy Saving 1/3 1 1 9 5 5 5 9 7 Maintenance 1/9 1/5 1/9 1 1 1 3 3 7 HVAC Control Algorithms 1/3 1/3 1/5 1 1 1 5 9 7

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3rd party Integration 1/3 1/3 1/5 1 1 1 5 9 7 Reliability 1/5 1/7 1/5 1/3 1/5 1/5 1 9 9 Complexity of Installation 1/7 1/3 1/9 1/3 1/9 1/9 1/9 1 5 Aesthetics 1/9 1/7 1/7 1/7 1/7 1/7 1/9 1/5 1

Number of comparisons is n(n-1)/2=9*8/2=36 . As the result of 19 pairwise comparisons , we have the reciprocal priority matrix shown above. 19 comparisons because in order to fill the lower triangular matrix, we use the reciprocal values of the upper triangular matrix.

Normalized Matrix Firs

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First Cost 0.362 0.589 0.503 0.336 0.208 0.208 0.160 0.139 0.153 Operating Cost 0.072 0.118 0.168 0.187 0.208 0.208 0.224 0.060 0.119 Energy Saving 0.121 0.118 0.168 0.336 0.346 0.346 0.160 0.179 0.119 Maintenance 0.040 0.024 0.019 0.037 0.069 0.069 0.096 0.060 0.119 HVAC Control Algorithms 0.121 0.039 0.034 0.037 0.069 0.069 0.160 0.179 0.119 3rd party Integration 0.121 0.039 0.034 0.037 0.069 0.069 0.160 0.179 0.119 Reliability 0.072 0.017 0.034 0.012 0.014 0.014 0.032 0.179 0.153 Complexity of Installation 0.052 0.039 0.019 0.012 0.008 0.008 0.004 0.020 0.085 Aesthetics 0.040 0.017 0.024 0.005 0.010 0.010 0.004 0.004 0.017

Having a comparison matrix, now we would like to compute priority vector, which is the normalized Eigen vector of the matrix. First, we sum each column of the reciprocal matrix. Then we divide each element of the matrix with the sum of its column, we have normalized relative weight. The sum of each column is 1. The new matrix is called “Normalized Matrix” or “Priority Vector”. Aside from the relative weight, we can also check the consistency of our reciprocal comparison matrix. To do that, we need what is called Principal Eigen value. Principal Eigen value is obtained from the summation of products between each element of Eigen vector and the sum of columns of the reciprocal matrix. To determine the priorities for criteria, we simply find the average of the various rows from the matrix of priority vector. 0.362 0.589 0.503 0.336 0.208 0.208 0.160 0.139 0.153

For example average of these numbers is 0.295194 which is the weight of first cost. Rank Criteria Weights 1 First Cost 0.295194 3 Operating Cost 0.151337 2 Energy Saving 0.210185 5 Maintenance 0.059171 6 HVAC Control Algorithms 0.0919 4 3rd party Integration 0.0919 7 Reliability 0.058517 8 Complexity of Installation 0.02729

It is believed that the paper has provided an adequate foundation to understand the core of the method and “consistency analysis” is ignored for avoiding complexity. With this method, the selection of many different systems can be accomplished without risk and errors. Some templates for criteria can be found in many resources.

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For Building Automation System (Wong & Li 2006) Work efficiency

Grade and level of BAS Ability of integration Complied with standard Use of internet protocol Reliability Efficiency (speed) Allow for further upgrade Maintenance factors Remote control and monitoring Life span

Cost effectiveness First cost Life cycle cost Or for an HVAC system, Cost Capital Cost Maintenance Cost Energy Cost Replacement Cost (life time, independence of brand, etc.) Related Building Costs (Mechanical room size, height etc.) Flexibility Operational Layout Adaptability Load Adaptability System Integration Ease Of Maintenance Flexibility Aesthetics Simplicity Environment Internal Air Quality Acoustics Sustainable Development (Leed, Breams etc.) Air Distribution Comfort Result: You must first select the criteria that apply to all the options, then you can accurately measure the suitability of all options for your intended goal. References: 1._Application of the analytic hierarchy process (AHP) in multi-criteria analysis of the selection of intelligent building systems, Johnny K.W. Wong , Heng Li, 2006 2._ Multiple Attribute Decision Making Methods and Applications, Gwo-Hshiung Tzeng, Jih-Jeng Huang, CRC Press, 2011 3._Handbook on Constructing Composite Indicators, OECD, 2012 4._How to Make a Decision: The Analytic Hierarchy Process, Thomas L. Saaty, 1994 5._Evaluating Intelligence In Intelligent Buildings Case Studies In Turkey,Maryam Farzin Moghaddam, ODTÜ,2012

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[Abstract:0065][Heating, Climatization and Air-conditioning Applications in Buildings] COMPUTATIONAL STUDY ON FLOW CHARACTERISTICS OF AN ENERGY EFFICIENT

AIR CURTAIN FOR LARGE BUILDINGS

Murat Çağlar1*, Yasin Şöhret2, Ahmet Gökşin1, T. Hikmet Karakoç3 1Havak Industrial Plants Inc., TR-34050, Istanbul, Turkey.

2Keciborlu Vocational School, Suleyman Demirel University, TR-32700, Isparta, Turkey. 3Faculty of Aeronautics and Astronautics, Anadolu University, TR-26470, Eskişehir, Turkey.

Corresponding e-mail: murat@havak.com SUMMARY Air curtains are essential devices for energy efficient buildings. As a known fact, they are commonly preferred in heating, ventilating and air conditioning applications of especially large buildings such as shopping malls, hospital, commercial buildings and so on. The main goal of air curtain usage is reducing heat transfer among outdoor and indoor environment. By this way, a partial insulation is obtained at the gate of the building. Thus, flow characteristics of an air curtain plays a vital role to gain a good insulation in the application field. The current paper presents flow analysis of an energy efficient air curtain developed by Havak Inc. for large buildings. The results obtained from the present paper can be useful for researchers who are interested in heating, ventilating and air conditioning, indoor air quality, energy efficient buildings topics. INTRODUCTION In last decade, interest of people on the comfortable living gradually increase. Especially technological developments in devices being used in commercial and residential buildings lead to achieve more comfortable living areas. Thermal comfort is one of numerous parameters affecting wellbeing in living areas. Difference between indoor and outdoor temperatures, indoor air humidity and blowing speed of ventilation air are known parameters having influence on the thermal comfort [1-3]. Many former study [4-8] have proven that heat loss from buildings to ambient occurs at walls, doors and windows. Reducing heat losses from walls may be possible with the aid of better insulation techniques compared to past. But windows, doors and gates are the main sources of infiltration. Especially, doors and gates of community life areas such as commercial buildings, schools, malls, and hospitals and so on are commonly used. In this framework, preventing heat loss at entrances and exits of community life areas is required. For this purpose, air curtains are mostly preferred to insulate indoor air from the outdoor air. On the other hand, air curtains play a crucial role on reducing heating, cooling and air conditioning costs in addition to providing thermal comfort implicitly. Operating values of an air curtain are set regarding operating conditions. As mentioned formerly, blowing speed of air is important for achieving the thermal comfort. Operating air speed should be determined with respect to thermal comfort conditions not to discomfort people. Thus, blowing characteristics of different air curtain types have been examined both numerically and experimentally in many former paper. Giraldez et al. [9] present a derived methodology for efficiency and performance prediction of an air curtain in their paper. Analytical method introduced in the paper was developed using results obtained from numerical analyses and gained data from experiments. Additionally, parametric study was included in the paper to reveal interaction among various parameters and efficiency. Cui and Wang [10] discuss numerical evaluation of an air curtain at design step. Authors emphasized necessity of computer aided numerical analyses for a more efficient air curtain design. Herein, numerical analysis methods are suggested to be a strong tool for optimization and efficiency improvement. In another research [11], air curtain applications for a cold storage room at summer and winter seasons are investigated energetically, environmentally and economically. In this manner, energy consumptions, energy losses, emitted carbon dioxide to the atmosphere and cost effectiveness of a high efficient air curtain and alternative solutions were compared. In the end of the study, the high efficient air curtain was found to be the best solution for separating indoor and outdoor airs with respect to energy consumption, costs and being environment friendly. Wang and Zhong [12] evaluated infiltration characteristics of a building entrance equipped with an air curtain. Numerical analyses are conducted for different ambient temperatures, pressures and usage frequency of the entrance parametrically. In Ref. [13], impact of opposing buoyancy force on an air curtain performance. The study was performed experimentally with a scaled model for a heated zone separated from a cold ambient. Moureh and Yaaghene [14] examined the flow characteristics of an air curtain both numerically and experimentally in their study. 1:5 scaled air curtain model was used during the experiments. PIV and LDS techniques were preferred to observe flow patterns through the air curtain. In the end of the study, it was found that external flow velocity has an inversely proportional impact on the efficiency of the air curtain. According to the summarized literature survey, analyzing the flow patterns through the air curtains are significant for efficient design. For this purpose, blowing characteristics of a genuine air curtain, designed and manufactured in Turkey, is numerically investigated in the current paper.

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METHODS AND MATERIALS Geometry At the first step of the study, computer aided design of the air curtain was done. The geometry of the air curtain is compiled of a crossflow type fan, two nozzles surrounding the fan, electrical preheaters and outlet grill. 3D model of the air curtain under investigation is illustrated in Fig. 1.

Fig. 1. 3D model of the examined air curtain Meshing Before computer aided numerical analyses, meshing the model is required. Regarding different flow types and boundary conditions, size of the mesh varies. For this reason, global and local meshing techniques are preferred. However, tetrahedron type meshing inflation are commonly applied. So, more sensitive results can be obtained where the velocity is approximately zero. Numerical Modelling ANSYS Fluent [15] software is used at this step of the study. It is known that, the main purpose of the air blow out from the air curtain is hindering the interaction among indoor and outdoor airs. So that, κ-ε modelling is preferred to understand turbulence mechanism better. Additionally, the pressure-based coupled solver is applicable for most flows, and flow regimes from low speed incompressible flow to high-speed compressible flow. RESULTS AND DISCUSSION

Fig. 2. Flow patterns within the air curtain in the end of primary analysis In the present paper, flow characteristics of a novel energy efficient air curtain is dealt with the aid of numerical methods. As shown in Fig. 1, in the end of the primary flow analysis design faults and consequently occurring turbulent flow zones within the

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air curtain are detected. Then, analyses are pursued to minimize turbulent flow zones and achieve laminar flow at the outlet grill. According to these results, nozzle design and location of the fan are optimized. However, increasing distance between fan and outlet grill leads more homogenous air distribution and more laminar flow at the outlet grill. Also, upper and lower nozzle positions are also revised for better air aspiration and preventing compress in addition to minimize recirculation zones.

Fig. 3. Flow patterns within the air curtain in the end of final analysis In conclusion, a laminar flow distribution is obtained at outlet grill section of the air curtain. On the other hand, rattle is reduced as a result of minimizing turbulent flow zones. Depending on laminar flow zone at the outlet grill and homogenous air distribution within the air curtain, more homogenous temperature distribution is gained after revising the internal design of the air curtain. Computer aided flow analysis has also beneficial effect on compact internal design, visually better and eurhythmic external design ACKNOWLEDGEMENT Authors are grateful to the Scientific and Technological Research Council of Turkey (TUBITAK) for funding this research under contract number of 7131275.

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REFERENCES 1. Nicol, J. F., Humphreys, M. A. 2002. Adaptive thermal comfort and sustainable thermal standards for buildings. Energy

and buildings. Vol. 34(6), pp. 563-572. 2. Yang, L., Yan, H., Lam, J. C. 2014. Thermal comfort and building energy consumption implications–a review. Applied

Energy. Vol. 115, pp. 164-173. 3. Karakoc, T.H., Goksin, A.H. 2011. Nemlendirme Teknigi, Havak Teknik Yayinlari. 4. Asdrubali, F., Baldinelli, G., Bianchi, F. 2012. A quantitative methodology to evaluate thermal bridges in

buildings. Applied Energy. Vol. 97, pp. 365-373. 5. Kosny, J., Desjarlais, A. O. 1994. Influence of architectural details on the overall thermal performance of residential wall

systems. Journal of Building Physics. Vol. 18(1), pp. 53-69. 6. Vatistas, G. H., Chen, D., Chen, T. F., Lin, S. 2007. Prediction of infiltration rates through an automatic door. Applied

thermal engineering. Vol. 27(2), pp. 545-550. 7. Bolattürk, A. 2006. Determination of optimum insulation thickness for building walls with respect to various fuels and

climate zones in Turkey. Applied thermal engineering. Vol. 26(11), pp. 1301-1309. 8. Karakoc, t.H. 2011. Klorifer Tesisati Hesabi: Verimli Sistemler. 9. Giráldez, H., Segarra, C. P., Rodríguez, I., Oliva, A. 2013. Improved semi-analytical method for air curtains

prediction. Energy and buildings. Vol. 66, pp. 258-266. 10. Cui, J., Wang, S. 2004. Application of CFD in evaluation and energy-efficient design of air curtains for horizontal

refrigerated display cases. International Journal of Thermal Sciences. Vol. 43(10), pp. 993-1002. 11. Gil-Lopez, T., Castejon-Navas, J., Galvez-Huerta, M. A., O’Donohoe, P. G. 2014. Energetic, environmental and economic

analysis of climatic separation by means of air curtains in cold storage rooms. Energy and Buildings. Vol. 74, pp. 8-16. 12. Wang, L. L., & Zhong, Z. 2014. An approach to determine infiltration characteristics of building entrance equipped with air

curtains. Energy and Buildings. Vol. 75, pp. 312-320. 13. Frank, D., Linden, P. F. 2015. The effects of an opposing buoyancy force on the performance of an air curtain in the

doorway of a building. Energy and Buildings. Vol. 96, pp. 20-29. 14. Moureh, J., Yataghene, M. 2016. Numerical and experimental study of airflow patterns and global exchanges through an air

curtain subjected to external lateral flow. Experimental Thermal and Fluid Science. Vol. 74, pp. 308-323. 15. http://www.ansys.com/Products/Fluids/ANSYS-Fluent

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[Abstract:0066][Heating, Climatization and Air-conditioning Applications in Buildings] FIRST LAW ANALYSIS OF SINGLE STAGE ABSORPTION HEATING SYSTEM USING

LIBR-H2O H2O-NH3 NASCN-NH3 AND LINO3-NH3

Veysel Ergül1, Yunus Emre Talu1, Hakan Demir1, Şevket Özgür Atayılmaz1 1Yıldız Technical University, Turkey

Corresponding email: veyselergul45@gmail.com SUMMARY In this study, absorption cycle was examined for heating purposes. Commonly natural gas fired boilers were used and its’ efficiency is approximately up to 95%. Heat pump systems’ coefficient can be higher than 100% because of free energy that taken from ambient. Vapor compression heat pumps use electricity as a source of energy. However, generator, absorber and pump instead of compressor have been used in absorption system in contrast with vapor compression system. Circulation pump is the only electrical driven component of system. Energy consumption of pump is neglectable because it is far less than system capacity. Absorption heat pumps use heat as their energy source, and can be driven with a wide variety of heat sources. LiBr-H2O, H2O-NH3, NaSCN-NH3 and LiNO3-NH3 were analyzed as working pairs of absorption heating system from the perspective of first law of thermodynamics with using Engineering Equation Solver (EES) programme. Gas Utilization Efficiencies (GUEs) were compared between solutions.

INTRODUCTION Vapor compression systems have been used in many cooling applications. Recently, global warming has made people to research for absorption system. Absorption refrigeration systems are helpful to reduce fossil fuel consumption since can be driven by waste heat, solar or geothermal energy [4]. Absorption refrigerant cycle can be used as heater. Natural gas fired boilers have been used for domestic hot water supply. Condensing boiler’s efficiency is up to %95. NH3-Water has been commonly used for absorption heating applications. Farshi et al. studied on absorption cooling systems. Found NH3-Water COP values up to 0,54 [3] In this paper single effect absorption system was numerically analyzed by using Engineering Equation Solver (EES) programme for 4 different working pairs in Table 1.1. Wei Wu et al. has a research about absorption heating and cooling technologies and gave LiBr-Water single effect refrigerant COP up to 0,9,[10] but LiBr-Water is not proper for using heating purposes because of water freezes at 0°C. However it analyzed for just 5°C, 10°C and 15°C evaporation temperatures. Farshi et al. found NH3-NaSCN COP up to 0,64. It is 18,5% higher than NH3-Water.[3] However in this paper GUE of NH3-Water was found 11,5% higher than NH3-NasCN for heating. Farshi et al. found NH3-LiNO3 COP up to 0,56. It is approximately equal to NH3-Water.[3] However in this paper GUE of NH3-LiNO3 was found 24,3% higher than NH3-NasCN for heating.

Table 1.1 Working pairs Absorbent Refrigerant

LiBr-Water LiBr Water Water-NH3 Water NH3

NaSCN-NH3 NaSCN NH3 LiNO3-NH3 LiNO3 NH3

METHODS In contrast with vapor compression system; a generator, an absorber and a pump instead of a compressor have been used in absorption system. Circulation pump is the only electrical driven component of system. Energy consumption of pump is neglectable because of it is far less than system capacity. Single effect system has two different working pressures. Circulation pump has used for increasing the pressure and receive the solution from absorber to the generator passing through the regenerator. While the absorber and the evaporator work at low pressure, the generator and the condenser work at high pressure. System working pressures are saturation pressure of refrigerant at evaporator and condenser temperatures, so pressure values determined by refrigerant type and working temperatures of the evaporator and the condenser. High concentrated solution coming from regenerator ‘7’ heated by an external heat source at generator to separate pure refrigerant from absorbent. After separation stage weak solution leaves separator as saturated concentration ’9’ and pure refrigerant leaves in the gas phase ’1’ at the temperature of generator.

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Pure refrigerant rejects heat in the condenser and left as saturated liquid ‘2’. In heating mode, condenser heats domestic heating supply. In cooling mode, condenser rejects heat to heat sink. Refrigerant which reduced pressure at adiabatic expansion valve enters evaporator ’3’. While rapidly evaporate at low working pressure extract free heat from low temperature heat source. Refrigerant leaves evaporator as saturated vapor ‘4’. In absorber, while weak solution which comes from generator by passing through regenerator absorbs refrigerant which comes from evaporator in the vapor phase, rejects heat. In heating mode, it heats domestic heating supply in conjunction with condenser. Solution leaves absorber as saturated concentration and go back to generator by passing through regenerator whereby pump.

Fig2.1 Absorption heating system

Equations which are seen in Table 2.1 created according to laws of conservation of mass and energy.

(1)

(2)

System was created by using EES to calculate GUE values depend on temperature variations of components.

Table 2.1 Thermodynamic equations Condenser Evaporator Absorber Generator

Regenerator Separator

Pump Expansion Valve 1 Expansion Valve 2

GUE was calculated for all combinations of conditions are as below, • 4 given working pairs (LiBr-H2O, H2O-NH3, NaSCN-NH3 AND LiNO3-NH3) • Evaporator temperature -30°C to 15°C with increasing 5°C • Absorber temperature 40°C to 60°C with increasing 5°C • Condenser temperature 40°C to 65°C with increasing 5°C • Generator temperature 120°C to 220°C with increasing 5°C • Regenerator efficiency %60, %70 and %80 GUE values will be compared according to working pairs and other conditions. Most efficient points will be identified. RESULTS

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Variation of GUE which depends on working conditions was calculated for each liquid pairs. GUE variation of LiBr-Water was drawn in Fig 3.1a and Fig 3.1b.

(a) (b) Fig 3.1 (a) Variation of GUE depending on generator temperatures for each evaporator temperatures. (ηreg =0,8 Tabs=55°C Tcon=60°C), (b) Maximum points of GUE at different evaporation temperatures for each regenerator efficiency. ( Tabs=55°C Tcon=60°C) GUE variation of Water-NH3 according to working conditions was calculated and shown in Fig 3.2 to Fig 3.4 also Max GUE points can be seen in these figures.

(a) (b) Fig 3.2 (a) Variation of GUE depending on generator temperatures for each evaporator temperatures. (ηreg =0,8 Tabs=55°C Tcon=60°C), (b) Maximum points of GUE at different evaporation temperatures for each regenerator efficiency. (Tabs=55°C Tcon=60°C)

(a) (b) Fig 3.3 (a)Variation of GUE depending on generator temperatures for each evaporator temperatures. (ηreg =0,8 Tabs=40°C Tcon=45°C), (b) Maximum points of GUE at different evaporation temperatures for each regenerator efficiency. (Tabs=40°C Tcon=45°C)

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Fig 3.4 Comparison of maximum GUE values at different absorption and condensation temperatures for each evaporation temperatures. (ηreg =0,8) GUE variation of NaSCN-NH3 according to working conditions was calculated and drawn in Fig 3.5 to Fig 3.7 also Max GUE points can be seen in these figures.

(a) (b) Fig 3.5 (a) Variation of GUE depending on generator temperatures for each evaporator temperatures. (ηreg =0,8 Tabs=55°C Tcon=60°C), (b) Maximum points of GUE at different evaporation temperatures for each regenerator efficiency. (Tabs=55°C Tcon=60°C)

(a) (b) Fig 3.6 (a) Variation of GUE depending on generator temperatures for each evaporator temperatures. (ηreg =0,8 Tabs=40°C Tcon=45°C) (b) Maximum points of GUE at different evaporation temperatures for each regenerator efficiency. (Tabs=40°C Tcon=45°C)

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Fig 3.7 Comparison of maximum GUE values at different absorption and condensation temperatures for each evaporation temperatures. (ηreg =0,8 ) GUE variation of LiNO3-NH3 according to working conditions calculated and drawn in Fig 3.8 to Fig 3.10 also Max GUE points can be seen in these figures.

(a) (b) Fig 3.8 (a) Variation of GUE depending on generator temperatures for each evaporator temperatures. ( ηreg =0,8 Tabs=55°C Tcon=60°C), (b) Maximum points of GUE at different evaporation temperatures for each regenerator efficiency. (Tabs=55°C Tcon=60 °C)

(a) (b) Fig 3.9 (a) Variation of GUE depending on generator temperatures for each evaporator temperatures. (ηreg =0,8 Tabs=40°C Tcon=45°C), (b) Maximum points of GUE at different evaporation temperatures for each regenerator efficiency. (Tabs=40°C Tcon=45°C)

Absorber (°C) Condenser (°C)

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Fig 3.10 Comparison of maximum GUE values at different absorption and condensation temperatures for each evaporation temperatures. (ηreg =0,8 ) H2O-NH3, NaSCN-NH3 and LiNO3-NH3 were compared based on maximum GUE values in different conditions and were drawn Fig 3.11 and 3.12.

Fig 3.11 Comparison of maximum GUE values at different evaporation temperatures for each mixtures at two different heating temperatures. (ηreg =0,8 Tabs=40°C Tcon=45°C, Tabs=55°C Tcon=60°C)

Absorber (°C)

Condenser (°C)

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Fig 3.12 Comparison of maximum GUE values at different evaporation temperatures for each mixtures at two different heating temperatures. (ηreg =0,6 Tabs=40°C Tcon=45°C, Tabs=55°C Tcon=60°C) In Fig 3.1a and Fig 3.1b LiBr-Water showed. Because of water freezing point of Water, LiBr-Water analyzed for just 5-10 and 15°C evaporator temperatures. As seen LiBr-Water is efficient but it is not proper to use for winter conditions as air source absorption heat pump. Because when domestic space needs heating, air temperature is usually low. Ground source or water source can be use instead of air source. Ground or water source temperature will be higher than air temperature [2] so it can be work with ground or water source. It should be examined deeply. As seen Fig 3.2a, Fig 3.3a, Fig 3.5a, Fig 3.6a, Fig 3.8a, Fig 3.9a evaporation temperature directly affected to GUE. Magnitude of GUE is higher at high evaporation temperatures in comparison with low evaporation temperatures. However it is not possible to estimate GUE values as a function of generator temperatures. In same figures when generator temperature increases, firstly GUE increases up to maximum point and then starts to decrease. As seen Fig 3.2b, Fig 3.3b, Fig 3.5b, Fig 3.6b, Fig 3.8b, Fig 3.9b, GUE directly influenced by regenerator effectiveness. When regenerator is more effective GUE values will be higher. Condenser and absorber temperatures are substantial for supply heat at desired temperature. But as seen Fig 3.4, Fig 3.7, Fig 3.10 increase of these temperatures are reduce GUE. Maximum GUE values in same conditions can be seen at Fig 3.11, Fig 3.12 for each working pair. In addition to this, maximum GUE values was determined for different condenser and absorber temperatures for different intended use of same working pair for two different heating implementation and it was showed as a variation of evaporator temperatures. Fig 3.11 was analyzed for 80%, Fig 3.19 was analyzed for 70%, Fig 3.12 was analyzed for 60% effectiveness of regenerator. CONCLUSION This study clearly shows that LiNO3-NH3 more efficient than other working pairs. For instance, while temperature of evaporator is -20°C, absorber is 55°C, condenser is 60°C and effectiveness of regenerator is 80%, max GUE of NaSCN-NH3 is 1,223 , Water-NH3 is 1,364, LiNO3-NH3 is 1,702 and in these conditions with same order, temperatures of generator which provides maximum GUEs are 165°C, 190°C, 180°C. Water-NH3 was indicated as a ratio of 11,5% more efficient than NaSCN-NH3; also LiNO3-NH3 was determined as 24,7% more efficient than Water-NH3, 39,1% than NaSCN-NH3. Generator temperatures should be analyzed from the perspective of second law of thermodynamics. If naturel gas fired absorption heat pump (GFAHP) by using LiNO3-NH3 utilize for heating instead of naturel gas fired boiler, naturel gas usage will decrease the about amount of 41,2% and also CO2 emissions. According to this clearly seen, when seasonal efficiency examined, operating cost of GFAHP will be lower than boilers’. However, GFAHP’s initial cost will be higher than initial cost of boilers’, so we suggest that a combine of initial and operating costs should be compared. The literature indicates that double effect absorption cooling systems are more efficient than single stage absorption cooling systems. In that case, double effect absorption heat pump should be analyzed to increase GUE. LIST OF SYMBOLS

c Specific heat[kj] h Enthalpy [kj/kg °K]

Flow rate [kg/s] P Pressure [kPa]

Heat transfer rate [W] T Temperature [°C] W Work [W] η Efficiency LIST OF SUBSCRIPT abs Absorber con Condenser eva Evaporator gen Generator pump Circulation pump reg Regenerator LIST OF ABBREVIATIONS COP Co-efficient Of Performance GUE Gas Utilization Efficiency EES Engineering Equation Solver

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REFERENCES 1. Dehua C, First law analysis of a novel double effect air-cooled non-adiabatic ammonia/salt absorption refrigeration cycle , Energy Conversion and Management 98 2015. 2. Ergül V, Talu Y.E., Gönül A., Comparative Economical Analysis of Air and Ground Source Heat Pumps for Different Climatic Regions of Turkey, XI. International HVAC+R Technology Symposıum, Istanbul, 08-10 May, 2014 3. Farshi L., Mosaffa A.H., Infante Ferreira C.A., Rosen M.A. Thermodynamic analysis and comparison of combined ejector–absorption and single effect absorption refrigeration systems, Applied Energy 133 (2014) 335–346. 4. Karamangil M.I, Coskun S. , Kaynakli O., Yamankaradeniz N A simulation study of performance evaluation of single-stage absorption refrigeration system using conventional working fluids and alternatives Renewable and Sustainable Energy Reviews 14 (2010) 1969 1978 5. Koroneos C., Nanaki E., Xydis G. Solar air conditioning systems and their applicability—An exergy approach, Resources, Conservation and Recycling 55 (2010) 74–82 6. Kühn A. Thermally Driven Heat Pumps for Heating and Cooling ISBN 978-3-7983- 2596-8 7. Patek, J., Klomfar J. Simple functions for fast calculations of selected thermodynamic properties of the ammonia-water system. Int. J. Refrig., 18(4). 228-234 2015. 8. Xu Z.Y., Wan R.Z. Experimental verification of the variable effect absorption refrigeration cycle Energy 77 (2014) 703e709 9. Solum C. Double Effect Absorption Refrigeration System With LiBr-H2O Fluid Thermodynamic Figures on the Effects of System Performance 10. Wu W, Wang B, Shi W, Li X. Absorption heating technologies: A review and perspective

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[Abstract:0067][Heating, Climatization and Air-conditioning Applications in Buildings] STATE OF THE ART OF HVAC TECHNOLOGY IN EUROPE AND AMERICA

Bjarne W. Olesen and Ongun Berk Kazanci

International Centre for Indoor Environment and Energy, Department of Civil Engineering, Technical University of Denmark, Kgs. Lyngby, Denmark

Corresponding e-mail: bwo@byg.dtu.dk SUMMARY The main purpose of heating, cooling and ventilation systems is to provide a comfortable, healthy and productive indoor environment for the occupants. Indoor terminal units, which have a direct effect on the occupants comfort, can be defined as the building elements that use different heat transfer mechanisms and media to emit and remove heat or moisture from indoor spaces (e.g. hydronic radiant heating and cooling systems, fan-coil units, and active beams). The main differences between HVAC systems in Europe, North America and other parts of the world are often the indoor terminal units. Type of energy sources and energy generators are very much similar. This paper will present the state of the art of energy efficient systems that will provide a good indoor environmental quality at a decreased energy use. Low Temperature Heating and High Temperature Cooling systems are an important requirement for increasing the energy efficiency of HVAC (heating, ventilation and air-conditioning) systems and for increasing the amount of renewable energy used. Especially, these types of systems are getting increasing attention in Europe and North America. In the present study, operation characteristics, capabilities and limitations of different terminal units were specified. Considered terminal units were radiant heating and cooling systems, all-air systems (mixing, displacement, and personalized ventilation), passive and active beams. INTRODUCTION Heating, cooling and ventilation systems play a significant role in the energy use and for the indoor environmental quality in a building. The main purpose of these systems is to provide a comfortable, healthy and productive indoor environment for the occupants. These goals should however be achieved with the lowest possible energy use. This can easily be done by sacrificing the requirement to the indoor environment, i.e. accepting higher room temperatures in summer and lower room temperatures in winter or decreasing the ventilation rate. This may however result in more uncomfortable places of work, increased health risk and lost productivity of the people working in the space. Besides decreasing the energy demand by building design and energy efficiency, another priority is to use as much as possible renewable energy sources like wind, solar and geothermal. The occupant comfort is directly influenced by the indoor terminal units, whereas the type of energy source (fossil fuel, solar, wind, and ground heat exchanger), energy generation (boiler, chiller, heat pump, heat exchanger) and energy distribution system (water, air, electricity) has an important influence on the energy consumption. The main differences between systems in Europe, North-America and other countries are the selected indoor terminal systems. Indoor terminal units are active building components that emit or remove heat and moisture to or from indoor spaces. These indoor terminals mainly rely on convection (natural or forced), radiation or both. A recent research project from International Energy Agency (IEA), Energy in Buildings and Communities (EBC) Program [1], Annex 59 – High Temperature Cooling and Low Temperature Heating in Buildings [2] is studying the currently existing terminal units. A sub-group within the project, Subtask B – Indoor temperature/humidity field and terminal units, is aiming to summarize the indoor heat and moisture sources and the current methods to address them. A further goal of the project is to provide improvement suggestions to the currently existing terminal units and HVAC systems. Terminal units differ in their capabilities of addressing sensible and latent loads, methods of heat emission or removal, maximum heating and cooling capacities, medium of energy distribution, and local or total volume conditioning. This paper summarizes the characteristics of the chosen terminal units (radiant systems, mixing, displacement, and personalized ventilation, passive and active beams) and it aims to function as a simple and reliable reference tool for these units. Capabilities (heating, cooling, ventilation, humidification and dehumidification of indoor air), method (heat transfer mechanism) of heat emission and removal from the indoor space, heating and cooling capacities, and the medium of energy distribution are specified for different terminal units. It is assumed that the reader is familiar with the described systems and concepts. The given values are only intended to provide guidance, and the indicated capacities could vary depending on the specific application. HYDRONIC RADIANT HEATING AND COOLING SYSTEMS A hydronic (water-based) radiant heating and cooling system refers to a system where the water is used as the heat carrier (medium of energy distribution) and more than half of the heat exchange with the conditioned space is by radiation [3]. It is possible to divide the radiant heating and cooling systems into three as follows:

Radiant heating and cooling panels, Pipes isolated from the main building structure (radiant surface systems), and Pipes embedded in the main building structure (Thermally Active Building Systems, TABS) [3].

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The heat carrier (water) circulating in the pipes has low temperatures in the heating and high temperatures in the cooling operation. In some TABS constructions (hollow core concrete decks) also air has been used as a heat carrier, and also electricity can be used in heating applications. Floor, wall and ceilings can be used as surfaces that provide heating or cooling to the space. Hydronic radiant surface systems require a ventilation system to address the latent loads and to provide the ventilation rates required for indoor air quality concerns [3]. Radiant heating and cooling systems enable lower air flow rates than all-air systems where the entire heating and cooling loads are addressed by the ventilation system [4]. The heat emission or removal from the space is achieved by a combination of radiation and convection. Total heat exchange coefficients (combined convection and radiation) for floor heating, wall heating and ceiling heating are 11, 8, 6 W/m2K, and for floor cooling, wall cooling and ceiling cooling are 7, 8, and 11 W/m2K, respectively [3]. The radiant heat transfer coefficient can be used as a constant value of 5.5 W/m2K, with an error of less than 4% [5]. The difference in the total heat transfer coefficients stems from the natural convection. An overview of natural convection coefficients is given in [6]. Based on the acceptable surface temperatures (comfort and dew-point concerns [3]), and assuming an operative temperature of 20°C and 26°C in the room for heating and cooling cases, the maximum heating and cooling capacities can be obtained. The maximum floor (occupied zone) heating and cooling capacities are 99 W/m2 and 42 W/m2, wall heating and cooling capacities are 160 W/m2 and 72 W/m2, and ceiling heating and cooling capacities are 42 W/m2 and 99 W/m2, respectively. In the perimeter zones of the floor, it is possible to obtain a maximum heating capacity of 165 W/m2 [3]. Different studies [3, 7, 8] have shown that the cooling capacity of a floor cooling system increases above the given maximum capacity of 42 W/m2 and may even exceed 100 W/m2, when there is direct solar radiation on the floor. Different construction types of radiant systems can be found in [3]. The design, test methods, control and operation principles of radiant panels are given in ISO 18566:2013 [9], while the design, dimensioning, installation and control principles of embedded radiant systems are given in ISO 11855:2012 [10]. ALL-AIR SYSTEMS Currently, there are eight commonly applied ventilation strategies in buildings. These strategies are mixing ventilation, displacement ventilation, personalized ventilation, hybrid air distribution, stratum ventilation, protected occupied zone ventilation, local exhaust ventilation, and piston ventilation [11]. For the ventilation systems, the main method of heat emission and removal is convection and the medium of energy distribution is air. A review of different airflow distribution and ventilation systems in buildings can be found in [11]. Mixing, displacement, and personalized ventilation systems are further described in this paper. Mixing ventilation Mixing ventilation (mixing room air distribution) intends to dilute the polluted and warm (or cool) room air with clean, cooler (or warmer) supply air. The aim is to achieve a uniform temperature and contaminant distribution in the occupied zone [12]. It is possible to heat or cool a space by mixing ventilation. It is also possible to provide dehumidified and conditioned outdoor air (fresh air). Typical supply air temperature range for heating and cooling is up to 34°C and down to 14°C, respectively [11]. The obtained heating and cooling effect will depend on the ventilation rate. Also in some countries, such as Denmark, the highest permissible supply air temperature is limited to 35°C by regulations [13]. It is not recommended to have a higher temperature difference than 10°C between the supply and room air to achieve proper mixing [12]. According to [14] a specific cooling load of 90 W/m2 can be handled with mixing ventilation systems. Displacement ventilation Displacement ventilation (displacement room air distribution) is based on displacing the polluted room air with fresh air (conditioned outdoor air) [11]. The cool fresh air is supplied with low velocity (0.25-0.35 m/s [15]) at or near the floor, and the supplied air rises by the effects of momentum and buoyancy forces [11, 15]. It is possible to provide cold, dehumidified and conditioned outdoor air with displacement ventilation. Although, it could be possible to provide warmer air than the room air with displacement ventilation (e.g. to heat an unoccupied room before the occupancy [16]) it is not common and it is not recommended due to the short-circuiting of the supply air. Typically, the supply air temperature can be down to 18°C [11]. The cooling load that a floor current displacement system can handle is 30-35 W/m2 according to [15] and 50 W/m2 according to [14]. Personalized ventilation Other than the two mainly applied total volume air distribution principles (mixing and displacement air distribution), another air distribution strategy is personalized ventilation, and it aims at supplying the clean and cool air close to an occupant before it is mixed with the room air [12,17]. The most important advantage of personalized ventilation compared to the total volume conditioning systems is its potential to provide clean, cool and dry air at inhalation [17, 18]. According to [11], the supply air temperature can be down to 20°C in cooling and up to 28°C in heating mode, however it should be noted that perceived air quality (PAQ) might be a problem with the increased supply air temperature [19,20] and ventilation effectiveness may decrease depending on the chosen air supply location and terminal. The required ventilation rates can be calculated based on EN 15251:2007 [21] (this standard is currently under revision [22]), CR 1752:1998 [23], and ASHRAE 62.1-2013 [24].

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BEAMS Although these systems are known as chilled beams, a recent guidebook [25] refers to them as beams, and this terminology will also be used in this paper. Beams (passive and active) are room air recirculation devices that can heat or cool (sensible) a space using water as the energy distribution medium. Active beams can also provide conditioned primary air to a space (they are coupled to the main air-handling unit) [25]. Fresh air is delivered to the space by a decoupled ventilation system in passive beam applications. Beams cannot directly humidify or dehumidify the room air since they operate in dry (non-condensing) conditions but it is possible to control the latent loads and to address the ventilation requirements with active beams [25]. The method of heat emission and removal from the space takes place mainly by convection. Passive beams The performance of passive beams relies on natural convection [25]. In passive beams, the medium of energy distribution from the plant is water. It is possible to heat and a cool space with passive beams but it is not possible to provide fresh air to the space. Although heating is possible with passive beams, in most applications, passive beams are used for cooling only and therefore a separate heating system should be used [25]. Also the ventilation needs should be addressed by a complementing system (e.g. by an air-handling unit) [25]. It is recommended to use passive chilled beams when the total sensible cooling load is up to 40-80 W/m2 [26]. Active beams The performance of active beams relies on convection that is caused by induction [25]. It is possible to heat, cool and provide fresh air to a space by active beams. In active beams, the medium of energy distribution is both air (fresh air, from the air-handling unit) and water from the heating or cooling plant. Active beams can typically be used when the total sensible cooling (air and water) load is less than 120 W/m2 in comfort conditions [25, 26]. The optimum operating range (for a satisfactory thermal comfort in sedentary type occupancy) is 60-80 W/m2 [26]. For the heating case, the optimum operating range is a heating load of 25-35 W/m2 and a maximum heating load of 50 W/m2 [26]. The specific heating and cooling capacities of beams can be found in [26] expressed in W/m. The testing and rating procedures of passive and active chilled beams are given in EN 14518:2005 [27] and in EN 15116:2008 [28], respectively. OTHER SYSTEMS Another type of terminal units that can be used for heating and cooling of buildings is a fan-coil unit. Information regarding fan-coil units can be found in [6, 14, 29]. The descriptions, characteristics, operation principles and other information regarding other terminal units that were not a part of this paper (radiators, radiant tubes, convectors, etc.) can be found in [14,29]. DISCUSSION AND CONCLUSION Capabilities, limitations and characteristics (heating and cooling capacities, governing heat transfer mechanisms and media of energy distribution) of the chosen terminal units are summarized in Table 1. Dynamic building simulation softwares, Computational Fluid Dynamics (CFD), and experimental methods could be used to evaluate the performance of different terminal units, in terms of energy performance, resulting temperature and humidity fields, and occupant thermal comfort. In addition to the characteristics that were addressed in this paper, there are several other factors to consider when selecting a terminal unit (excluding the capital and operational expenditures):

The chosen type of heating and cooling strategy (terminal unit) will have a remarkable effect on the occupant thermal comfort, and criteria such as noise, draft, vertical air temperature difference, temperature drifts, etc. could be limiting factors for the choice and application of different terminal units. The international standards (EN 15251:2007 [21], EN ISO 7730:2005 [30], ASHRAE 55-2010 [31]) should be followed to provide the optimal thermal comfort for the occupants.

It is crucial to consider the transportation and auxiliary energy use (pumps, fans, valves, dampers, sensors, etc.) associated with each terminal unit.

Availability (depending on the location, natural resources, district heating or cooling network, etc.) of the energy sources and sinks, and the possibility of coupling with the terminal unit should be considered.

Control possibilities and principles (e.g. individual room or zone control, control based on flow rate, supply temperature, average temperature, as a function of air temperature, operative temperature or outside temperature) and dynamic behaviors of the terminal units should be considered, e.g. ventilation systems try to keep a constant room temperature while TABS allows a certain temperature drift and keeps the operative temperature within the comfort range rather than a constant value [4]. Another example of the dynamic behavior is the difference between a TABS and a radiant panel, where the radiant panel will be able to affect the thermal conditions in the room faster than a TABS construction, due to its significantly lower thermal mass.

When deciding on which terminal unit to use in a space, all of these issues (occupant thermal comfort, transportation and auxiliary energy use, possible use of heat sources and sinks, control and dynamic behavior) and the capabilities, limitations, and capacities provided in this paper should be considered.

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ACKNOWLEDGMENTS The authors’ participation in IEA EBC Annex 59 was funded by EUDP (Energiteknologisk udvikling og demonstration), project no. EUDP-12-I 2012. REFERENCES [1] International Energy Agency. (2015, February 22). Retrieved from International Energy Agency’s Energy in Buildings and Communities Programme: http://www.iea-ebc.org/ [2] IEA EBC Annex 59. (2015, February 22). Retrieved from High Temperature Cooling & Low Temperature Heating in Buildings: http://www.annex59.com/ [3] Babiak, J., Olesen, B. W., & Petráš, D. (2009). Low temperature heating and high temperature cooling. Brussels: REHVA - Federation of European Heating, Ventilation and Air Conditioning Associations. [4] Olesen, B. W. (2012). Using Building Mass To Heat and Cool. ASHRAE Journal, 44-52. [5] Olesen, B. W., Michel, E., Bonnefoi, F., & De Carli, M. (2000). Heat Exchange Coefficient Between Floor Surface and Space by Floor Cooling - Theory or a Question of Definition. ASHRAE Transactions, Part I, 684-694. [6] Liu, X., Jiang, Y., & Zhang, T. (2013). Temperature and Humidity Independent Control (THIC) of Air-conditioning System. Springer-Verlag Berlin Heidelberg. doi:10.1007/978-3-642-42222-5 [7] Olesen, B. W. (1997). Possibilities and Limitations of Radiant Floor Cooling. ASHRAE Transactions, Part I, 42-48. [8] Simmonds, P., Holst, S., Reuss, S., & Gaw, W. (2000). Using Radiant Cooled Floors to Condition Large Spaces and Maintain Comfort Conditions. ASHRAE Transactions, 695-701. [9] ISO 18566. (2013). Building environment design - Design, test methods, control and operation of radiant heating and cooling panel systems. Geneva: International Organization for Standardization. [10] ISO 11855. (2012). Building environment design - Design, dimensioning, installation and control of embedded radiant heating and cooling systems. Geneva: International Organization for Standardization. [11] Cao, G., Awbi, H., Yao, R., Fan, Y., Sirén, K., Kosonen, R., & Zhang, J. (2014). A review of the performance of different ventilation and airflow distribution systems in buildings. Building and Environment(73), 171-186. doi:10.1016/j.buildenv.2013.12.009 [12] Müller, D. (Ed.), Kandzia, C., Kosonen, R., Melikov, A. K., & Nielsen, P. V. (2013). Mixing Ventilation - Guide on mixing air distribution design. Brussels: REHVA - Federation of European Heating, Ventilation and Air Conditioning Associations. [13] DS 469, 2nd edition. (2013). Heating- and cooling systems in buildings. Charlottenlund: Danish Standards. (in Danish). [14] Recknagel, H., Sprenger, E., & Schramek, E.-R. (2011). Taschenbuch für Heizung und Klimatechnik. München: Oldenbourg Industrieverlag. (in German). [15] Awbi, H. (2003). Ventilation of Buildings (2nd ed.). London: Spon Press. [16] Skistad, H. (Ed.), Mundt, E., Nielsen, P. V., Hagström, K., & Railio, J. (2004). Displacement ventilation in non-industrial premises. Brussels: REHVA - Federation of European Heating, Ventilation and Air Conditioning Associations. [17] Melikov, A. K. (2004). Personalized ventilation. Indoor Air, 14 (Suppl. 7), 157-167. [18] Melikov, A. K., Skwarczynski, M. A., Kaczmarczyk, J., & Zabecky, J. (2013). Use of personalized ventilation for improving health, comfort, and performance at high room temperature humidity. Indoor Air(23), 250-263. [19] Fang, L., Clausen, G., & Fanger, P. O. (1998). Impact of temperature and humidity on the perception of indoor air quality. Indoor Air, 8(2), 80-90. [20] Fang, L., Wyon, D. P., Clausen, G., & Fanger, P. O. (2004). Impact of indoor air temperature and humidity in an office on perceived air quality, SBS symptoms and performance. Indoor Air, 14 (Suppl. 7), 74-81. [21] EN 15251. (2007). Indoor environmental input parameters for design and assessment of energy performance of buildings addressing indoor air quality, thermal environment, lighting and acoustics. Brussels: European Committee for Standardization. [22] Olesen, B. W. (2012). Revision of EN 15251: Indoor Environmental Criteria. REHVA Journal, 6-12. [23] CR 1752. (1998). Ventilation for buildings - Design criteria for the indoor environment. Brussels: European Committee for Standardization. [24] ASHRAE. (2013). ASHRAE Standard 62.1, Ventilation for Acceptable Indoor Air Quality. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers. [25] Woollett, J., & Rimmer, J. (Eds.). (2014). Active and Passive Beam Application Design Guide. Brussels: REHVA - Federation of European Heating, Ventilation and Air Conditioning Associations. [26] Virta, M. (Ed.), Butler, D., Gräslund, J., Hogeling, J., Kristiansen, E. L., Reinikainen, M., & Svensson, G. (2007). Chilled Beam Application Guidebook. Brussels: REHVA - Federation of European Heating, Ventilation and Air Conditioning Associations. [27] EN 14518. (2005). Ventilation for buildings - Chilled beams - Testing and rating of passive chilled beams. Brussels: European Committee for Standardization. [28] EN 15116. (2008). Ventilation in buildings - Chilled beams - Testing and rating of active chilled beams. Brussels: European Committee for Standardization. [29] ASHRAE. (2012). ASHRAE Handbook - HVAC Systems and Equipment. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers.

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[30] EN ISO 7730. (2005). Ergonomics of the thermal environment - Analytical determination and interpretation of thermal comfort using calculation of the PMV and PPD indices and local thermal comfort criteria. Brussels: European Committee for Standardization. [31] ASHRAE. (2010). ASHRAE Standard 55, Thermal Environmental Conditions for Human Occupancy. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers.

Table 9. The chosen terminal units and their corresponding capabilities, limitations and operational characteristics (Y: Yes, N:

No, NC: Not Common).

Capabilities Method of heat emission or removal Capacity (W/m2) Medium of energy distribution

Name of terminal Type Heating Cooling Ventilation

(fresh air) Humidification + Dehumidification Convection Mainly

convection Radiation Mainly radiation Heating Cooling Air Water Electricity

Radiant systems

Floor* Y Y N N Y N Y Y 99 42 Y Y Y

Wall Y Y N N Y N Y Y 160 72 N Y Y

Ceiling Y Y N N Y N Y Y 42 99 Y Y Y

Air systems**

Mixing ventilation Y Y Y Y N Y N N 34°C 14°C Y N N

Displacement ventilation Y/N Y Y Y N Y N N NC 18°C Y N N

Personalized ventilation Y/N Y Y Y N Y N N NC 20°C Y N N

Beams Passive Y Y N N N Y N N NC 80 N Y N

Active Y Y Y Y*** N Y N N 50 120 Y Y N

*: Floor in the occupied zone. **: For air systems, typical maximum and minimum supply air temperatures are provided [11]. The heating and cooling capacity will depend on the ventilation rate. ***: Humidification and dehumidification is possible with the primary air and it should be done by the air-handling unit. Beams operate in dry, non-condensing conditions.

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[Abstract:0071][Energy Efficient Buildings] ENERGY EFFICIENCY APPROACHES IN DESIGN OF METRO STATIONS

İbrahim Ethem Özbakır1 and Levent Tosun1

1BİLGE Engineering and Consulting Co. Ltd, Ankara, Turkey Corresponding email : ozbakir@metu.edu.tr, ltosun@metu.edu.tr

SUMMARY During the last decades extensive investment made in urban transport, especially in construction of metro lines, in Turkey. It seems that this trend will go on for several more decades. Metro stations require intensive heating, ventilating and air-conditioning (HVAC) systems. Those systems, together with sanitary installations, consumes energy in the form of of electricity for operation. During the design of new metro stations there are several approaches to decrease energy consumption and increase energy efficiency. In this paper, architectural and mechanical engineering issues are examined and some proposals are made to achieve this goal. INTRODUCTION Demand of transportation at dense city centers enforces railway systems which generally use underground routes and stations. By ”rule of thumb” a city reaching to one million population should evaluate metro system as one of the sustainable choice to satisfy the demand. Metro systems do not only shorten the travel time, noise level and decrease traffic / air pollution created by private cars but also decrease the amount of energy consumption in passenger transportation, in cities. Even though energy consumption per passenger is smaller in metro systems compared to private car systems the total value is still quite high. So it needs to be discussed and energy efficiency ways during the design stage need to be found to decrease the value. During the last decades extensive investment made in construction of metro lines in Turkey and it seems this trend will continue for several decades. Keeping in mind, also, that Turkey is a net energy importer country energy efficiency issues are important in design of metro systems. ENERGY CONSUMPTION AND DISCIPLINES RELATED WITH IT Metro systems require electric energy for the operation and large part of the energy is spend for ‘traction of the trains’. Although there is no data available for the ‘energy consumption of the stations’ it may be assumed as 1/3 of the total value. Stations shall have building services such as mechanical (HVAC, sanitary, drainage, fire extinguishing), electrical (lighting, power), electronic (signalling, telecommunication) and electro-mechanical (elevator, escalator) systems. Like other buildings, architecture, mechanical and electrical engineering are the relevant disciplines in energy consumption and efficiency issues in station design. At this paper only architectural and mechanical ways to decrease energy consumption will be discussed. ARCHITECTURAL DESIGN APROACHES TO DECREASE ENERGY CONSUMPTION Thermal insulation of the construction material: When ‘architecture’ and ‘energy conservation’ terms come together people get used to think about thermal insulation of the fabric of the building. In metro stations number and also area of the rooms requiring heating and cooling are quite small. If a metro station is at ground or above ground level these rooms should have thermal insulation. However, if a metro station is underground type these rooms face free or paid areas or technical corridors. Temperatures of such neighbour rooms are at moderate levels. Therefore, for cities like İstanbul, İzmir (which has rather hot winter outside temperatures) thermal insulation of the fabric of the heated and cooled rooms is not crucial. But because of winter outdoor temperature (-12ºC), the walls of the rooms of underground stations of Ankara metro system should have thermal insulation. Positioning and sizing of electrical (power) rooms:

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Electrical (power) rooms like transformer, MV switchgear, LV main panel room dissipate heat energy. In order to operate properly, those rooms should be kept at 40ºC maximum temperature. This goal is achieved by ventilation or DX air-conditioning systems depending upon summer outdoor DB temperature, both consuming electric energy. If such rooms are positioned in stations such that to have more outdoor wall and ceiling area facing underground, some of the dissipated heat will sink into soil and mechanical systems to cool the room temperature shall operate for shorter time yielding less energy consumption. Increasing the volume of such rooms will also help to decrease energy consumption. Avoiding reserve rooms: In some stations because of excavation and structural enforcements some extra space is created and architects design them as “reserve rooms”. Such non-functional rooms increases not only constructional, mechanical and electrical investment costs but also electricity consumption as well. Best solution for such extra spaces is to add them into electrical (power) rooms since no additional energy consumption will be necessary. Another solution might be to add them into technical corridors since technical corridors shall have less ventilation rates than reserve rooms. Grouping air-conditioned rooms together and positioning: Electronic (signalling, telecommunication) rooms, because of their function, are cooled in winter and summer by split or VRF type air-conditioners. Staff (station operation, personnel, ticket, health, security) rooms are heated in winter and cooled in summer by split or VRF type air-conditioners. Grouping such rooms together will decrease heat loss and positioning them away from the electrical (power) rooms which dissipate heat will decrease heat gain resulting less energy consumption. An example of such grouping is shown in Figure1. Figure 1. Grouping of air-conditioned rooms (example).

Utilizing skylights as natural exhaust ventilation stacks: Underground stations have quite large free and paid areas which require artificial lighting and consuming electricity even day time. Designing skylights will positively effect the psychology of the passengers and increase daylight zone resulting decrease in energy consumption. However, one step further, instead of simple skylights architects may use ‘smoke and natural ventilators’. During daily working conditions, such equipment will serve for ‘lighting’ purpose and also can be used for ‘smoke exhaust’

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during fire. Additionally, it will extract the exhaust air of public free and paid areas of the concourse level of the station, by natural way. Since no electricity is required for the operation energy shall be conserved. A cross section of such a station is shown in Figure 2. Figure 2. Skylight to be used as natural ventilation stack (example).

Providing suspended ceiling to staff rooms: Staff rooms sometimes have suspended ceilings, sometimes not, depending upon the decision of the authorities. Suspended ceilings will cause additional construction cost, but decrease HVAC investment cost and energy consumption in terms of heat loss and heat gains and ventilation flow rate (if the design criteria is accepted as ‘air change per hour (a.c.h.)’) throughout the whole life of the station. Since staff room areas are relatively small (8-15 m2) and their height is relatively great (4.5-5.5 m) constructing suspended ceiling will also avoid ‘deep well effect’ from interior architectural point of view. Designing waste water sump pits at upper levels: Metro stations have wet cores (staff and passenger toilets, janitor’s room, staff sink) creating waste water. Since the collecting pipes, in the station, are below city waste water network pipe level, sump pit and pump system should be designed to collect and raise the waste water. In order to decrease head of the pumps and save investment cost and decrease energy consumption it is a good practice to place the wet cores and sump pits at upper floors like ‘concourse’ or ‘technical’ levels. ‘Platform’ levels might have such pits wherever the station is the first and the last station of the metro line to serve to train drivers (if the system operates with drivers). HVAC DESIGN APPROACHES TO DECREASE ENERGY CONSUMPTION Özbakır states ‘Metro stations should be equipped with a series of heating, ventilating and air-conditioning installations to function properly and at the same time to satisfy passenger and staff comfort’ and explain HVAC design in stations of underground rail system [1].

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Applying realistic air flow rates for ventilation: Technical specifications used for the design and construction of metro systems are usually taken from twenty-twentyfive years old documents and only small modifications are made. According to those technical specifications air flow rates for ventilation system design are expressed as ‘a.c.h.’ criteria. Table-1 shows some basic rooms and related air flow rate criteria usually used. Table-1. Minimum air flow rate criteria for metro stations (existing).

Room name Air flow rate (a.c.h.) Platform, Free and paid areas of concourse level 8 Staff 6 Toilet 10 Mechanical 6 Electrical 10 Electronical 4 - 6 Battery 6 - 15

Generally, if the function of the room is not known at the design phase, using a.c.h. criteria might be useful. However in metro stations every room has specific function and should be ventilated accordingly. Public areas like platforms, free and paid areas, passenger corridors are large areas with great heights. Therefore 8 a.c.h. value yields to huge and unnecessary ventilation flow air. Realistic criteria might be ‘m3/h per passenger.’Each staff room will have 1-2 persons as occupants. So 6 a.c.h. value, again, will give unnecessary ventilation flow. This fact also effect heat loss and heat gain therefore air-conditioning investment cost and energy consumption increases. Fresh air and exhaust air requirement should depend upon ‘number of occupants’ and ‘area of such rooms’. Toilet room exhaust air value should depend upon to the number of units that exist, not to the volume of the room. Air flow rates for mechanical and electronical rooms are very high; since they are no-man rooms and almost no air pollution source exists in them. As mentioned before, electrical rooms are kept at certain indoor temperature (40 °C) by the help of fresh air and exhaust air ventilation system, if the outdoor temperature in summer time is at an acceptable level. For this case a.c.h. criteria can not be applied. If the outdoor temperature in summer time reaches to indoor design temperature, then DX air conditioning system shall be used. In this case, ventilation flow rate might be decreased to corridor criteria (1.1 m3/h,m2). Air flow rates used for battery rooms are triple times more than the figures given in ASHRAE Handbook [2], again yielding high energy consumption in terms of ventilation and air-conditioning. Mechanical engineers usually accept that American Society of Heating, Refrigeration and Air-Conditioning Engineering (ASHRAE) provides realistic flow rates for ventilation design. Using ANSI / ASHRAE Standard 62.1-2013, Ventilation for Acceptable Indoor Air Quality [3] one might suggest the minimum air flow rates as given in Table-2. Table-2. Minimum air flow rate criteria for metro stations (suggested).

Room name Air flow rate Platform, Free and paid areas of concourse level 9 m3/h, person + 1.1 m3/h,m2 Staff 18 m3/h, person + 2.2 m3/h,m2 (*) Toilet (staff) 90 m3/h, unit Toilet (public) 125 m3/h, unit Mechanical 9 m3/h,m2 Electrical (Cooled by ventilation system) Calculate according to heat dissipation value of

the equipments Electrical (Cooled by air-conditioning system) 2.2 m3/h,m2 (*) Electronical 2.2 m3/h,m2 (*) Battery 4 a.c.h.

(*) : Figures are doubled from ASHRAE Standard to provide better environmental comfort.

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Using heat reclaim ventilators for staff rooms: Staff rooms should have forced ventilation for 24 hours, if they are underground. Ceiling hanged compact ventilation units are advised to use for both exhaust and fresh air. Since staff rooms are air-conditioned, exhaust air shall heat fresh air in winter and cool in summer, resulting energy saving, if heat reclaim ventilators are used. Of course, high enthalpy exchange efficient units, like cellulosic type, to transfer not only heat but also humidity is a better choice. Using inverter type air-conditioning equipment: Instead of using on-off type air-conditioning equipment, inverter type will decrease energy consumption. Using thermostat controlled ventilation systems : Usually small electrical rooms (LV panel, escalator panel, etc.) are ventilated on a.c.h. criteria. However, these are no-man rooms and at off-peak hours inside temperatures are usually quite lower than acceptable levels. Therefore thermostat controlled ventilation systems will consume less energy. No fresh air supply for the battery rooms: Because of old written technical specifications, sometimes, ducted ‘fresh air supply’ ventilation system is required for the battery rooms by authorities. However, standards require ‘exhaust ventilation to extract gasses’ which might occur in the room. If make-up air is taken into room from neighbouring technical corridor through a transfer grille placed at the wall, then system will be simplified and less energy will be consumed. Using plug type fans and high efficiency electric motors: In general plug type fans shall consume less energy compared to V-belt centrifugal fans. New generation electric motors of fans will also decrease energy consumption. SANITARY DESIGN APROACHES TO DECREASE ENERGY CONSUMPTION Grouping and type of hot water supply: Some sanitary appliances in wet cores (staff wash basin and sink, janitor’s wash basin and sink, sump pit room wash basin) require hot water supply. In order to save energy, such sanitary appliances should be grouped together. Electrical hot water producing unit might be selected as ‘storage type’ and the maximum pipe length from unit to appliance advised to be within 7 m. If sump pit room is far away from hot water producing unit, ‘instantaneous type’ unit should be preferred instead of storage type since the time of usage is very short and rare. Selecting proper type of domestic water supply system: Sanitary appliances in wet cores shall require domestic water supply. To avoid shortages and break-downs in city water network to effect the metro station wet cores water storage tank and booster set is necessary. Two solutions are available for the design. First solution is to connect city network service pipe only to water storage tank, use booster set continuously to provide domestic water to sanitary appliances. Second solution is to connect city network service pipe to, both, water storage tank and domestic water supply header. When city network has enough pressure, water will directly go to consuming points. When the pressure drops booster set will automatically function; it will take water from storage tank, pressurize and deliver to water supply header. Second solution will save more energy than the first solution, because under normal conditions stations will have no water pressure problem and shortages and break-downs in city network are quite rare in big cities in Turkey. Using high efficiency electric motors : Waste water and drainage water pumps should be desgned with high efficiency electric motors to consume less energy.

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RESULTS Metro stations have many opportunities at design phase to decrease energy consumption and use energy more efficiently. Architectural approaches for this goal might be summarized as; thermal insulation of the construction material, positioning and sizing of electrical (power) rooms, avoiding reserve rooms, grouping air conditioned rooms together and positioning, utilizing skylights as natural exhaust ventilation stacks, providing suspended ceiling to staff rooms, designing waste water sump pits at upper levels. On the other side, mechanical approaches like; applying realistic air flow rates for ventilation, using heat reclaim ventilators for staff rooms, using inverter type air-conditioning equipment, using thermostat controlled ventilation system, having no fresh air supply for the battery rooms, using plug type fans and high efficiency electric motors, grouping hot water using wet cores, choosing the type of hot water supply, selecting proper type of domestic water supply system might be evaluated for the same purpose. REFERENCES 1. Özbakır, E.2005. Yer Altı Raylı Sistem İstasyonlarında Isıtma, Havalandırma, Klima Tesisatı Tasarımı, VII. Ulusal Tesisat Mühendisliği Kongresi Bildiriler Kitabı. 2. ASHRAE Handbook.2014, ASHRAE Heating, Ventilating and Air Conditioning Applications, Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. 3. ASHRAE.2013. ANSI/ASHRAE Standard 62.1-2013, Ventilation for Acceptable Indoor Air Quality, Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc.

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[Abstract:0072][Modeling and Software] THE GENERATION OF TYPICAL METEOROLOGICAL YEAR AND CLIMATIC DATABASE

OF TURKEY FOR THE ENERGYANALYSIS OF BUILDINGS

Serpil Yılmaz1, İsmail Ekmekçi2 and Mustafa Yılmaz1 1 Marmara University, İstanbul

2 İstanbul Commerce University, İstanbul Corresponding Author: serpilbozkurtyilmaz@gmail.com

SUMMARY For sustainable development, a reduction in energy demand is essential. This could be achieved through improving energy efficiency, effective energy conservation and management. The weather conditions of a given region are the most important considerations for the proper design of space AC systems. In this study, The typical meteorological year and climatic database of Turkey for the energy analysis of buildings were generated by SQL database programmimg language. The Finkelstein–Schafer statistical method was applied to analyze the hourly measured weather data of a 23-year period (1989–2012).and select representative typical meteorological months (TMMs). The selection criteria was based on 13 meteorological parameters. These parameters are the daily mean, maximum and minimum values and ranges of temperature, dew-point and wind velocity and the daily values of global solar radiation. According to results of TMY, climatic database of turkey including daily or hourly climate variables was created in sql data tables. 1. INTRODUCTION The design of energy requirements and thermal comfort of buildings requires an updated and very accurate climatological and solar database. A climatological and solar database database is very important for calculation of energy efficiency. The hourly amounts of about 10–13 meteorological parameters such as solar radiation, dry bulb temperature, relative humidity, wind speed, atmospheric pressure and etc. are usually needed for energy simulation. A representative database for a year duration is known as a typical meteorological year (TMY), a term mainly used in the USA, or a test reference year (TRY) or a design reference year (DRY), terms mainly used in Europe. TMY, TRY or DRY consists of individual months of meteorological data sets selected from different years over the available data period, which is called a long term measured data

The primary objective of these methods is to select single years or single months from a multi-year database, preserving a statistical correspondence. This means that the occurrence and the persistence of the weather should be as similar as possible in the TMY to all available years. These different TMY methodologies have been developed with selection criteria based on solar radiation or on solar radiation together with other meteorological variables [1,2,3,4,5] The literature review conducted in this work shows that one of the most common methodologies for generating a TMY is the one proposed by Hall et al. using the Filkenstein–Schafer (FS) statistical method . The other methodologies cited above for generating TMY use a modified version of it. This method is an empirical approach that selects individual months from different recorded years from. The selection criteria was based on 13 meteorological parameters. These parameters were the daily mean, maximum and minimum values and ranges of temperature, dew-point and wind velocity and the daily values of global solar radiation. However, four of 13 parameters were considered to be less effective, and therefore, are given zero weight. These variables are the ranges of daily dry-bulb temperature, 228wet-bulb temperature and wind speed, and daily minimum wind speed. Except for a few changes to the weighting criteria, which account for the relative importance of the solar radiation and meteorological elements, there has been no change in the original methodology which has been adopted by different countries [6,7,8,9]. 2. REVIEW ON TYPICAL METEOROLOGICAL YEAR The literature review conducted within the framework of this work shows that one of the most common methodologies for generating a TMY is the one using the Filkenstein–Schafer (FS) statistical method [9]. The other methodologies cited above for generating a TMY use a modified version of it. This method is an empirical approach that selects individual months from different years from the period of record. The selection criteria were based on 13 meteorological parameters. These parameters are the daily mean, maximum and minimum values and ranges of temperature, dew point and wind velocity and the daily values of global solar radiation. However, 4 of the 13 parameters are considered to be less effective and, therefore, are given zero weight. These variables are the ranges of daily dry bulb temperature, wet bulb temperature and wind speed, and daily minimum wind speed. A TMY consists of the months selected from the individual years and sorted to form a complete year. In the literature, there are many attempts to produce weather databases for different locations . The main objective of these methods is to select representative months from the multi-year database.

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Except for a few changes to the weighting criteria, which account for the relative importance of the solar radiation and meteorological elements, there has been no change in the original methodology and it has been adopted by different countries: for example, by date of publication, for Holmet Stations [10], Athens [11], Egypt [12] Ibadan, Nigeria [13], Hong Kong [14], Nicosia,Cyprus [15], Saudi Arabia [16], Malaysia [17], Damascus, Syria [18], Recently, ASHRAE has started an international Project to develop TMY data throughout the world, the international weather year for energy calculations (IWEC) [5]. Most recently, using the FS method, Kalogirou developed TMYs for the city of Nicosia, Cyprus. The study of Kalogirou included additional variables such as illuminance, visibility, precipitation and snow fall data [19]. The objective of the present work is to select and implement TMY generating methodologies using long term hourly measured meteorological and global solar radiation data. For Turkey, only three attempts have been found in the literature for the generation of TMY datasets [20,21,22]. Pusat S, Ekmekçi I, Akkoyunlu T generated TMY for 8 cities. Ecevit et al. generated the TMY for Ankara [22]. They stated that solar radiation data was unreliable in Turkey. Therefore, they evaluated the possibility of using the daily sunshine duration or the ratio of the daily sunshine duration to the day-length instead of daily global solar radiation, as the ninth parameter, in obtaining TMY. They used the data of Ankara covering the period 1979-1999. On the other hand, there is no any other attempt to validate the results of their study. In the paper of Üner and İleri, TMYs for 23 cities representing demographic and climatic conditions of Turkey were investigated by using actual recordings (1990-1996) [26]. They generated the typical meteorological database of 23 locations for building simulations and air-conditioning design. The only deficiency in this study is the number of years used in the generation of TMY. There isn't enough study to generate TMY datasets for Turkish locations in the literature. In Ref. [22], TMY datasets was generated just for Ankara, and number of years used is not enough for TMY generation.

3. PROBLEM DEFINITION IN MEASUREMENTS AND DATA In this study, the meteorological data was obtained from The State Meteorological Affairs General Directorate (DMI) and cover a period of 1989-2012 for 81 cities throughout the Turkey. Meteorological stations are located in city centers and there is generally only one station in each city .There were missing and invalid measurements in the data and they were filled as null. So, the data were checked for wrong entries and missing data. The missing and invalid measurements, accounting for approximately 0.30% of the whole database, were replaced with the values of preceding or subsequent days by interpolation. In the calculations, the year was excluded from the database if more than 15 days measurements were not available in a month. 4. TMY SELECTION METHOD For each station, nine daily meteorological parameters: maximum air temperature (Tmax), minimum air temperature (Tmin), mean air temperature (Tmean), maximum air relative humidity (RHmax), minimum air relative humidity (RHmin), mean air relative humidity (RHmean), maximum wind speed (Wmax), mean wind speed (Wmean) and global solar radiation (G) were employed to create an indicator for selecting typical months (TMM). The weighting factors used are selected according to existing experience on the influence of the meteorological parameters used on the simulated application. Three sets of weighting factors, all oriented towards energy simulation applications were used, as shown in Table 1. Table 1 Weighting Factors for TMY type PRESENT (FS) Weather İndex [23] [24] [17] [25,26] 1/24 5/100 5/100 1/20 Maximum dry bulb temperature 1/24 5/100 5/100 1/20 Minimum dry bulb temperature 2/24 30/100 30/100 2/20 Mean dry bulb temperature 1/24 2.5/100 1/20 Maximum dew point temperature 1/24 2.5/100 1/20 Minimum dew point temperature 2/24 5/100 1/20 Mean dew point temperature 2/24 5/100 5/100 1/20 Maximum wind speed 2/24 5/100 5/100 1/20 Mean wind speed 12/24 40/100 40/100 5/20 Total horizontal solar radiation 5/20 Direct Solar Radiation 10/100 Relative Humidity In the first step , for a given parameter xi, a long-term cumulative distribution function (CDFm) of xi for each month covering the period of 23-year (1989–2012) was created. A short-term cumulative distribution function (CDFy,m) of xi for year y and month m was also generated.. Finkelstein-Schafer (FS) statistics are the most common methodology for creating CDF functions while generating typical weather data. This method is an empirical methodology for selecting individual months from different years over the available period.According to FS statistics (Filkenstein and Schafer, 1971 ), if a number, n, of observations of a variable X are available and have been sorted into

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an increasing order X1 , X2,…,Xn, the cumulative frequency distribution Function (CDF) of this variable is given by a function Sn(X) which is defined as follows:

The FS by which comparison between the long-term CDF of each month and the CDF for each individual year of the month was done is given by the following equation:

İ , ∑ , ,

where İ , is the Finkelstien–Schafer statistics of the parameter for year y and month m, j is interval number of data and is the total number of data intervals. In the second step, the weighted sums of İ were computed by , ∑ . , with the condition :

1

where is the weighting factors for the FS of the variable and Np is the total number of the parameters. In this case, the weighting factor for Tmax, Tmin, RHmax, RHmin is 0.04, for Tmean, RHmean, Wmax,Wmean is 0.08 and for global radiation is 0.5 . All individual months are ranked in ascending order of WS values.

( 1 )

(2 )

( 3 )

( 4 )

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5. RESULTS

CITY CITY CODE NAME

1 2 3 4 5 6 7 8 9 10 11 1217020 BARTIN 2004 2005 2004 2005 2004 2008 2003 1989 2003 2005 2004 200517022 ZONGULDAK 2004 2008 2012 1990 1997 2004 2005 1989 2012 1989 2012 199417026 SİNOP 2004 1990 2011 1990 2005 2002 1990 1990 2008 1989 2007 198917030 SAMSUN 1995 1998 1994 2001 1995 1998 1995 1999 1989 1997 1995 199717033 ORDU 2012 1990 2004 1995 2009 1992 2009 1991 2007 1997 2006 199317034 GİRESUN 2009 1989 2009 1990 2012 1990 2012 1989 2009 1990 2009 199017037 TRABZON 2004 1989 2004 1990 2004 1989 2003 1993 2004 1989 2004 198917040 RİZE 2009 1990 2010 1990 2010 1991 2009 1995 2009 1990 2009 199017045 ARTVİN 2011 1990 2007 1990 2010 1991 2008 1995 1995 1989 2012 199117046 ARDAHAN 1998 2002 1994 1990 2011 2009 2011 1992 2009 1989 2009 200317050 EDİRNE 2011 1994 2011 1992 2011 2002 2010 2010 2010 1994 2009 199317052 KIRKLARELİ 2008 1989 2011 1990 2011 1990 2011 1989 2011 1989 2011 198917056 TEKİRDAĞ 2011 1990 2011 1990 2011 1993 2010 1998 2009 1989 2008 198917062 İSTANBUL 2006 1990 2006 1990 2005 1990 1991 1994 2007 1991 2001 199017066 KOCAELİ 2009 1990 2011 1990 2009 1990 2010 2007 2007 1990 2012 199017069 SAKARYA 2011 1990 2012 1990 2008 1992 1998 2012 2006 1990 2012 199017070 BOLU 2009 1990 2009 1992 2011 1993 2000 1994 1998 1990 2012 199017072 DÜZCE 2009 1989 2011 1990 2008 1993 2009 1993 2006 1990 2009 199017074 KASTAMONU 2012 1990 2011 1992 2012 2000 2003 2003 2010 1990 2009 198917078 KARABÜK 2010 2008 2011 2000 2011 2009 2010 2010 2000 2000 2008 200717080 ÇANKIRI 2009 1990 2005 1990 1990 1990 2002 1991 1991 1990 2009 199017084 ÇORUM 2009 1990 2006 1990 2010 1990 1999 1989 2003 1989 2008 198917085 AMASYA 2004 1994 2010 1990 2005 1990 2008 1994 2010 1989 2008 198917086 TOKAT 2011 1989 2011 2006 2009 1989 2010 1989 2009 1990 2009 199017088 GÜMÜŞHANE 2007 2004 1991 1990 2010 1990 2001 1989 2008 1989 2009 199017089 BAYBURT 2012 1990 2011 1990 2011 1990 2011 1991 2006 1990 2008 199017090 SİVAS 2011 1989 2007 1992 2010 1993 2008 2008 2006 1990 2011 199517094 ERZİNCAN 2011 2003 2010 1995 2010 1989 1998 1989 2009 1992 2009 199717096 ERZURUM 2006 1995 2007 1990 2005 1990 2006 2005 2006 1990 2007 199717097 KARS 2000 1992 2007 1993 2009 1994 2008 1991 2000 1992 2003 199117099 AĞRI 2009 1989 2009 2000 2010 2009 1996 1994 2009 1990 2009 200317100 IĞDIR 2009 1990 2009 1990 2009 1990 2009 2003 2005 1989 2009 199017112 ÇANAKKALE 2011 1991 2011 1992 2011 2010 1994 2008 2010 1990 2012 198917116 BURSA 2011 1991 2011 1990 2005 1990 1999 1989 2009 1990 2009 199017119 YALOVA 2011 1990 2011 1990 2011 1990 2011 1994 2011 1990 2009 199417120 BİLECİK 2011 1990 2011 1990 2011 1990 2011 1994 2011 1990 2009 199417126 ESKİŞEHİR 2012 2012 2011 2010 2009 2007 2011 2007 2011 2007 2008 201217130 ANKARA 2007 1990 2007 1992 2006 2011 2004 2000 2004 1990 2008 199017135 KIRIKKALE 2009 1989 2007 1990 2010 1989 2008 1989 2006 1990 2012 199017140 YOZGAT 2012 1990 2010 1991 2009 1990 2011 1989 2008 1990 2012 198917152 BALIKESİR 1993 1990 1995 1990 1990 1991 1994 1996 1995 1990 1995 198917155 KÜTAHYA 2008 1990 2007 1998 2009 1995 1999 1995 2006 1989 2012 199017160 KIRŞEHİR 2009 1990 1991 1990 2005 1989 1999 1992 1998 1990 2009 198917165 TUNCELİ 2009 1991 2009 1991 2007 1990 2001 1998 2008 2007 2009 199117172 VAN 2006 1991 2007 1993 2007 1994 1996 1999 2006 1990 2008 199017186 MANİSA 2011 1991 2012 1990 2006 1991 2005 1993 1995 1991 2005 198917188 UŞAK 2011 2006 1999 2004 2012 1996 2006 1999 2011 1990 2012 198917190 AFYONKARAHISAR 2004 1990 2011 2005 2009 1993 2010 1989 2000 1990 2011 198917192 AKSARAY 2011 1993 2008 1990 2010 2007 2007 2007 2009 1990 2009 198917193 NEVŞEHİR 1996 1989 2011 1990 2011 1989 2011 1989 1998 1989 2009 198917196 KAYSERİ 2009 1989 2009 1992 2009 1989 2006 1996 2009 1990 2009 199017199 MALATYA 2012 1994 2011 1993 2012 1995 2006 1995 2009 1992 2008 199317201 ELAZIĞ 2009 1991 2011 1990 2009 1991 2004 1998 1994 1990 2009 199317203 BİNGÖL 2011 1991 2009 1990 2010 1995 2004 2002 2008 1990 2009 199117204 MUŞ 2006 1993 2009 1990 2010 1991 2006 2004 2008 1991 2008 199117207 BİTLİS 2009 1993 2009 2004 2008 1997 2006 2000 2008 1991 2009 199717210 SİİRT 2009 1991 2011 1996 2010 1995 2009 2005 2005 1990 2007 200017220 İZMİR 2011 1990 2012 1990 2009 2012 1996 2011 1990 1990 2010 198917234 AYDIN 2007 1991 2011 1990 2009 2009 2000 1992 2011 1991 2012 199717237 DENİZLİ 2011 1990 2011 1990 2010 1997 1998 1994 2011 1998 2007 199517238 BURDUR 2011 1990 2011 1990 2009 1993 2006 2006 2006 1990 2012 199017240 ISPARTA 2006 1990 2005 1999 2012 2010 2002 2007 1993 2007 2012 199017244 KONYA 2004 1990 1994 1991 1991 1995 1991 2002 2002 1997 2001 199717246 KARAMAN 2009 1990 2011 1990 2009 1991 2007 1996 2006 1990 2009 200317250 NİĞDE 1996 1989 2005 1990 2010 1991 1999 1996 2009 1989 2009 199717255 KAHRAMANMARAŞ 2011 1991 2007 1991 2008 2007 1998 2008 2006 1991 2012 199717261 GAZİANTEP 2006 1991 2011 1991 2009 1995 2004 1999 2006 1994 2008 199317262 KİLİS 2006 1990 1991 1991 2005 2005 1998 1992 2006 1991 2012 199317265 ADIYAMAN 2009 1990 2011 1991 2004 2002 1990 1990 2001 1998 2009 199817270 ŞANLIURFA 2005 1990 2010 1991 2010 1995 2009 1991 2007 1990 2005 200517275 MARDİN 2006 1990 2011 1991 2011 1994 2011 1989 2011 1990 2009 198917280 DİYARBAKIR 2009 1990 2011 1991 2010 2001 2006 2001 2008 1990 2009 199917282 BATMAN 2009 1990 2011 1990 2011 2010 2007 2007 2011 1990 2009 199017285 HAKKARİ 2011 1990 2012 1993 2010 1989 2011 1989 2010 1989 2009 199017287 ŞIRNAK 2011 2000 2011 2003 2011 2002 2007 2001 2011 1991 2009 199917292 MUĞLA 2011 1990 2011 2001 2012 2002 1996 2000 2010 1991 2009 198917300 ANTALYA 1995 2006 1998 2005 1998 2006 1999 2002 1998 2006 1997 200417340 MERSİN 2009 2004 2009 1991 2009 1993 2007 1999 2009 1990 2009 199017351 ADANA 2006 1990 2011 1990 2010 1993 2010 1996 2009 1989 2007 198917355 OSMANİYE 2012 1990 2012 1990 2012 1989 2012 1992 1997 1990 1997 199017372 HATAY 2006 1991 2007 2005 2009 2003 2007 2007 2007 1991 2007 1989

TYPİCAL METEOROLOGICAL YEARS MONTHS ( 1- 12 )

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6. CONCLUSIONS Energy consumption in Turkey is increasing continuously parallel to its development. Because of its limited energy resources,Turkey is heavily dependent on imported oil and gas . Therefore, every means to use energy in a much more rational way should be taken into consideration. Heating, cooling, ventilating and air conditioning (HVAC) systems are major energy users in residential and commercial buildings. The first step in the design of air-conditioning systems is the calculation of heating and cooling loads of the building that depend on its characteristics, the indoor conditions to be maintained,and on outside weather conditions. If the air-conditioning system is expected to provide the indoor conditions specified (comfort conditions) at all times, it should be designed for peak conditions that are determined by the most extreme weather data recorded for the locality in which the building is located. This approach, however, will result in oversized air conditioning equipment,which in turn, will increase the initial equipment cost and the operating cost. It is very important to represent the climate of a location. In this study, the typical meteorological year (TMY) for 81 cities of Turkey was calculated. It will be very useful source for building simulations to estimate the annual energy consumptions of buildings . 7. ACKNOWLEDGEMENT I would like to express my special appreciation and thanks to my co-supervisor Prof. Dr. İsmail EKMEKÇİ , you have been a tremendous mentor for me. I would like to thank you for encouraging my research and for your priceless advices on both research as well as on my career. I would also like to thank my advisor Assist. Prof. Mustafa YILMAZ and Msc advisor Prof. Dr. Birol Kılkış for his guidance and help.I would also like to thank Research & Technical Services Manager of Ashrae, Mr Michael Vaughn and TC 4.2 member Mr Didier Thenevard for very valuable support in USA .A special thanks to my family. I would like express more appreciation to my beloved husband Levent YILMAZ who spent sleepless nights with and was always my support in the moments when there was no one to answer my queries.. 8. REFERENCES [1] Klein SA. A method of simulation of solar processes and its applications. J Solar Energy 1975;29. [2] Klein SA, Beckman WA, Duffie JA. A design procedure for solar heating systems. J Solar Energy 1976;113. [3] Schweitzer S. A possible ‘‘average’’ weather year on Israel’s coastal plain for solar system simulations. J Solar Energy 1988;21. [4] Hall IJ, Prairie RR, Anderson HE, Boes EC. Generation of typical meteorological years for 26 SOLMET stations. Sandia Laboratories Report, SAND 78-1601, Albuquerque, NM, 1978. [5] Crow LW. Weather year for energy calculations. ASHRAE J 1984;6. [6] Feuermann D, Gordon JM, Zarmi Y. A typical meteorological day TMD approach for predicting the long-term performance of solar energy systems. J Solar Energy 1985;35. [7] S. Wilcox and W. Marion, “User’s manual for TMY3 data sets,” Technical Report, NREL/TP-581-43156, April 208. [8] Gazela, M., and Mathioulakis, E., 2001. “A new method for typical weather data selection to evaluate long-term performance of solar energy systems”, Solar Energy, Vol. 70, No. 4, pp. 339-48. [9] Yang HX, Lu L.The development and comparision of typical meteorological years for building energy simulation and renewable energy applications . ASHRAE Trans 2004:110(2):424-31 [10] Hall IJ, Prairie RR, Anderson HE, Boes EC. Generation of typical meteorological year for 26 SOLMET stations. Sandia Laboratories Report, SAND 78-1601. Albuquerque, New Mexico; 1978. [11] Pissimanis D, Karras G, Notaridou V, Gavra K. The generation of a typical meteorological year for the city of Athens. J Solar Energy 1988;40. [12] Shaltout MAM, Tadros MTY. Typical solar radiation year for Egypt. Renewable Energy 1994;4(4):387–93. [13] Fagbenle RL. Generation of a test reference year for Ibadan, Nigeria. Energy Conversion and Management 1995;36(1):61–3. [14] Wong WI,Ngan KH. Selection of an ‘’example weather year ‘’ for Hong Kong. Energy build 1993:19 (4) 313-6 [15] Petrakis M, Kambezidis HD, Lykoudis S. Generation of a "typical meteorological year" for Nicosia, Cyprus. Renewable Energy 1998;13(3):381–8. [16] Said SAM, Kadry HM. Generation of representative weather-year data for Saudi Arabia. Applied Energy 1994;48(2):131–6. [17] Rahmana IA, Dewsbury J. Selection of typical weather data (test reference years) for Subang, Malaysia. Building and Environment 2007;42(10):3636–41. [18] Skeiker K. Generation of a typical meteorological year for Damascus zone using the Filkenstein–Schafer statistical method. Energy Conversion and Management 2004;45(1):99– [19] Kalogirou, S.A., 2003. “Generation of typical meteorological year (TMY-2) for Nicosia, Cyprus”,Renewable Energy, Vol. 28, No. 15, pp. 2317-34. [20] Ecevit, A., Akinoglu, B.G., and Aksoy, B., 2002. “Generation of a typical meteorological year using sunshine duration data”, Energy, Vol. 27, No. 10, pp. 947-54. [21] Pusat S, Ekmekçi İ, Akkoyunlu T. Generation of typical meteorological year for different climates of Turkey.

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[22] Merter U, Arif I. Typical weather data of main Turkish cities for energy applications. Int J. Energy Res 2000; 24(8) : 727 – 48 [23] Finkelstein JM,Schafer RE.Improved goodness of–fit tests. Biometrica 1971:58(3):641-5 [24] ASHRAE, Ashrae handbook Fundamentals, 2001. [25] Bahadori MN, Chamberlain NJ. A simplification of weather data daily and monthly energy needs of residential buildings 1986:36(6):499-507. [26] Siurna DL, D’Andrea LJ, Hollands KGTA. Canadian representative meteorological year for solar system dimulation In:10th annual conference of the solar energy society of Canada (SESCI’84). Calgory, Alberta, Canada:August 1-6 1984.

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[Abstract:0073][Exergy] AN EXERGY-BASED AUTOMATION SYSTEM IN ESER LEED PLATINUM BUILDING

Ayşe Gülbeden 1, Prof. Dr. Şefik Bilir2 , Prof. Dr. Birol Kilkis3

1Eser Holding, Ankara, Turkey 2 Selçuk University, Konya, Turkey

3Baskent University, Ankara, Turkey Corresponding email: bkilkis@baskent.edu.tr

SUMMARY

Different building loads like steam, hot water, chilled water, power, and service hot water change almost instantly with different proportions and the electromechanical systems bundle like a cogeneration plant, absorption chiller, wind turbines, solar PV, and solar collectors, and heat pumps cannot match and follow these constantly changing load magnitudes and proportions. In a TUBİTAK funded research project an ASHRAE BACNet protocol compatible electromechanical system automation program based on exergy efficiency was developed and tested. A one-year run of the program at full scale and scope in the Eser LEED Platinum building in Ankara has showed that the Rational Exergy Management Model Efficiency has increased by 20% on an annual average basis on top of the already achieved high energy and exergy efficiencies through LEED certification. This program may be applied on 15-minute time intervals to existing buildings, retrofit buildings (provided that sufficient Thermal storage systems are implemented), or simulation and equipment selection and optimum system design of new buildings to be architectured over a suitable building simulation program.

INTRODUCTION

A green and hybrid electromechanical system consists of several energy sources, energy conversion systems, and varying time based performance values, as a requirement of the system. Eser Green Building is also designed and constructed as a high performance green building and has various green and hybrid systems incorporated within its electromechanical structure. The building has platinum certification from LEED. Winter and summer operation diagrams of ESER Green Building are given in Figure 1 and Figure 2. It is almost impossible for these different systems to work together in harmony and supply various energy and power demands of the building and achieve the desired energy savings with the use of the existing Building Management Systems. An automation algorithm for high performance buildings, based on exergy balance between supply and demand was developed. This software is called as “Rational Exergy Automation (AEO) Program”. The main objective of the algorithm is to deliver exergy from on-site sustainable systems and other equipments to various demand points with maximum supply and demand exergy balance. Increasing the balance reduces exergy destructions and thus compound CO2 emissions [1, 2, 3, 4]. The method is based on Rational Exergy Management Model (REMM) [1].

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Figure 1. Winter operation scheme for ESER Green Building.

Figure 2. Summer operation scheme for ESER Green Building. The Rational Exergy Efficiency, ψRi, for such a system (i) is proportional to the supply and demand exergy balance. The Rational Exergy Efficiency can be calculated by Equation 1 if exergy is destroyed prior to application(s).

(1)*

236

Here, εsup is the supply exergy and εdem is the demand exergy for unit power consumption. Supply and demand exergies are based on the ideal Carnot Cycle and they can be, respectively, calculated by Equation 2 and Equation 3.

(2)

(3)

Here;

Tref

, Ambient reference temperature [K], T

f, Supplied exergy (or equivalent ) temperature [K],

Tapp

Temperature at exergy demand point [K]. * if exergy is destroyed before major application(s).

The basic exergy block shown in Figure 3 is used for a set of various exergy supplies and

demands.

Figure 3. Basic Exergy Block.

By using this algorithm hourly (or on any selected time interval) heat load demand of the building is exergy-optimally allocated to different systems shown in Figure 1 and 2 supplied to meet the rational exergy efficiency in a highly efficient manner. In other words, supply and demand exergies are matched by using Rational Exergy Management Model (REMM).

The hourly heat load demands are satisfied by the thermal storage buffer tanks (real interfaces) and electric power panel boards.

DIVIDE and CONTROL

By using the Divide and Control method (Figure 4) the Rational Exergy Automation can be used both for the existing buildings and at the design level by operating the system virtually. Hence, the selection of the most efficient electromechanical system components becomes possible. Rational Exergy Automation (AEO) communicates with Conventional Building Management Systems by using Direct Data Exchange (DDE).

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Figure 4. Logic of the developed algorithm. The software permits the use of two different languages (Turkish and English) and two separate unit systems (SI and IP). The software offers three different objectives options to the user. These are: as to operate the system at

- Low operating cost, - High exergy efficiency, or - Low CO2 emission levels.

At selected time intervals the software prepares reports giving user input data, performance of all electromechanical system components, Renewable Energy Ratio, Rational Exergy Efficiency, CO2 emission (kg/hr) according to 1st and 2nd law of Thermodynamics, commercial values of CO2 (TL/hr) and fuel saving (%). Average Rational Exergy Efficiency of the whole system is calculated by Equation 4 in the algorithm. Here i and j sub-indices respectively represent the supply and demand points/systems or loads. Qij (kW-h), shows the required energy to supply demand (j) by system (i). ηij shows the 1st law efficiency of the process for supplying the demand (j) by system (i).

(4)

The algorithm calculates total CO2 emission ( kg/hr) by Equation 5 Here, ci is CO2 emission coefficient based on fossil fuels and cj, indicates CO2 emission coefficient for existing power plants. Assumed value of ci for natural gas is 0.2 kg CO2/kW-h, for sustainable systems it is 0 and cj can be assumed as 0,40 kg CO2/kW-h. ηT is overall efficiency of the power plant and it can be assumed as 0.27 for Turkey.

(5)

Thermal efficiency according to 1st law is calculated by Equation 6. Here Qi is total heat load, Ei is total electricity load (grid supplied), QT is total heat supply and ET is total power supplied with renewable energy.

(6)

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Alternative Energy Ratio is calculated by Equation 7. Here, QH is the total heat consumption of the system and QAH is supplied heat load by the alternative energy systems. Similarly, QE is the total electricity consumption of the system and QEH is supplied electricity load by the alternative energy systems.

(7)

Equation 8 shows the Primary Energy Savings percentage calculation. [CHP]η1E is partial power efficiency and [CHP]η1H is partial heat efficiency of CHP. [REF]η1H is reference heat efficiency (assumption value 0.95) and [REF]η1E reference power efficiency (assumption value 0.52). RefΨR is 0,21. Other reference values are given by EU/2004/8/EC directive [5].

R

R

H

H

H

HR

fREFCHP

REFCHP

PES

2Re2

11

1

1

1

1

(8)

CO2 emission reduction rate (According to baseline building) is calculated by Equation 9. Here, is the baseline assumption value (between 0.04 and 0.06).

(9)

Total Energy Cost (TL/hr) (Natural gas and grid electricity cost for baseline condition) is calculated by Equation 10.

(10)

Total energy cost ( TL/hr) (If natural gas powered CHP is operating) is calculated by Equation 11.

(11)

Sample on-line automation screen copies are given in the following figures. Supplied electricity from grid and the PV can be viewed in the software. Environmental variables such as pressure, wind velocity and direction around the building and global radiation can be monitored. Moreover, operation mode (winter/summer) and fuel type can be selected from the interface.

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Figure 5. Sample output from winter operation with AEO.

Figure 6. Sample output from winter operation.

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RESULTS

Table 1. Annual Energy Consumptions in ESER Green Building. ESER GREEN BUILDING ANNUAL ENERGY CONSUMPTION (kWh/m2)

YEARS ELECTRICITY kWh/m2

NATURAL GAS kWh/m2

TOTAL CONSUMPTION kWh/m2

Baseline 78,4 71,1 149,5

Proposed 41,1 98,6 139,7

2010 71,9 52,5 124,4

2011 91,8 89,8 181,6

2012 90,2 71,6 161,8

2013 56 46,1 102,1

2014 47,4 52,5 99,9 Table 1 shows the baseline, proposed and, for 5 years period, actual electricity and natural gas consumptions of the ESER Green Building. In the baseline scenario, the building is heated by natural gas boiler and cooled by electrical chiller. In the proposed case, simulated and expected values of consumption based on existing design is shown. 2010 data may be neglected because commissioning was still in progress. During 2011- 2012 years, the building was operated with conventional method. During 2013-2014 years new developed software was used.

Figure 7. Comparison of the Annual energy consumptions.

Figure 7 shows the comparison between baseline, proposed scenario and actual annual consumptions. As mentioned above, during 2010 commissioning was still in progress thus real consumption at that year is neglected. Real consumption of 2011-2012, especially 2011, was significantly higher than the both baseline and proposed scenarios. The large drop in the actual consumption values is clearly seen in Figure 7, for the years 2013 and 2014 where the software was used. During the 2013-2014 period annual consumption was stabilized. Moreover, approximately 30% and 35% less consumption, with respect to the proposed and baseline scenarios respectively was achieved. This results show that proper automation strategy and applications is crucial to succeed of complicated hybrid systems. The Rational Exergy Management Model based automation algorithm was proven successfully to decrease energy consumption and CO2 mitigation.

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As mentioned before, the software and the system were designed for both new buildings and existing buildings and retrofitted buildings thus, the method and software offer an important energy consumption and CO2 mitigation decrement potential.

ACKNOWLEDGEMENT

This project was supported by TUBİTAK 3120176 project. This support is greatly appreciated.

SYMBOLS

[CHP]η1E CHP partial power efficiency [CHP]η1H CHP partial heat efficiency [REF]η1E Reference power efficiency ( assumption value 0.52) [REF]η1H Reference heat efficiency ( assumption value 0.95) ci CO2 emission coefficient based on fossil fuels, kg CO2/kW-h (Assumed value for natural gas is 0,2

kg CO2/kW-h , and for sustainable systems it is 0) cj CO2 emission coefficient for existing power plants, kg CO2/kW-h. (assumed value 0,40 kg CO2/kW-

h) E Power (kW-h). Ei Total Electricity Load (grid supplied) (kWh) ET Total power supplied with renewable energy ( kWh ) Q Heat Load for the system or a system component (kW-h). QAE Electricity Power provided from renewable energy sources (or energy). Unit exergy value is assumed

as 1 QAH Heat Power provided from renewable energy sources with εAH exergy (or energy). QE Total Power demand of the building. QH Total heat load demand of the building. Qi Total Heat Load (kWh) QT Total heat supplied , (kWh) T

app Temperature at exergy demand point [K].

Tf, Supplied exergy (or equivalent ) temperature [K],

Tref

, Ambient reference temperature [K], TL Turkish Lira εdem Exergy demand for any task εHH Average heat load exergy of the building heat loads εsup Actual supplied exergy for any task ηT Overall efficiency of the Power Plants (Assumed value for Turkey is 0.27).

REFERENCES

1. Kilkis, B., Kilkis, Ş. Energy and Exergy Efficiency Comparison of Poly-Generation and Co-generation Systems, Conference

Proceedings, (In Serbian), Proceedings of the 40th Congress on HVAC&R – KGH, Vol. 22, pp: 474-486, 2-4 December, Belgrade, 2009.

2. Kilkis, Ş. Sustainable Development of Energy, Water and Environment Systems Index for Southeast European Cities, Journal of Cleaner Production, 1-13, 2015.

3. Kilkis, Ş. Energy System Analysis of a Pilot Net-Zero Exergy District, Energy Conversion and Management 87, pp. 1077–1092, 2014.

4. Kilkis, Ş. Exergy Transition Planning for Net-zero Districts, Energy 92, pp. 515-531, 2015. 5. EN, Directive 2004/8/EC of the European Parliament and of the Council of 11 February 2004 on the Promotion of

Cogeneration Based on a Useful Heat Demand in the Internal Energy Market and Amending Directive 92/42/EEC, L 52/50 Brussels, 2004.

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[Abstract:0074][Energy Efficiency in Healthcare Buildings] ENERGY MANAGEMENT IN HOSPITALS

M. Zeki YILMAZOGLU

Gazi University Faculty of Engineering Department of Mechanical Engineering Maltepe ANKARA zekiyilmazoglu@gazi.edu.tr

SUMMARY Hospitals are one of the buildings that consume energy intensively. The energy use intensity (EUI) of a hospital is approximately 800-1000 kWh/m2. Decreasing the EUI and increasing the efficiency can be achieved by an effective energy management model. In this study, points to take into consideration about energy management in hospitals are discussed. Energy audits, benchmarking studies, related metrics, surveys, and the importance of the PDCA cycle are mentioned. In addition, the focus points in energy efficiency measures are indicated. It is proposed that as an energy policy, implementations of the energy management in hospitals can be actualized with support mechanisms in a pilot region which will be selected by the Ministry of Health. The results of energy management can be disseminated in accordance with the Ministry of Energy. INTRODUCTION Hospitals are critical buildings where different air conditioning requirements are needed for different zones and energy, particularly electricity, must be provided without interruption. Energy management in a building with these features are vital. The purpose of the energy management is the observation of minimum cost and environmental effect in the manufacturing of a product or providing a service. The definition can be detailed as follows. Energy management is the reduction of the unit cost of a product or service by optimization and regulation strategies for the reduction of energy requirements. The Energy Management Department in a plant is responsible for reducing the energy costs and the environmental impact without affecting the production and quality. In this responsibility, the Energy Management Department is in charge of monitoring the share of energy costs in production or service, recovery and recycling potentials (especially for heating and cooling applications in a building service), following the latest technologies for maintenance and expansions, obtaining the highest efficiency for energy consuming devices, and minimization of the losses. Energy Use Intensity (EUI) is a definition or metric which expresses the annual energy consumption per unit area for buildings. Lombard et al. [1] calculated EUIs of different types of buildings. The results showed that EUI for schools and hospitals were calculated to be 262 kWh/m2 and 786 kWh/m2, respectively. Hospital EUI were found to be 1200-1500 kWh/m2 in an Energystar project report [2]. Both of these studies showed that the hospitals are in the second range among the different types of buildings. In the buildings that consume energy intensively such constitution of energy management department is definitely a must. Energy Management Department present the development reports and future projections to the senior management directly. In this study, points to take into consideration about energy management in hospitals were discussed. Energy audits, benchmarking studies, related metrics, surveys, and the importance of the PDCA cycle were mentioned. In addition, focus points in energy efficiency measures were indicated. Simple measures can be taken on energy efficiency focus are indicated. It is proposed that as an energy policy implementations of the energy management in hospitals can be actualized with support mechanisms in a pilot region which will be selected by the Ministry of Health. The results of energy management can be disseminated in accordance with the Ministry of Energy. ENERGY MANAGEMENT MODEL The most important issue in energy management is the support of the senior management. Senior management has to be convinced of the importance of the energy management. Inherently, financial support for the proposed projects covered by the senior management. In some projects with higher payback period, it is sometimes hard to find the support of the senior management. However, it must be noted that energy cost is a part of the unit production or service cost and generally energy part of the cost can be decreased by improving the energy efficiency. For the purpose of energy management internet-based commercial softwares, applications, cloud-based systems etc. have been under development and used. These softwares create a database for the management and energy costs of the plant can be monitored according to various metrics in real time. Having historical data and comparison of these with the new data can guide the energy manager for the measures to be taken in some points or systems. However, the reasons of the sharp peaks and declines have to be investigated in detail. In the buildings like hospitals, these peaks are generally stem from the seasonal variations. For instance, electricity consumption has peaks in summer season due to increasing cooling demand.

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Detailed energy audits are the prime mover of the implementation of energy management model. During the audit phase energy consumption values in the previous years have to be investigated in detail and the reasons of the seasonal consumptions for the buildings, particularly for the hospitals, have to be examined. In addition, the habits of the users in hospitals should be examined. Many office equipment (with the exception of the connected to the medical devices) never turned off and the lights of the personnel rooms and warehouses switch on permanently in the hospitals depending on the usage habits. In an existing facility technical services staff who will implement this program need to be convinced in order to be successful in energy management program. Energy manager, serves as a bridge between senior management and technical services, also need to be convinced of the need for energy management both departments. Surveys may also be useful in order to determine the user behavior. Air conditioning and energy related issues of the hospital staff can be determined by these surveys and problem areas can be identified in this way without relying on measurements. For example, with a survey study applied to the operation room staff some clues can be obtained on laminar flow units. An example survey that can be applied to the operation room staff is given in Table 1. The satisfaction status in this survey is based on ASHRAE seven-point scale. Similar surveys have to be performed for patients and technical staff. Following these studies problematic areas in terms of energy consumption in the hospital will be easily determined. Addressing these problematic areas or issues firstly in the creation of the main strategy is of paramount importance to ensure convictions to the energy management model. In addition to this survey questions about the strengths and weaknesses of the system, operating staff must be asked. Moreover, one may be asked about the parameters that disrupts thermal comfort in the operating rooms. These are hot or cold surfaces on the floor, ceiling, and walls, infiltrations, noise level, the response time of the system, local thermal discomfort and so on. Table 1. Thermal comfort survey for operation room staff Gender Male Female Age <30 31-40 41-50 51-

60 >60

Thermal Comfort inside the operation room

1 Hot

2 Warm

3 Slightly warm

4 Neutral

5 Slightly cool

6 Cool

7 Cold

Operation room temperature 1 Not satisfied

2 3 4 5 6 7 Very satisfied

Temperature fluctuation during the operation

1 Constant

2 3 4 5 6 7 Not constant

Humidity 1 Too dry

2 3 4 5 6 7 Too humid

Odor 1 Always

2 3 4 5 6 7 Never

Air quality 1 Stuffy

2 3 4 5 6 7 Airy

Air conditioning system control 1 Not satisfied

2 3 4 5 6 7 Very satisfied

Pressurization 1 Not familiar

2 3 4 5 6 7 Familiar

Laminar Flow Unit 1 Not familiar

2 3 4 5 6 7 Familiar

The adoption of the energy management strategy in a hospital should not be limited to only the senior management and technical staff. Energy management strategy should be explained to all parties and need for measures to be communicated to all staff. For this purpose, a meeting has to be organized in order to create an awareness. In this meeting, the energy management strategy of the administration has to be shared with attendants and some basic tips for energy efficiency should be explained. Starting from the historical data on energy consumption of the building long term projections have to be shared.Short and medium term energy efficiency increment projects should be realized adhering to the time schedule and should be completed. For example, in a replacement project with low efficiency induction motors to higher efficiency induction motors in AHUs time table has to be needed. In any efficiency improvement project the budget has to be approved by senior management and economic analysis has

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to be performed. Although the simple payback method gives an idea about the implementation of the proposed project, detailed economic analyses with the time value of the money have to be presented to the administration. The most implemented model in energy management is PDCA (Plan-Do-Check-Act) cycle. PDCA cycle is a total quality management approach that can be applied to the system at any size on any subject. PDCA is based on KAIZEN (continuous improvement) strategy. Planning is the most important step in PDCA cycle that has to be based on detailed energy audits. In the energy audits, initially current energy consumption metrics have to be determined. For example, energy consumption per square meter or energy consumption per bed. After performing energy audits for the building, short, medium, and long term efficiency improvement projects have to be determined with respect to the results of the detailed energy audits. Payback time and decrement in energy consumption have to be calculated. A detailed schedule for each project has to be planned. In this step, energy management team has to work in accordance with hospital staff in order to avoid future disruptions. In the first place, in order to prove the impacts of the energy management model with respect to the KAIZEN strategy by small measures will provide an increased commitment to the model. The second step in the PDCA cycle requires actualization of project which the budget, schedule, and economic analysis were determined. Check step is the controlling of the success of implemented project with measurements. In this step, the measurement methodology and devices with proper measurement ranges have to be determined. For example, if a boiler’s performance will be investigated after the project a calibrated flue gas analyzer is needed to determine the exhaust losses. Also, thermal comfort can be measured in an operation room. In this case, an anemometer (hot wire or equivalent), a radiant temperature device, a humidity and temperature measurement device etc. have to be used. The deviations are determined with respect to the measurement results. If the application is considered as successful standardization takes place. If the deviations are greater than expected the measures investigated again. Gantt charts which depend on time and work packages of the project are an indispensable tool for these types of applications. Any efficiency improvement projects should be handled according to the PDCA cycle into the Gantt charts. Table 2 shows an example of Gantt chart used for lighting system renovation of a hospital. As shown in Table 2, initially current situation has to be investigated for the comparison of the results after renovation. Required lighting intensities of the zones differ due to the aim of the usage of the zone. CIBSE [3] lighting guide shows particular lighting levels and practices to be applied in the hospitals. The measurements have to be performed on the ground or on the work space. After the renovation project, new lighting intensities have to be measured in order to check the required lighting level. TS EN 12646-1:2013 gives required lighting intensities for different zones. In a lighting system renovation project, the effects of the energy consumption have to be calculated. The report should consists of annual energy consumption data before and after implementation, economic analyses, decrement in CO2 emissions, etc. Table 2. Lighting system renovation project in a hospital (Gantt chart example) Weeks 1 2 3 4 5 6 7 8 Current situation lighting intensity measurements Comparison with standards Lighting system required level determinations Purchase order Installation Comparison with standards Reports Energy audits are essential for the creation of the PDCA cycle in an energy improvement project. The measurement devices to be used in these analyses should be selected properly and these have calibration certificates. The report also consist of the calibration certificates and the measurement precisions of the devices in order to calculate error percentage. Efficiency improvement projects will be determined after pre-audits and detailed audits. Some of the improvement projects with short term payback time have to be accomplished instantly. Preventing steam or pressurized air leakages are the examples of such projects. Thus, observation of these supply and return lines during the pre-audit phase will be useful in determination of the possible leakage points. Outputs that will be achieved by the implementation of these projects increase the confidence of the stakeholders in the energy management strategy. In some instances, long term retrofitting projects with higher payback time can be implemented. For example, in parallel to technological developments energy efficiency of the energy consuming devices may be increased. This example is similar to developments in a chiller. It might be mandatory to replace the chiller with the developments in refrigerants and agreements between countries which involve ozone depletion potential and global warming potential of the refrigerants. Although some changes on the chiller let the staff to run the device, it is not practical and economical. In such cases, changing the chiller is the most efficient way due to developed new technologies and more efficient refrigerants. Ice storage with dual operating mode chillers substantially decrease the cost of the cooling system. According to the data from the Ministry of Health [4], the number of hospitals and beds in 2014 were stated as 874 and 123703, respectively. The occupancy rate of these beds were stated as 69% and 72% in 2013 and 2014. The number of university hospitals was 69 in 2014. A technical report on the energy costs of the hospitals has not been prepared yet. Some non-

245

governmental organizations conducted workshops on the energy efficiency in hospitals in recent years. However, none of the reports have tangible results. A benchmarking study in a pilot area, selected by the Ministry of Health, has to be started. In this benchmarking study, the results have to be obtained with different metrics such as kWh/m2, kWh/bed, and MWh/(air conditioned area/patient) [5]. In the last metric the total number of the patients should be considered without a discrimination inpatient or outpatient. Energy efficiency applications that can be performed in hospitals are presented in ASHRAE AEDG (Advanced Energy Design Guide) according to the energy consumption points [6]. ASHRAE and DIN standards present information about design temperature, humidity, air change rates, outside air requirements etc. However, there is no standard in a consensus of Infection Committees in hospitals. The Ministry of Health should stipulate all of the hospitals with a regulation on energy management applications. However, proper subsidized mechanisms with efficiency improvement projects have to be adopted. The validations of these projects have to be performed by third parties principally certificated universities. Hospitals should not be considered as ordinary buildings due to their dynamics. Therefore, the buildings such as hospitals require a detailed certification program. The scope of the program should include AHUs, measurements of air change rates for different zones, enumeration of particles, classification of the zones, HEPA filters, laminar flow units, chillers, boilers, lifts, cooling towers, thermal storage etc. ENERGY EFFICIENCY FOCUS POINTS IN HOSPITALS Ensuring the efficient use of energy and determining the improvement potential some of the focus points are given as follows. In a hospital building, these focus points should be considered together [8-9].

1. Heating and cooling systems 2. Air handling units 3. Hot water and/or steam 4. Lighting 5. Other energy consuming devices 6. Lifts 7. Laundry and catering services 8. Refrigeration systems 9. Medical gases 10. Building envelope 11. Pressurized air systems

Heating and cooling systems are directly related to thermal comfort in a zone. It is of importance to provide thermal comfort conditions for patients and staff. In order to provide thermal comfort conditions in a zone control systems and sensors have to be selected and positioned properly. The feedbacks on thermal comfort in a zone from the staff working in that zone with the survey, given in Table 1, supply information about the situation of the air conditioning devices. ASHRAE 170 2013 provides design temperatures and humidity for different zones in a hospital [7]. In operation rooms (Class B and C), the design temperatures and humidity are specified as 20-24°C and 20-60%, respectively. For instance, in neonatal intensive care unit design temperature and humidity are specified as 22-26°C and max. 60%, respectively. It might be problematic to provide thermal comfort conditions in transition seasons. Sudden decrements and peaks for the outside air temperature and humidity make difficult to provide thermal comfort conditions in a zone. Thermostats are used to control the zone temperature. However, improper positioning of these devices distort the thermal conditions of the zone. Solar effects, draughts, radiators, etc. may generate wrong signals to the controller unit of the air handling device. A dead temperature band has to be defined for thermostat controls. The dead band might be 20°C to 26°C. Air handling units are set to be out of service within this dead band range and consequently, energy consumption is decreased. Separate controls are required for different zones in a hospital. ASHRAE 170 provides minimum air change rates and minimum outside air requirements for different zones in a hospital. Also pressurization conditions of the zones are specified. For instance, an operation room has to be kept at a positive pressure in order to prevent air leakage from the corridor side. For Class B and C operation rooms’ minimum air change rate and the minimum outside air requirement are specified as 20 ad 4, respectively. For intensive care units these values are 6 and 2, respectively. Building Energy Management Modules can be effective to ensure these controls. Therefore, for new hospital constructions building energy management modules have to be integrated with proper sensors and controls. In order to control these values in a zone the fans of the AHU have to be inverter controlled. As is known, halve the fan speed means eight times reduction in power consumption. Hot water is a service that should be provided at all times in a hospital. Hot water is generally produced in boilers and the combustion efficiency of the boilers should be monitored periodically. O2 trim is a control method in order to provide efficient combustion in boilers. Of course, issues to be considered here are the proper locations and maintenance of the sensors. Piping and related instruments are the veins of the building heating-cooling system and any bottleneck in the piping system negatively affects the equipment such as boilers, pumps, control valves etc. Legionella is another issue in hospitals that is related to the water system (piping, water tanks, cooling towers etc.). Lighting systems constitute 20-30% of total electricity consumption in hospitals. A well-designed lighting and automation systems reduce the share of lighting consumption. Occupancy sensors have to be used in staff rooms or warehouses. Daylighting possibilities should also be considered where available. Efficient interior and exterior lighting systems and efficient ballasts should be used in order to decrease the

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consumption. Laundry and catering services have a considerable amount of share in total energy consumption. Hood design and the locations of the cooking appliances can reduce the energy consumption of the exhaust fans. Also, heat recovery potential has to be considered especially from air to water which might be used to heat the service table instead of electric resistances. The kitchen and laundry staff should be trained on energy efficiency particularly the maintenance of the devices. It is possible to save energy in refrigeration systems by defrosting. In addition, increasing the set temperature 1°C decreases the energy consumption between 2-4%. Medical gases, particularly O2, has a vital role in the hospitals. Oxygen is stored cryogenically and during phase change energy can be recovered. Induction motor inventory can provide an energy-saving reference. Replacing older induction motors with higher efficiency types will provide a significant energy saving. Lifts are also use these types of induction motors in older versions and these consume significant amount of electricity especially in high-rise hospital buildings. The new generation of lift systems use polyurethane straps instead of traditional steel wire ropes, gearless permanent magnet synchronous motors instead of asynchronous motors, and closed loop modern drives instead of traditional drives. The new generation type lifts can also generate and store electricity to be used in the case of power failure. Electricity is the most important energy type for a hospital. Electricity has to be supplied to the hospital in 8 seconds via standby generators. Of course, insulation of building envelope and piping system substantially reduce the energy consumption. In addition, cogeneration and trigeneration systems can be used to reduce the energy costs. CONCLUSIONS Hospitals are critical buildings where different air conditioning requirements are needed for different zones and energy, particularly electricity, must be provided without interruption. Decreasing the EUI and increasing the efficiency can be achieved by an effective energy management model. In this study, points to take into consideration about energy management in hospitals were discussed. Energy audits, benchmarking studies, related metrics, surveys, and the importance of the PDCA cycle were mentioned. In addition, focus points in energy efficiency measures were indicated. Simple measures can be taken on energy efficiency focus are indicated. It is proposed that as an energy policy implementations of the energy management in hospitals can be actualized with support mechanisms in a pilot region which will be selected by the Ministry of Health. The results of energy management can be disseminated in accordance with the Ministry of Energy. REFERENCES

1. Lombard L.P., Ortiz J., Pout C., A review on buildings energy consumption information, Energy and Buildings, 40, 394-398, 2008.

2. Internet,http://www.energystar.gov/buildings/facility-owners-and-managers/existing-buildings/use-portfolio-manager/understand-metrics/what-energy, Access date: 27/012014.

3. CIBSE Chartered Institution of Building Services Engineers, Lighting Guide 2: Hospitals and Healthcare Buildings 2009.

4. Internet, http://rapor.saglik.gov.tr/istatistik/rapor/index.php, Access date:18/02/2016. 5. CIBSE Chartered Institution of Building Services Engineers, Lighting Guide 2: Hospitals and Healthcare Buildings

2009. 6. ASHRAE, Advanced Energy Design Guide for Large Hospitals, Achieving 50% Energy Savings Toward a Net Zero

Energy Building, 2012. 7. ANSI/ASHRAE/ASHE Standard 170 2013, Ventilation of Healthcare Facilities. 8. Kapoor R., Kumar S., Energy Efficiency in Hospitals Best Practice Guide, USAID ECO-III Project, 2011. 9. Internet,http://newscenter.lbl.gov/feature-stories/2009/06/02/working-toward-the-very-low-energy-consumption-

building-of-the-future/, Access date: 27/01/2014.

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[Abstract:0075][Renewable Energy Systems and Applications]

ENERGY AND EXERGY ANALYSIS OF WATER AND AIR COOLED PVT SYSTEMS WITH HEAT PIPE TECHNOLOGY FOR SUSTAINABLE BUILDINGS

Birol Kılkış1, Tuğberk Ozar2, Ali Suavi Aktun3 1 Fellow ASHRAE, Baskent University, Ankara, Turkey

2 Enover, Inc., Ankara, Turkey 3 Enover, Inc., Ankara, Turkey

Corresponding email: bkilkis@baskent.edu.tr

SUMMARY Especially under high irradiance and high ambient temperatures the temperature of the PV cells increases significantly. This increase of temperature causes efficiency reduction down to almost 8~9% and thus results a reduction of total power generation by about 15~20%. The main function of a PVT system is to preserve the nominal efficiency of PV cells at ambient conditions as much as possible by cooling the PV cells either by circulating air or water at the back of the panels. In water circulated PVT designs, not only efficiency lost is prevented by a large amount, but also pre-heating of hot domestic water can be achieved. The problem about circulating water behind PV panels is high pumping costs. The aim of this study is to design a PV panel cooler with heat pipes and to reduce the pumping costs through an optimum design approach using CFD techniques. According to CFD analyses, up to 12% of the original power generation capacity can be achieved by the new heat pipe technology. This paper also compares air cooling by finned heat pipes using natural convection and chimney effect. These results however show that technically feasible option is water cooling by heat pipes and by generating useful heat if there is a hot water demand of suitable magnitude and load profiles. INTRODUCTION Within the scope of the project, Enover INC. shall carry out studies on prevention of efficiency loss of PVs due to temperature rise of solar cells during the generation of electricity from solar energy by cooling panels with heat pipes. For this purpose, necessary cooler designs have been made with Catia V5R21 and the designs have been simulated with Autodesk Simulation CFD 2014 software by Enover R&D Department. As is known efficiency of PV cells is about %15. Therefore about %85 of solar radiation can not be converted into electrical energy. So, except the reflected radiation almost %70 of total radiation causes temperature rise on solar cells. This situation causes efficiency losses (figure 1).

a) b)

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c) d) Figure 1. a) Impact of operating time to PV panels with different baseboards, b) Impact of solar radiation to PV panels with different baseboards, c) Impact of air temperature to PV panels with different baseboards and d) Impact of wind speed to PV panels with different baseboards [1]. In PVT systems, water is circulated on entire rear surface of the panels thereby hot water can be obtained while the panel is cooled [2]. Also circulating a working fluid with a high flow rate results a small increase in the electrical efficiency of PV panels. On the other hand, high flow rate will also increase the unwanted electricity cost to pump the fluid. To prevent pumping costs, operating the working fluids at low flow rate causes fluid’s temperature to increase, however PV panel temperature will be high as well [3]. But, especially in consequence of installation locations like remote areas, deserts, etc, locations and degree of solar insolation, cooling with water is not possible or even when it can be done very high power pumps are needed for long distance water circulation. In the case of cooling with heat pipes, heat absorbed on solar cells will be transferred to air by natural convection or when it is beneficial such as using in houses, hotels, malls, etc., transferred to domestic water by forced convection. In this way the aim is reducing or resolving head losses so the pumping costs. In order to reduce margin of errors of CFD analysis of further designs, firstly CFD analysis of a standard PV panel without heat pipes will be held. The next step will be comparing the further designs’ CFD analysis with the original PV’s CFD results. Then, deciding the optimum design and product. Due to these simulations, which will be held on PCs, the effectiveness of the designs can be observed before prototyping and physical testing. Advantages of CFD analysis;

• Low cost; in most engineering Works, computer-based numerical analyses are much less cost than physical experimental studies.

• Time saving; through numerical analyses companies find the opportunity to study on many different designs and configurations in a very short time rather than the physical experiments for all designs.

• Ability to obtain full information; computer-based analysis of a problem will provide designers much more detailed and complete information than the physical tests. Thus major increase of performance can be obtained with minor revisions.

• Realistic conditions; realistic boundary and initial conditions, ambient conditions, material properties etc. can be easily simulate in numerical calculations.

Main equations, Reynold’s Transport Theorem,

. . ∙

. . ∙

. . (1)

249

Continuity equation;

(2)

Conservation of momentum; X – momentum,

(3)

Y – momentum,

(4)

Z – momentum,

(5)

Conservation of energy;

(6)

THERMOSYPHONS and HEAT PIPES The heat pipe is a device of very high thermal conductance. The first heat pipe idea suggested by Gaugler in 1942. However, until Grover’s heat pipe in the early 1960s it had remained as only a theory. George Grover’s heat pipe’s capabilities have driven attention caused significant developments. The heat pipe is similar in some respects to the thermosyphon. The thermosyphon is shown in Figure 2a. A small quantity of water is placed in a tube from which the air is then evacuated and the tube sealed. The lower end of the tube is heated causing the

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liquid to vaporize and the vapor to move to the cold end of the tube where it is condensed. The condensate is returned to the hot end by gravity. Since the latent heat of evaporation is large, considerable quantities of heat can be transported with a very small temperature difference from end to end. Thus, the structure will also have a high effective thermal conductance. The thermosyphon has been used for many years and various working fluids have been employed. One limitation of the basic thermosyphon is that in order for the condensate to be returned to the evaporator region by gravitational force, the latter must be situated at the lowest point. The basic heat pipe differs from the thermosyphon in that a wick, constructed for example from a few layers of fine gauze, is fixed to the inside surface and capillary forces return the condensate to the evaporator. (Figure 2b.) In the heat pipe the evaporator position is not restricted and it may be used in any orientation. If the heat pipe evaporator happens to be in the lowest position, gravitational forces will assist the capillary forces. The term ‘heat pipe’ is also used to describe high thermal conductance devices in which the condensate return is achieved by other means, for example centripetal force, osmosis or electrohydrodynamics [4].

a) Thermosyphon b) Heat pipe Figure 2. a) Thermosyphon, b) Heat Pipe To see how effective a thermosyphon is experimentally, it is possible to use an electric towel heater. Heat transfer rate difference between water and a thermosyphon system can be seen in Figure 3. With a 500W electrical heat source in ten minutes temperature distribution on a thermosyphon system is around 45oC higher than the water filled towel heater.

Min. 1 Min. 2 Min. 3 Min. 4 Min. 5

Min. 6 Min. 7 Min. 8 Min. 9 Min. 10 Figure 3. Temperature distribution of a water filled system and a thermosyphon

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MODELING of a PV PANEL In order to make a conclusion on a cooler design analysis, as a first step it is a good idea to analyze a PV panel without a cooling system with the same boundary conditions. To model a PV panel, firstly, we need to know about the panel’s materials. As shown in the Table 1 and Figure 4 a standard PV panel’s first layer is a glass cover with the thickness of 3.2 mm, and then solar cells with a thickness of 0.2 mm. The final layer of it is coating called dymat with a thickness of 0.35 mm. To keep all the layers together a special glue called as EVA is used between all the layers with the thickness of 0.45 mm. Table 1. PV Panel Metarials

Conductivity Density Specific Heat Thickness Low Iron Glass 0.937(W/mK) 2530(kg/m3) 0.21(j/g K) 3.2 mm EVA 0.35(W/mK) 0.95(g/cm3) 1400(j/kg K) 0.45 mm Silicon(Solar Cell) 0.1872(W/mmK) 0.002333(g/mm3) 0.66481(j/g K) 0.2 mm Dymat 0.23(W/mK) 1200(kg/m3) 1250(j/kg C) 0.35 mm

Figure 4. Layers of a PV panel Points to consider while making the CAD model of a PV panel are, firstly the dimensions a panel is 1960x950x4 mm. However, only a symmetrical slice of a PV panel with the dimensions of 1960x156x4 mm is enough for the CFD analysis. Thereafter, the PV slice is positioned into an air volume which dimensions are 2827,508 x4417,502 x156 mm with the angel of 45o. The reason of this angel is according to International Standard IEC 61215 Second edition 2005-04 during the performance tests of a PV panel, the panel must be placed with the angel of 45o between the panel and ground (Figure 5). All those conditions are same for all the designs.

Low iron glass EVA Solar Cell EVA Dymat

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Figure 5. Positioning of the CAD model CFD ANALYSIS of a PV PANEL Boundary conditions; Again according to IEC 61215, irradiance on the panel is 800W/m2. Assuming that except reflecting from front glass cover, energy production, %70 of this radiation is heat load on the solar cells. So that, the first boundary condition is 560W/m2 heat flux on the solar cells. As shown in Figure 6 number back (1) and top (2) sides of ambient air are open to atmosphere with the gage pressure of 0 Pa. During IEC 61215 performance tests of the PV panels air is blown with 1 m/s speed and 20oC temperature from in front of the panel. To apply that into CFD analysis 1 m/s air velocity is defined from the front (3). However to simulate hot air conditions its temperature is defined 35 oC for all the simulations.

Figure 6. Defining the boundary conditions

45◦

Air volume

1 P

gage

= 0

Pa

2 Pgage= 0 Pa

3 V

air=

1 m

/s

Tre

f= 3

5oC

253

Figure 7. Meshing

Figure 8. Velocity of air Figure 9. Traces Results; According to CFD analysis, solar cells’ mean temperature is around 70oC with 800W/m2 irradiance, 35oC ambient temperature and 1 m/s air flow (Figure 11). As shown in Figure 10, according to I-V curves for a standard PV panel with 70oC of cell temperature total power generation is around 148W. If cell temperature can be decreased to 55oC, power generation will be around 160W.

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Figure 10. I-V Curves

Max Temperature: 80,56◦C Min Temperature: 43,98◦C

Mean Temperature: 69,25 ◦C Figure 11. Temperature distribution on solar cells

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Version 1, First Design of PV Cooler In the first design of the cooler, cooling PV panel with natural convection has been tried. However, no matter supporting heat transfer area with aluminum fins and heat pipes (figure 12 - 13), cell temperature cannot be reduced by natural convection.

For CFD simulations, all the PV materials have been defined as shown in Figure 4 and the analysis has been solved with the same boundary conditions as in Figure 6. The heat pipe has been defined as a solid material with the conduction of 11000W/mK. Results,

Figure 12. First version of the cooler

Figure 13. Fins and heat pipes of version 1

Heat pipe

Aluminum fins

Aluminum fins

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Figure 14. Meshing of version 1

Figure 15. Velocity distribution of version 1 Figure 16. Traces of version 1

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Max Temperature: 96,32◦C Min Temperature: 60,11◦C

Mean Temperature: 77,36 ◦C Figure 17. Temperature distribution on solar cells Version 2, Cooling Heat Pipes With Water in a Manifold After failure of cooling with natural convection, a manifold system has been simulated and it has been observed that cooling with water is much more effective than air. On this design heat pipes were fixed underneath the panel and tied up to two different manifolds on is at the top of the panel and the other is at the middle of it as seen in Figure 18. The reason of designing two manifolds is in real life making heat pipe too long reduces the effectiveness of them.

Figure 18. Version 2, cooling with water

Manifolds

Heat Pipes

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For the simulation all the materials and boundary conditions are as same as the previous analyses as well as the heat pipe material. On this design manifolds are condenser of the heat pipes. There is a water circulating pipe inside of the manifold and condensation occurs on the external wall of the pipe. As boundary conditions, water inlet velocity has been defined as 100 mm/s and water temperature has been defined as 35oC same as ambient as seen in Figure 18 and Figure 19. On later designs water pipe number has been increased.

Figure 19. Manifold design of version 2 and boundary conditions Results,

Figure 20. Meshing of version 2

Water inlet V=100 mm/s T=35 oC

Heat Pipe

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Max Temperature: 67,14◦C Min Temperature: 35,12◦C

Mean Temperature: 51,23◦C Figure 21. Version 2, Temperature distribution on solar cells As seen on CFD results, with circulating water in the manifolds, the mean temperature of the cells has been reduced to 51oC which means that there will be less efficiency drops due to heat load so that power generation increases by about 10%. Version 3 As further step, on the next design heat pipe number per a slice has been increased from one to two, also manifold have been designed as including two parallel water pipes as shown in Figure 22 and Figure 23.

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Figure 23. Version 3 manifold

Figure 22. Version 3 As a result of increasing heat pipe numbers, cooling performance off the design has increased significantly. As seen in Figure 24 mean temperature decreases from 51oC to 43oC. This result shows that cooling with heat pipes provide almost 12% increase of power generation. On the other hand, to cool down the heat pipes on versions 2 and 3 cooling water must be circulated inside the manifolds. That causes unwanted electricity consumption due to pumping. But head lose inside the manifolds will be less than circulating the water underneath the panels as PV/T systems. Disadvantage of water cooling designs version 2 and 3 is in large solar fields it is not possible to circulate water in all the PV manifolds due to pumping costs and lack of water in deserts. The solution for large solar fields seems that using heat storage tanks at the top of the panels without circulating any liquid as in Figure 25. More importantly using a phase change material (PCM) inside the tank makes possible to use latent heat to storage heat in day times and due to high temperature difference between days and nights the heat stored in day times can be transferred to ambient at nights. That transfer causes solidifying of PCM in the tank.

Manifolds

Heat Pipes

Water inlet V=100 mm/s T=35 oC

Heat Pipes

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Max Temperature: 56.66◦C Min Temperature: 35.1◦C

Mean Temperature: 43.33◦C Figure 24. Version 3, Temperature distribution on solar cells

Figure 25. A complete system scheme for PVT storage [5]

PCM

PCM Tank

Immersing Heat Pipes

Hot Water

To Hot Water d

PV

Attached Heat Pipes

Tc

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PVT

Temperature Peaking

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REFERENCES 1. Xingxing Zhang, Xudong Zhao, Jingchun Shen, Xi Hu, Xuezhi Liu and Jihuan Xu. 2013. Design, fabrication and

experimental study of a solar photovoltaic/loop-heat-pipe based heat pump system. Solar Energy 97 551–568 2. Mohd Yusof Othman, Adnan Ibrahim, Goh Li Jin, Mohd Hafidz Ruslan, Kamaruzzaman Sopian. 2013. Photovoltaic-thermal

(PV/T) technology The future energy technology. Renewable Energy 49 171-174 3. Mohd Nazari Abu Bakar, Mahmod Othman, Mahadzir Hj Din, Norain A. Manaf, Hasila Jarimi. 2014. Design concept and

mathematical model of a bi-fluid photovoltaic/thermal (PV/T) solar collector. Renewable Energy 67 153-164 4. David Reay, Peter Kew. 2006. Heat Pipes, Theory, Design and Applications Fifth Edition.

5. Kılkış B. and Kılkış Ş. Combined Heat and Power with Renewable Energy Systems, 370 Pages, (In Turkish) TTMD Publication No: 18 Doga Publications Inc. Istanbul, 2016

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[Abstract:0076][Economy of Energy and Environment] POTENTIAL FOR CO2 EMISSION REDUCTION AT SETTLEMENT SCALE THROUGH COST

OPTIMAL AND NEARLY ZERO ENERGY BUILDINGS

Ece Kalaycıoğlu1, Touraj Ashrafian1, Ayşe Zerrin Yılmaz1, Nazenin Moazzen1 1 Istanbul Technical University, Istanbul Turkey

Corresponding mail: yilmazzer@itu.edu.tr

SUMMARY The European Directive on the Energy Performance of Buildings (EPBD Recast, Directive 2010/31/EU), introduced the "nearly zero energy level" and “cost optimality” concepts in order to take also economic effects into account while analyzing the energy performances of the buildings with different measures. In this paper the studies to define the cost optimal and nearly zero energy levels for different building typologies in Turkey for an Ecological Settlement Design have been introduced. The study covers defining reference buildings for five different functions, as residential, office, commercial, industrial and educational buildings. Nine buildings have been studied as case study buildings considering five different residential types, one office, one school, one shopping mall and one industrial building. Reference buildings have been defined for these case study buildings and energy efficiency improvement measures and packages have been applied to achieve the cost optimal levels and nearly zero energy levels. Global costs and investment costs were calculated for every improvement measure and package. Building based CO2 emission reductions have been calculated for a settlement which has been designed as ecological city and it has been concluded that building originated CO2 reduction level can reach to 66% in the settlement even with individual HVAC systems for each building. Besides individual HVAC Systems the effect of district energy systems on CO2 emission reduction has also been studied. INTRODUCTION As it is known building energy efficiency is the key problem of building sector in all over the world since buildings are consuming more than 40% energy of the world and therefore they are responsible from the 1/3 of CO2 emission. Therefore, together with the EU Countries Turkey has to follow the building energy efficiency directives. The new European Directive on the Energy Performance of Buildings (EPBD Recast, Directive 2010/31/EU), introduced the "nearly zero energy level" and “cost optimality” concepts in order to take also economic effects into account while analyzing the energy performances of the buildings with different measures [1,2,3]. Cost optimum energy efficiency level is the energy efficiency level which is providing lowest global cost during the economic life cycle of the building. Global cost is mainly the sum of construction cost, maintenance cost and energy cost of the building. Nearly zero energy building is the building with very low energy demand and this demand is provided by renewable energy as much as possible. In order to close the gap between cost optimum and nearly zero energy level it is crucial to find feasible solutions for renewable applications to the buildings. The studies are going on in Turkey to define the cost optimum energy efficiency level and nearly zero energy level under national conditions [4]. TUBITAK Project with the number of 113M596 is the adaptation of the EPBD Methodology to Turkey and definition of the reference buildings starting from residential types. In the following Figure two of selected reference buildings and ID form for one of reference buildings are given as sample.

Figure 1. Sample reference buildings and ID form [4] In this paper the results of energy efficiency studies for a new settlement which is Kocakır reserved area for urban transformation in Eskişehir. For

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energy efficiency in buildings nine building typologies have been studied to define reference buildings, energy efficiency improvements for cost optimum energy efficiency and nearly zero energy levels. The selected building typologies are single family house, attached and detached apartment buildings with three and five stories, school building, office building, industrial building and shopping mall. ENEGY EFFICIENCY FOR BUILDINGS IN CASE STUDY SETTLEMENT Reference values for the selected reference buildings have been defined for Eskişehir conditions basing on statistical values and the existing building and system standards [5,6,7]. Then the energy consumption level has been calculated for each reference building through dynamic simulation tool named Energy Plus. The single measures and packages for energy efficiency improvements have been applied to the reference buildings to calculate energy consumption and global cost for each case. Energy consumption is primary energy consumption and global cost is the sum of investment cost, maintenance cost and energy cost for the estimated economic lifecycle of building. The following Figures 2 and 3 are given as sample for the office buildings to show the energy consumption levels of different energy efficiency measures and packages and global cost of them.

Figure 2. Primary energy consumption for different improvements in Eskişehir office building

Figure 3. Primary energy consumption and global cost for Eskişehir office building

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Primary energy consumption, CO2 emission and global cost of reference buildings, cost optimum buildings and nearly zero energy buildings are given in Table1 for the selected building typologies under Eskişehir conditions. As it can be seen in this table it is possible to achieve 35% and 66% carbon emission reduction from the buildings if the buildings are designed to provide the cost optimum level and nearly zero energy level respectively. Case study area is Kocakır reserve area in Eskişehir and the estimated total building area in the settlement will be around 3.350.000 m2. Considering this total building area it is possible to provide 37.000 Ton and 68.000 Ton CO2 reduction in a year from the settlement with the cost optimum level and nearly zero energy level buildings respectively. Table 1. Primary energy consumption, CO2 emission and global cost of case study buildings

Single family house

Primary energy consumption, kWh/m2-y

Reference building 214,7 Cost optimum building 109,4 Nearly zero energy build. 32,7

Carbon emission, Kg CO2/m2-y

Reference building 52,12 Cost optimum building 27,18 Nearly zero energy build. 7,71

Global cost, TL/m2

Reference building 1098 Cost optimum building 991 Nearly zero energy build. 1090

Attached apartment

Primary energy consumption, kWh/m2-y

Reference building 118,14 Cost optimum building 83,05 Nearly zero energy build. 37,46

Carbon emission, Kg CO2/m2-y

Reference building 29,65 Cost optimum building 20,60 Nearly zero energy build. 9,39

Global cost, TL/m2

Reference building 946 Cost optimum building 867 Nearly zero energy build. 922

Detached apartment

Primary energy consumption, kWh/m2-y

Reference building 106,55 Cost optimum building 55,55 Nearly zero energy build. 27,62

Carbon emission, Kg CO2/m2-y

Reference building 26,12 Cost optimum building 13,77 Nearly zero energy build. 6,67

Global cost, TL/m2

Reference building 913 Cost optimum building 831 Nearly zero energy build. 856

Office building

Primary energy consumption, kWh/m2-y

Reference building 172,65 Cost optimum building 96,65 Nearly zero energy build. 72,45

Carbon emission, Kg CO2/m2-y

Reference building 44,39 Cost optimum building 24,89 Nearly zero energy build. 18,57

Global cost, TL/m2

Reference building 1043 Cost optimum building 877 Nearly zero energy build. 928

School building

Primary energy consumption, kWh/m2-y

Reference building 92,31 Cost optimum building 73,16 Nearly zero energy build. 25,00

Carbon emission, Kg CO2/m2-y

Reference building 23,68 Cost optimum building 18,61 Nearly zero energy build. 6,01

Global cost, TL/m2

Reference building 919 Cost optimum building 884 Nearly zero energy build. 898

Industrial building Primary energy consumption, kWh/m2-y

Reference building 286,70 Cost optimum building 229,81 Nearly zero energy build. 180,40

Carbon emission, Reference building 73,76

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Kg CO2/m2-y Cost optimum building 59,56 Nearly zero energy build. 46,64

Global cost, TL/m2

Reference building 1202 Cost optimum building 1091 Nearly zero energy build. 1124

Shopping mall

Primary energy consumption, kWh/m2-y

Reference building 369,45 Cost optimum building 181,65 Nearly zero energy build. 94,41

Carbon emission, Kg CO2/m2-y

Reference building 94,93 Cost optimum building 46,77 Nearly zero energy build. 23,96

Global cost, TL/m2

Reference building 1870 Cost optimum building 1575 Nearly zero energy build. 1595

DISTRICT ENERGY SYSTEMS FOR THE CASE STUDY SETTLEMET AREA Meeting the all energy demand by renewable systems can’t be possible and/or feasible for single building since the proper application area may not be big enough to install the systems. Therefore individual applications of building integrated renewable energy systems usually have longer payback time. On the other hand the efficiency is limited for the building based energy system such as individual HVAC system, because the heat rejection during the energy production is usually lost. Therefore central renewable energy systems and district heating/cooling systems to feed the buildings in a settlement are more efficient to reduce the carbon emission, as combined heat and power systems are utilized and losses of the distribution systems are minimized. In this study, for the district energy system analyses, all the buildings summarized above are assumed to be sited in the case study settlement area. In the Table 2, building numbers and annual heating, cooling and electricity demands are summarized. The energy demand values are derived from nearly zero energy building results. For the district energy system calculations, individual buildings energy generation units are excluded and electricity, hot water for heating and chilled water for cooling are assumed to be met by district systems.

Table 2. Buildings Annual Energy Demands in the Settlement

Building Type Total Area in the

Settlement (m2)

Annual Electricity Demand (kWh/year)

Annual Heating Demand

(kWh/year)

Annual Cooling Demand

(kWh/year) Single Family House 29,920 944,354.43 540,162.00 - Attached Apartment 1,572,142 83,796,921.21 35,009,857.60 - Detached Apartment 647,671 25,579,722.65 22,721,949.03 - Office/Commercial Buildings 557,464 22,545,077.15 7,687,735.57 14,235,493.17

Industrial Building 80,000 27,347,498.56 1,578,083.42 4,623,545.64 School Building 131,912 2,592,065.30 2,197,699.00 1,139,065.42 Shopping Mall 64,118 4,151,645.03 1,653,044.74 2,287,901.33 TOTAL 166,957,293.34 71,388,531.36 22,286,005.55

As schematically shown in the Figure 4, in the case study settlement area, for district heating 140MW CHP system with %50 heating efficiency and %35 electricity efficiency is utilized. For district cooling, 28 MW chiller system with 3 COP is used. For electricity, beside of the electricity generation of CHP, PV panels of totally 46 MWp on building roofs and a separate PV panels on the site area are used. Network losses for hot and chilled water are assumed to be %15 of the total energy distributed.

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Figure 4. District Energy System Schematic Design According to the hourly simulation results shown in Table 3, CHP and PV systems together can meet the whole electricity demand of the settlement. However, in hourly basis, on some hours CHP and PV system cannot meet the electricity demand (Figure 5), as on some other hours more electricity than demanded is produced. For that reason, in total 7,234.00 MWh-y and %4 percent of the electricity is met from the grid.

Table 3. Annual District Electricity Production Annual Electricity

Production (MWh-y) Percentage of Meeting

the Demand CHP System 58,424.90 %35 PV Panels 109,817.30 %66

Figure 5. Hourly Electricity Demand and Electricity Production of CHP and PV Panels for a week period. In the case study settlement to meet the annual heating demand, 166,928.20 MWh-y energy and 14,515,496.60 m3-y natural gas is utilized. In the Table 4, the CO2 emissions of whole settlement area can be compared for stand-alone heating and cooling systems and district heating and cooling systems. For stand-alone energy systems, electricity grid losses are also added to the calculations as for district energy systems distribution losses are already included.

Table 4. Stand-Alone Energy Systems and District Energy Systems CO2 Emissions Stand-Alone Systems / Grid Energy CO2 Emissions (kg

CO2 -y)

District Energy Systems CO2 Emissions (kg CO2 -y)

Improvement Percentage

68,339,547.78 49,594,770.88 %27 CO2 emission for district energy systems is basically the reason of natural gas consumption. In the case study settlement area, energy acquisition potential from organic wastes and waste water recycle is also calculated and annually 10,525.0 MWh energy production potential is expected. Additionally, in the case study area cultivation of proper trees is planned to use wood chips as energy source instead of natural gas. Generally energy content of the wood chips can be assumed as 800 kWh/m3, thus annually 195,504 m3 wood chips should be used to meet all the energy demand and to zero natural gas based CO2 emissions.

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CONCLUSION In this paper the results of energy efficiency studies for different building typologies and the effect of district energy systems in Eskişehir Kocakır case study area have been summarized. As it can be seen from the results it is possible to achieve very big amount of CO2 emission reduction through cost optimal level and nearly zero energy level buildings with individual energy systems. But especially for a new settlement or transformation for an urban quarter it is strongly recommended to reduce the energy demand as much as possible through correct design scenarios and then to provide energy demand from central renewable energy systems and district energy systems, such as combined heat and power systems, to increase the efficiency and consequently to reduce CO2 emission. Eskişekir Kocakır case study shows that the buildings’ energy demand in the whole district can be provided from the sources of the district without any imported energy. REFERENCES

1. Energy Performance of Buildings Directive, EPBD recast, Directive 2010/31/EU of the European Parliament and of

Council, Official Journal of the European Union, L153/13-35, 2010. 2. CABAU, E., The EU Commission’s upcoming proposal on the cost optimal framework methodology, International

seminar on EPBD, Milano, 2011. 3. COMMISSION DELEGATED REGULATION (EU) No 244/2012, Supplementing Directive 2010/31/EU of the European

Parliament and of the Council on the energy performance of buildings by establishing a comparative methodology framework for calculating cost-optimal levels of minimum energy performance requirements for buildings and building elements, Official Journal of the European Union, L81/18-36, 2012.

4. Binalarda Maliyet Optimum Enerji Verimliliği Seviyesi İçin Türkiye Koşullarına Uygun Yöntemin ve Referans Binaların Belirlenmesi, TÜBİTAK 113M596, 2016.

5. TÜRK STANDARDLARI ENSTİTÜSÜ, TS 825: Binalarda Isı Yalıtım Kuralları, T.C. Resmi Gazete, 26979, Ankara, 2008.

6. T. C. BAYINDIRLIK ve İSKÂN BAKANLIĞI, Binalarda Isı Yalıtımı Yönetmeliği, T.C. Resmi Gazete, 27019, Ankara, 2008.

7. www.tuik.gov.tr

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[Abstract:0077][Comfort Cooling]

COMPARISON OF VACUUM COOLING AND CONVENTIONAL COOLING

*Hande Mutlu OZTURK, **Günnur KOÇAR and ***Harun Kemal OZTURK *Pamukkale University, School of Tourism and Hotel Management, Denizli, Turkey

**Ege University, Institute of Solar Energy Dept. of Energy Tech., İzmir, Turkey ***Pamukkale University, Faculty of Engineering Mech. Engineering Dept., Denizli, Turkey

Abstract— Vacuum cooling is known as a rapid evaporative cooling technique for any porous product which has free water.

This technique is used for pre-cooling of leafy vegetables and mushroom, bakery, fishery, sauces, cooked food and particulate foods. Vacuum cooling has some advantages such as short processing time, extension of product shelf life and improvement of product quality and safety. The aim of this paper is to apply vacuum cooling technique for the cooling of the cooked potato and show the pressure effect on the cooling time and temperature decrease. The results of vacuum cooling are also compared with conventional cooling (cooling in refrigerator) for the different temperatures. It has been seen that cooked potato can be cooled much more rapidly and efficiently when it is compared with conventional cooling.

Keywords— Conventional cooling, cooked potato, parsley, leek, pressure, vacuum cooling.

INTRODUCTION Vacuum cooling is a widely used rapid cooling method, which has been proven to be one of the most efficient cooling methods available and therefore, it is extensively used for cooling some agricultural and food products [1–3]. Vacuum cooling mainly depends on latent heat of evaporation to remove the sensible heat of cooled products and the quantity of the heat removed from the product is directly related to the amount of water evaporated of the product surface. Thus vacuum cooling method can be considered a rapid and evaporative cooling method. Generally, vacuum cooling can be applied for any porous product which has free water [4-8]. Vacuum cooling can be considered one of the most effective cooling methods to cool fresh fruit, vegetables, cut flowers, meat production, fish and sauces [9-11]. The main components of a typical vacuum cooler are vacuum chamber, vacuum pump and vapor condenser. Vacuum is provided in the vacuum chamber with the vacuum pumps and vapour–condenser. The function of the vacuum chamber is to keep the products to be cooled with vacuum cooling. When the vacuum pump start to run and vacuum established, the pressure inside the chamber is reduced to the saturation pressure corresponding to the initial temperature of the product, therefore some water boils away from the food until new equilibrium condition is achieved. Vacuum cooling causes to evaporate the large amount of vapour in from the food in the chamber. This vapour evacuated from the chamber is removed by the vacuum pump and/or through condensation when a vapour condenser is installed inside the chamber. Vacuum cooling can be applied any food product including free water and whose structure will not be damaged by water removal from the product [8]. Cooling occurs due to the evaporation of water from the food. When water evaporates, it needs to absorb heat in order to maintain higher energy level of molecular movement at gaseous state. The amount of heat required is called latent heat, which must be supplied from the product or from the surroundings that consequently are refrigerated. The temperature at which water starts to evaporate is directly dependent on the surrounding vapour pressure. Vacuum cooling is one of the most effective rapid cooling methods for providing all of these. The main requirements for using the vacuum cooling are: (a) the product should have a large surface area for mass transfer, (b) product water loss should not represent an economic or sensory problem, due to weight reduction and possible changes in structure or appearance [6]. Any porous food can be cooled with vacuum cooling because the water vapour generated within the sample easily diffuse to the surrounding atmosphere. The heat and moisture transfer is a complicated process and therefore it has been investigated by many researchers. Vacuum cooling is extensively used for cooling some agricultural and food products [12-14]. It has been widely applied in pre-cooling treatment for lettuce ([11], [15-20]), cut flowers [21], mushrooms ([22]). There are several advantages of vacuum cooling. First of all, foods can be cooled in extremely short period. The difference between the cooling rate of vacuum cooling and conventional cooling is due to different cooling mechanisms involved.

Vacuum cooling process Vacuum cooling is based on the rapid evaporation of moisture from the surface and within of the products due to the low surrounding pressure. Water evaporation directly depends on the surrounding pressure and the decreased pressure causes the temperature decrease. Water evaporates at 100 C at atmospheric pressure of 101.3 kPa (1 atm), however, water starts to *Hande MUTLU OZTURK is with the Pamukkale University, Tourism Faculty, Denizli, 20070 Turkey (e-mail: hmozturk@pau.edu.tr.com)

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evaporate at lower temperature (below 100 C) when the pressure is decreased to below atmospheric pressure. When any free water containing product is placed in a closed chamber and the pressure is decreased with a vacuum pump to below the atmospheric pressure, due to the pressure difference between the water in the product and the surrounding will cause evaporation and the vapor move from the product to the surrounding atmosphere. Heat removed from the product during this process will be equal to the latent heat required for evaporation. As a result, product temperature starts to decrease with decreasing of the pressure and cooling is achieved. In order to remove the large amount of water vapor and keep the cooling cycle within a reasonable length of time, the vapor-condenser is used to economically and practically handle the large volume of water vapor by condensing back to water and then draining it through the drain valve. For maintaining steady cooling process, it is necessary to evacuate the chamber continuously. Desired final temperature of the product can be controlled adjusting the final surrounding pressure. Process of a vacuum cooling can be given as follow: Vacuum chamber is used to keep the food products. After placing the food into the vacuum chamber, the door is closed and the vacuum pump is switched on. When the pressure is reduced and water starts to evaporate, the food temperature begins to decrease. Cooling of the food continues until it reaches to the desired product temperature. When the determined temperature is achieved, the pump is stopped, the ventilation valve is opened and atmospheric air is allowed to enter into the chamber. After the process is finished, finally, the cooled products are removed from the chamber.

2.1. Theoretical Approach

In this section, a simple theoretical analysis of vacuum cooling process based on thermodynamic principles is presented. This analysis is limited to the mass loss based on temperature drop observed during vacuum cooling process. Average Specific Heat (Cavg) of any vegetable can be calculated by the following expression:

Cavg=3349a+837.36 (J/kgK) (1)

where a is the water content. For instance, water content of cooked potato is about 80% by mass. Therefore, specific heat of cooked potato is:

Potato=3516,56 J/kgK (2)

Then the heat required to lower the temperature of a 1 kg cooked potato from 100°C to 5°C could be calculated by the following expression:

Q=mCT (3)

For the other vegetables leek and parsley same equations could be used for calculating Average Specific Heat (Cavg) and heat requirement.

2.2. Energy Analysis

In following two sections, energy analysis of the conventional and vacuum cooling system will be given. For the both vacuum cooling and conventional refrigeration system, as an example cooked potato will be given. The calculation will be carried out for cooked potato so that the temperature of cooked potato is desired to decrease from about 100 C (cook temperature) to 5 C (storage temperature).

2.3. Energy Analysis of Conventional Refrigeration System

For the conventional refrigeration system, it is considered that the cooked potato is placed in an isolated and cooled area. The refrigerated area is considered at 5C and cooked potato at 100C temperature is placed into the refrigerator. Flowing equation can be used to calculate the heat removed from the cooked potato to cool from 100C to 5C:

potato potatop TCmTCmQ otatoairvairL (4)

If temperature of the refrigerator is considered at 5C, first term at the right side of the equation can be neglected, and in this case, the amount of heat is needed to remove from refrigerator will be as;

potato potatopotato TCmQL (5)

From Equation (2), in order to cool 1 kg cooked potato from 100C to 5C, removed heat from cooked potato will be;

07.334LQ kJ (6)

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If the process is considered reversible and the cycle is considered as Carnot refrigerator, the input work will be

LH QQW (7)

For the Carnot cycle, QH can be calculated from following equation;

22.442278)95273(07,334 LHLH TTQQ (8)

Work can be found as;

15.10807.33422.442 revW kJ (9)

2.4. Energy Analysis of Vacuum Cooling System

As expressed earlier, for the vacuum cooling, the pressure of the cooling area needs to be decreased. For the vacuum cooling, pressure decreased from atmospheric pressure to the vacuum pressure at the constant volume. In this case, reversible work will calculate with;

VdPWrev (10)

If the vacuum chamber volume is taken as 29 l (0.029 m3) and the atmospheric pressure is considered as 101.3 kPa,

93.23.101029.0 revW kJ (11)

It can be seen easily that vacuum cooling is consume about 37 times less energy than conventional refrigeration system for this example. It also should be noted that the energy consumption for the conventional refrigeration systems increase linearly with the weight of the products. However, the energy consumption for the vacuum cooling does not change with the amount of the product.

Materials and Methods 3.1. Vacuum Cooling System, Measurements and Data Collection

The basic components of a vacuum cooling system used in this study are a vacuum chamber, vacuum pump and vapour condenser (heat exchanger). The experimental apparatus is presented in Fig. 1. The vacuum chamber (Memmert VO-200, Schwabach, Germany) is used to keep the food product and cooled in. The rotary vane vacuum pump is used to generate vacuum (Edward, RV8, New Jersey, USA) with 1.5x10-3 mmHg (2x10-3 mbar). Pumping speed of 8.5 m3/h and rotary speed 1800 rpm is used to evacuate the air in the vacuum chamber and the vapour evaporated from the products from atmospheric pressure to the given vacuum pressure. Since a great amount of vapour is generated during vacuum cooling, vapour condenser is installed between the vacuum chamber and vacuum pumps in order to condense the vapour back to water to be discharged through the water evacuation. Variation of surface and center temperature of the products determined with two calibrated thermocouples (Highly accurate immersion/penetration probe, precision of ±0.01C, TESTO, Lenzkirch, Germany). Thermocouples are inserted into the samples and connected to the data logger (TESTO 350 M/XL-450, Lenzkirch, Germany). Humidity and temperature (highly accurate reference humidity/temperature probe, precision of ± 1 % and ±0.4 C, TESTO, Lenzkirch, Germany) of vacuum chamber have been measured with the same probe and data are recorded to the data logger in the vacuum chamber. Pressure (low pressure probe, TESTO, Lenzkirch, Germany, the accuracy of ±0.1%) has been measured from the pipe between the vacuum pumps and vacuum chamber and data are recorded to the control unit (TESTO 350 M/XL-450, Lenzkirch, Germany). Vapour evacuated from the vacuum chamber is condensed in the heat exchanger via coolant (POLYSCIENCE 9506, Niles, Illinois, USA). Center and surface temperature of the iceberg lettuce, ambient temperature and humidity of vacuum chamber have been measured and recorded with the data logger in the vacuum chamber (see Figure 1). On the other hand, control unit that records the pressure data is located outside the vacuum chamber (see Figure 1). The data logger measures and saves readings without any connection to the control unit in the vacuum chamber for each 10 seconds. The control unit displays the measurement data and controls both the data logger and control unit. Pressure data also recorded for each 10 seconds. Measured data are transferred from control unit and data logger to the computer using ComSoft3 Software (TESTO, Lenzkirch, Germany). Air filter is used to grasp the dirt before the pumps. Before cooling, vacuum pump was warmed up for half an hour so that the system was stable. Experiments carried out for three different pressures (0.7 kPa, 1 kPa and 1.5 kPa) and three replicates were performed for each pressure and average data were used.

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Figure 1. Schematic diagram of the vacuum cooler system.

3.2. Conventional Cooling

Conventional cooling was carried out in a no-froze refrigerator (Beko D 9470 NE, Gebze, Kocaeli, Turkey). The average temperature of the refrigerator was set to 6C, -16C and -20C and the cooked potato were placed in the refrigerator. The experiments were ended center temperature of cooked reach to 6 C. Three replicates were performed for each experiment.

3.3. Weight Measurement

The weights of the cooked potatoes before and after the cooling process are determined with an electronic balance (Precisa XT 1220 M). The weight difference is the mass loss during the vacuum cooling process. The accuracy of the balance is ±0.001 g.

3.4. Thermal View

Before and after the vacuum cooling of the products, thermal views (Flir Systems, Danderyd, Switzerland) have been taken and views are transferred from thermal camera to the computer by using ThermaCAM QuickView (Flir Systems, Danderyd, Switzerland).

3.5. Plant Material

Freshly harvested potato was bought from a local distribution centre and transported to the Pamukkale University-Clean Energy Center, Denizli, Turkey. Experiments have been carried out for fresh potato. Potato was weighted before and after the cooking and after the cooling.

Results and Discussion The aim of this study is to determine the effect of the pressure on the vacuum cooling of parsley, cooked potato and leek and comparison of the results with conventional cooling. In order to determine the mass loss and mass loss ratio before and after the vacuum cooling, the weights of the parsley, cooked potato and leek have been taken. In order to determine the variation of the center and surface temperature of the parsley, cooked potato and leek, the vacuum chamber humidity and temperature, variation of pressure during the vacuum cooling are measured for three different set pressure 0.7 kPa and 1 kPa. In Figure 2, 4, 6, 8, 10 and 12, the pressure, humidity and temperature variation for the two different pressures (0.7 kPa and 1 kPa) for the parsley, cooked potato and leek are given. As can be seen from the Figures, vacuum chamber temperature has not been changed during cooling period, and it is nearly equal to ambient temperature or initial temperature for parsley, cooked potato and leek. Since cooling effect for vacuum cooling directly comes from water evaporation from the cooled product, almost no temperature change occurs at the ambient (temperature in the vacuum chamber). However, vacuum chamber humidity fluctuates through the process as can be seen in the Figures.

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It can be seen from the Figures 2, 4, 6, 8, 10 and 12 that vacuum pressure in the vacuum chamber decreased rapidly from atmosphere to about 2 kPa in 200 second (3.33 min), then decline slightly. When it reaches to set pressure it keeps almost constant value. When the pressure is lower or equal to the saturated pressure at the local temperature, water starts to boil in the food and the evaporation of the water causes to the cooling. For three different set pressures (0.7 kPa and 1 kPa), the surface temperature is lower than the centre temperature as expected. The cooling effect comes from water boiling from the samples, and therefore evaporation and cooling of the sample starts from the surface. Therefore, the surface temperature of the samples for the different pressures lowers than the centre temperature. However, with decreasing the pressure, evaporation and cooling occurs through the parsley, cooked potato and leek and temperature decrease together. The total cooling time is dependent on the type and shape of the products, porosity, pore size, surface are and the pore distribution within the samples, and the availability of free water in the pores, and set pressure. Temperature of cooked potatoes desired to decrease from about 95 C (cooking temperature) to 5 C (storage temperature), on the other hands, for the parsley and leeks, the temperature decrease from the 25 °C ambient temperature to 5 C storage temperature. When cooling time of parsley, cooked potato and leek compared with each other for both 0.7 kPa and 1 kPa, it can be seen the short cooling time recorded for cooked potato as it is not expected. It is because of the fact that although the parsley and leek cooled from the temperature of the 25 C while cooked potato cooled from the 95 C, the free water contents of the parsley and leek lower than the cooked potato. Therefore cooked potato cooled in a short time. Figures 3, 5, 7, 9, 11 and 13 shows the thermal camera pictures of temperature for the parsley, cooked potato and leek. It can be seen that surface temperature of the food product higher than the center temperature. The reason is that the temperature of the products surface come effected by the ambient temperature when it is taken from the vacuum chamber to the environment. Weight loss occurs during vacuum cooling since cooling effect directly comes from water evaporation (boiling) from food products; parsley, cooked potato and leek. Weight losses of parsley, cooked potato and leek during vacuum cooling for two different pressures are given in Table 1, 2 and 3. Weight loss and the percentage weight loss are closely related to final set pressure. As shown in the Tables, cooling time depends on set pressure and for low pressure cooling time is shorter. Also, final temperature depends on set pressure and it is not possible to achieve 5 C storage temperatures for 1 kPa pressure fro all products. Conventional cooling was carried out in a refrigerator at the set temperature of 5C and -18C for parsley, cooked potato and leek. The results have shown that, surface temperatures cool much faster than the centre temperatures (see Figure 14, 18 and 22 for 5 °C and Figures 16, 20 and 24 for -18 °C). Also, temperature differences centre and surface get lower as the ambient temperature (temperature in the refrigerator) decreases. Figures 15, 17, 19, 21, 23 and 25 shows the thermal camera pictures of temperature for the parsley, cooked potato and leek for the 5 °C and -18 °C of conventional cooling. It can be seen that surface temperature of the food product change with the cooling temperature and surface temperature drop the very low temperature It can also be concluded that provides conventional cooling much slower than vacuum cooling. A comparison of vacuum cooling at 0.7 kPa pressure is about 55 times faster than conventional cooling (cooling time is 20730 s) with the ambient temperature of 5C for the cooked potato. Same results was found for other products as well. For the case, cooling with vacuum is about 10 times faster than conventional cooling (cooling time is 3930 s) with -18C. Cooling time and mass loss of parsley, cooked potato and leek have been given for the conventional cooling at the Table 1, 2 and 3. As can be seen from the Tables, mass loss is higher at the cooling of 5C than -18C. Mass loss ratio is higher for vacuum cooling than conventional cooling. However, cooling time for vacuum cooling is shorter than the conventional cooling.

Conclusions In this study, two different cooling methods have been tested; vacuum cooling and conventional cooling. Results show that the vacuum cooling is a rapid and efficient cooling method when it is compared with conventional cooling method. On the other hand, it has been noted that the mass loss is higher for vacuum cooling when it is compared with conventional cooling. It can be concluded that for the high vacuum pressure it is not possible to achieve desired storage temperature of 5C. The results also show that the temperature decrease of cooked potato at the surface and at the centre is decrease nearly together. However, for conventional cooling, the surface temperature of parsley, cooked potato and leek decrease faster that the centre temperature. Also, for the low temperature of -18C, surface of the parsley, cooked potato and leek are freezing which is not desired. This study confirmed that vacuum cooling is an efficient method and is suitable for cooling of cooked food such as potato.

ACKNOWLEDGEMENTS The authors are grateful to TUBITAK (Scientific and Technological Research Council of Turkey) for support of the project named developing a vacuum cooling system and application in the food industry (Project Number: 106 M 262).

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Figure 2. Variation of pressure, center and surface temperature of parsley, temperature and humidity of vacuum chamber with time for set pressure of 0.7 kPa.

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Figure 3. Temperature variation of parsley before and after vacuum cooling for 0.7 kPa.

Figure 4. Variation of pressure, center and surface temperature of parsley, temperature and humidity of vacuum chamber with time for set pressure of 1 kPa.

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Figure 5. Temperature variation of parsley before and after vacuum cooling for 1 kPa.

Figure 6. Variation of pressure, center and surface temperature of cooked potato, temperature and humidity of vacuum chamber with time for set pressure of 0.7 kPa.

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Figure 7. Temperature variation of cooked potatoes before and after vacuum cooling for 0.7 kPa.

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Figure 8. Variation of pressure, center and surface temperature of cooked potate, temperature and humidity of vacuum chamber

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Figure 9. Temperature variation of cooked potatoes before and after vacuum cooling for 1 kPa.

Figure 10. Variation of pressure, center and surface temperature of leek, temperature and humidity of vacuum chamber with time for set pressure of 0.7 kPa.

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Figure 11. Temperature variation of leek before and after vacuum cooling for 0.7 kPa.

Figure 12. Variation of pressure, center and surface temperature of leek, temperature and humidity of vacuum chamber with time for set pressure of 1 kPa.

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Figure 14. Variation of center and surface temperature of parsley with time for 5°C storage temperature.

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Figure 16. Variation of center and surface temperature of parsley with time for -18°C storage temperature.

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Figure 18. Variation of center and surface temperature of cooked potato with time for 5°C storage temperature.

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Figure 19. Temperature variation of cooked potato for conventional cooling at the temperature of 5 °C.

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Figure 20. Variation of center and surface temperature of cooked potato with time for -18°C storage temperature.

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Figure 21. Temperature variation of cooked potato for conventional cooling at the temperature of -18 °C.

Figure 22. Variation of center and surface temperature of leek with time for 5°C storage temperature

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Figure 23. Temperature variation of leek for conventional cooling at the temperature of 5 °C.

Figure 24. Variation of center and surface temperature of leek with time for -18°C storage temperature

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Figure 25. Temperature variation of leek for conventional cooling at the temperature of -18 °C.

Table 1.Variation of mass lost and mass ratio with pressure and temperature for parsley

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Table 2.Variation of mass lost and mass ratio with pressure and temperature for leek

Table 3.Variation of mass lost and mass ratio with pressure and temperature for cooked potato

Vacuum Pressure (kPa) Temperature (°C)

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120,718 98,846 57,015 47,950 g

Mass Before Cooling (g): 112,549 95,532 55,657 47,304 g

Mass After Cooling (g): 8,169 3,314 1,358 0,646 g

Mass Loss (g) 6,77 3,352 2,38 %1,34

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93,334 71,819 25,556 68,608

Mass Before Cooling (g): 90,885 69,317 25,302 68,098 Mass After Cooling (g): 2,449 2,502 0,254 0,510 Mass Loss (g) 2,62 3,48 0,99 0,74

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Mass Before Cooking (g): 154,094 160,656 159,842 126,658 Mass After Cooking (g):

154,132 162,800 162,600 128,101 Mass After Cooling (g):

135,245 144,209 158,440 126,997 Mass Loss (g) 18,887 18,590 4,16 1,104 Mass Loss Ratio (%) %12,25 % 12,89 %2,56 % 0,86

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REFERENCES [1] Brosnan T, Sun Da-Wen. Precooling technique and applications for horticultural products-a review. International Journal of Refrigeration 2001; 24(2):154–70. [2] Sun Da-Wen, Wang L.J. Heat transfer characteristics of cooked meats using different cooling methods. International Journal of Refrigeration 2000; 23 (7):508–16. [3] Wang L.J, Sun Da-Wen. Modelling vacuum cooling process of cooked meat-part 2: mass and heat transfer of cooked meat under vacuum pressure. International Journal of Refrigeration 2002; 25(7):861–2 [4] Wang, L. J., & Sun, D.-W.. Rapid cooling of porous and moisture foods by vacuum cooling technology. Tends in Food Science & Technology 2002; 15, 174–184. [5] Houska, M.; Podloucký, S.; Zitné, R.; Grée, R.; Sesták, J.; Dostàl, M.; Burfoot, D. Mathematical model of the vacuum cooling of liquids. Journal of Food Engineering 1996; 29, 339–348. [6] Mc Donald, K.; Sun, D.W. Vacuum cooling technology for the food processing industry: A Review. Journal of Food Engineering 2000; 45, 55–65. [7] Dostál, M.; Petera, K. Vacuum cooling of liquids: mathematical model. Journal of Food Engineering 2003; 61, 533–539. [8] Wang, L.; Sun, D.W. Rapid cooling of porous and moisture foods by using vacuum coolingtechnology. Food Science and Technology 2001; 12, 174–184. [9] Shewfelt RL, Phillips RD. Seven principles for better quality of refrigerated fruits and vegetables. In: Refrigeration science and technology proceedings. new developments in refrigeration for food safety and quality, Lexington (KY, USA). 1996, p. 231–6. [10] Sullivan GH, Davenport LR, Julian JW. Precooling: key factor for assuring quality in new fresh market vegetable crops. In: Janick, editor. Progress in new crops. Arlington: ASHS Press; 1996. p. 521–4. [11] Tambuna AF, Morishima H, Kawagoe Y. Measurement of evaporation coefficient of water during vacuumcooling of lettuce. In: Yano Nakamura, editor. Developments in food engineering. UK: Chapman and Hall; 1994. p. 328–30. [12] Thompson J; Rumsey T R. Determining product temperature in a vacuum cooler. ASAE Paper No: 84-6543 1984. [13] Sun D-W; Wang L J. Heat transfer characteristics of cooked meats using different cooling methods. International Journal of Refrigeration 2000, 23, 508-516. [14] McDonald K; Sun D-W. Vacuum cooling technology for the food industry: a review. Journal of Food Engineering 2000; 45(2), 55-65. [15] Shewfelt RL, Phillips RD. Seven principles for better quality of refrigerated fruits and vegetables. In: Refrigeration science and technology proceedings. new developments in refrigeration for food safety and quality, Lexington (KY, USA). 1996; 231–6. [16] Sullivan GH, Davenport LR, Julian JW. Precooling: key factor for assuring quality in new fresh market vegetable crops. In: Janick, editor. Progress in new crops. Arlington: ASHS Press1996; 521–4 [17] Tambuna AF, Morishima H, Kawagoe Y. Measurement of evaporation coefficient of water during vacuum cooling of lettuce. In: Yano Nakamura, editor. Developments in food engineering. UK: Chapman and Hall1994; 328–30. [18] Varszegi T. Vacuum cooling of vegetables. Hungarian Agricultural Engineering 1994; (7) 67–8. [19] Haas, E., & Gur, GFactors affecting the cooling rate of lettuce in vacuum cooling installations. International Journal of Refrigeration 1987; 10, 82–86. [20] Rennie, T. J., Raghavan, G. S. V., Vigneault, C., and Gariepy, Y. Vacuum cooling of lettuce with various rates of pressure reduction. Transactions of ASAE 2001; 44, 89–93. [21] Sun Da-Wen, Brosnan T. Extension of the vase life of cut daffodil flowers by rapid vacuum cooling. International Journal of Refrigeration 1999; 22: 472–8. [22] Frost CE, Burton KS, Atkey PT. A fresh look at cooling mushrooms. Mushroom Journal 1989; 193:23–9.

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[Abstract:0078][Renewable Energy Systems and Applications] ON THE USE OF A DIRECT SOLAR FLOOR FOR HEATING AND DOMESTIC HOT

WATER IN A RURAL HOUSE

1Farid Boudali Errebai*, 2Lotfi Derradji, 3Mohamed Amara Corresponding email: b-farid@hotmail.com

1, 2, 3BUILDING PHYSICS AND ENVIRONMENT DEVISION (DPBE), NATIONAL CENTRE FOR BUILDING INTEGRATED STUDIES AND RESEARCH (CNERIB), ALGIERS, ALGERIA.

SUMMARY

The solar energy potential in Algeria is very important, the insolation duration on almost the entire national territory exceeds 2000 hours annually and reaches 3900 hours in the south of Algeria and the daily received energy on a horizontal area of 1 m2 is 5 kWh over most of the Algeria, or around 1700 kWh/m2/year in the north to 2263 kWh/m2/year in the south, which represents a significant solar potential. This work aims to study the energy potential of Direct Solar Floor (DSF) for floor heating and hot water provider in an isolated rural house. The winter energy balance of the house was calculated by the Pleiades+Comfie dynamic simulation software. The results obtained showed that the energy needs for heating are 4709 kWh and 3720 kWh for domestic hot water. To be able to provide 70% of the energy needs, an area of 8 m² of solar panels was necessary in order to meet these needs. Nomenclature

cw Specific heat capacity of water, (Wh/kg.K) nsem,nor number of weeks of the month, holidays excluded, - npp number of draw-off points, - np number of puisages, - Nsubnor normal number of sub-periods of the month, - Qecs domestic hot water, (kWh) Qpecs Heat losses in the circuit of the hot water, (kWh) Qpch losses of heat in the heating circuit, (kWh) Tef cold water temperature, (°C) Tam room temperature (° C) (°C) Tec Average temperature of the hot water equal to 50 ° C (°C) Ti basic average indoor temperature [1], (°C) v volume of water equal to 1.2 l by draw-off point, (l) Vuw volume of mixed hot water used during the week, (L) ρw density of water. (kg/l)

INTRODUCTION

The solar energy potential in Algeria is very important, the sunshine duration of almost the entire national territory exceeds 2000 hours annually and can reach 3900 hours on the Hauts Plateaux and South. The energy received daily on a horizontal surface of 1 m2 is 5 kWh over most of the country, or around 1700 kWh/m2/year in the north to 2263 kWh/m2/year in the south of the Algeria, this represents a significant solar potential.

It is clear that this amount of energy which is in its primary form has to be converted into a useful form to use it. For solar energy, there are three main use: the passive use, photovoltaic use and the active solar use.

The active solar thermal energy use recovers heat from sunlight in a fluid, sometimes air, usually water, through the implementation of solar collectors.

The operation of a solar water heater is simple (figure 1). The fluid heated by the solar collectors flows directly into a slab floor, without going through a storage tank. The floor concrete mass ensures the energy storage function and phase-shift of its restitution in the heated volume. This low temperature heating installation does work with a better yield since all intermediate losses (heat exchangers) are eliminated.

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Since the development of the DSF by heating technology, several hundred installations have been carried out almost exclusively in France, mainly in detached houses and, in some cases, in the small buildings or hospitals. Measuring campaigns made on various houses helped to verify that the solar panels effectively contribute to the energy savings [2-5].

This article aims to present a case study of a detached rural house with low energy consumption, insulated with expanded polystyrene (EPS) and equipped with a direct solar floor. In this study, dynamic simulations have been done on this house to determine the required heating power and domestic hot water to determine the surface of flat solar panels in order to provide at least 70% of the thermal energy needed for heating and hot water.

Figure 1. Schematic of the Direct Solar Floor

METHODS

1. General description of the investigated house The investigated house has a living area of 80 m², it is located in the Algiers region, specifically in the village of Souidania. This region belongs to the climatic zone A which is characterized by a cool winter and a hot, humid summer (see figure 2).

The house contains two bedrooms, bedroom 1 oriented to the south-west and bedroom 2 which is oriented to the north-west, a living room with two large windows facing south to make the most of natural light in letting in maximum light, the kitchen is on the eastside, and finally the bathroom and toilets that are directed to the north (figure 3).

Figure 2. Rural housing prototype Figure 3. Layout plan of the house prototype

2. The initial conditions 2.1. Solar radiation

The direct and diffuse solar radiation on the ground is presented for each month of the year in figure 4. 2.2. Global radiation captured by an inclined captor

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The figure 5 shows the annual mean values of daily and monthly irradiation of the Algiers region on a plane facing south and inclined at an angle equal to 40°, 50° and 60° to the horizontal plane and without any barrier that to could act as a mask [6]. 2.3. The maximum monthly temperatures, minimum and average

The figure 6 shows the variation of the maximum external temperature, minimum and average.

2.4. Monthly temperature of cold water

The temperature of cold water in a given location varies during the year of almost sinusoidally between a minimum in winter according to the location to a maximum in summer. Figure 7 shows the change in average monthly temperature of the cold water during the year for Algiers.

Figure 4. Direct and scattered irradiation on the horizontal plane in Algiers during a year

Figure 5. Overall irradiation sensed by the captor plane behind a standard glazing on a south plane for inclinations of 40°, 50° and 60° in Algiers

Figure 6. Average monthly maximum, minimum and average temperatures during a year in the Algiers region

Figure 7. Change in average monthly temperature of cold water during a year in the Algiers region

Numerical simulation

1. Heating Needs In determining the heating requirements, a thermal balance of the house was made with the thermal dynamic simulation software "Pleiades + COMFIE" (figure 8a and 8b). The "Pleiades + COMFIE" software incorporates several thermal data libraries on materials and constructive elements, joinery, surface states, albedo and plant screens. The software also includes hourly building management scenarii for a typical week (occupancy, internal gains, heating set-point temperatures and cooling).

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a)

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Figure 8. PLEIADES simulation software + COMFIE

2. Requirements hot water The energy requirements for hot water depending on the type of use of the area and of the reporting period is calculated analytically with Equation 1 and the following assumptions :

- Hot water at 50 ° C,

- The cold water inlet temperature depends on the climate zone, it is presented in figure 7.

Qecs = ρw .cw .Vuw .(Tec - Tef).nsem,nor(Wh) (1)

3. Energy distribution of losses for a hot water installation

3.1. Loss of energy on hot water In calculating the energy losses associated with the distribution of hot water we use Equation 2.

The following assumptions are made for the calculation of these losses:

- distribution losses in a hot water draw-off are not considered.

- It is considered that at the end of the draw-off, the hot water contained in the network cools and its energy is lost.

- It retains, in addition, the following conventions:

o the amount of water contained in the network is equal to 1.2 l for each point of use, this corresponds to 5 m length of a 16/18 network,

o water is circulated at a Tec = 50 °C, o the number of puisages np is 4.

This leads to the following formula for lost energy:

Qpecs= cw .npp .v . (Tec – Tam).Nsubnor.np (Wh) (2)

3.2. Loss of energy on hot water heating The calculation is performed by equation 3.

Qpch= cw .v . (Tec – Tam).Nsubnor (Wh) (3)

RESULTS

1. The monthly energy needs The energy requirement for heating and hot water are shown graphically in figure 9.

We note that monthly heating energy requirement depends essentially on the period and the external climatic conditions of the month, while the energy requirements for hot water do not change significantly because it does not depend strongly on the climate (except for the cold water temperature), but depend on the daily hot water consumption for each individual.

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Figure 9. The monthly energy needs for heating, cooling and domestic hot water

2. Determination of optimal solar collector area The 50° inclination of the solar captors is selected for it gives the highest energy during the heating period among the 3 inclinations considered at the beginning (40°, 50° and 60°), namely 515 kWh from 40°, 535 kWh from 50 kWh and 530 kWh from 60° C for the months of November and December, January, February and March. The 50° inclination of the solar captors is the same as the inclination of the roof of rural prototype.

Depending on the heating demand for hot water and also depending on the monthly losses of energy for heating and hot water, we sketch the curves representing the needs and energy intake of the flat solar collector, inclined 50° to the horizontal and oriented south.

Figure 10. Difference between the period of the heating needs and hot water and the period in which solar energy is most

available for multiple sensing surfaces

The figure 10 shows the energy requirements for heating and hot water and the curves of global irradiation on a flat captor of 1 m², 2 m², 4 m², 6 m², 8 m² and 10 m².

We remark that the period with the greatest amount of solar energy does not coincide with the period when energy demand is the highest. While the energy consumption for hot water remains relatively constant throughout the year, this is when the heat requirements for heating are the highest that solar energy is the least abundant. If it is desired that the solar panels are also used for heating purposes, it is necessary to provide a sufficiently large surface area, which will eventually lead to stagnation of the solar circuit in summer.

In figure 10, we see the curves to ensure that many of the needs for heating and hot water requires that the solar collector area between 4 m² and 10 m². To choose the optimal surface, it must take into account several aspects, such as:

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- The feasibility of installing the captor on the roof, - Supply more energy for the needs, - The economic aspect.

Considering these aspects, we decide for an area of 8 m², the captor is easily installed on the roof of the house and also with 8 m² captor it is easy to meet the need for hot water.

3. Energy requirements for housing After having pre-dimensioned solar captor, were sketch in figure 11, the needs of the rural house for heating and hot water, thus we distinguish three areas:

- Area 1: This is the additional energy to solar energy necessary to ensure the energy needs of the house, this energy or part of it, must been provided either by a storage tank or another system.

- Area 2: the usable solar energy, this energy is not fully provided by the solar captor,

- Area 3: solar energy in excess, this energy is not directly used.

Figure 11. Representation of hot water and heating needs and solar gain of a house

DISCUSSION

This article is aimed at a solar installation for heating and producing hot water for a rural house designed for low energy consumption whose annual energy needs for heating are 4709 kWh and 3485 kWh for hot water.

The results show that 74 % of the energy consumption or 3720 kWh, are consumed during the coldest months (December, January and February) of the year. The remaining of the energy is consumed during the months of October, March and April. The domestic hot water consumption stays almost constant for each month and it is between 275 kWh in July and 345 kWh in January.

The results show also that an 8 m2 solar panels are required to provide 74% of annual energy for heating and producing hot water (by solar energy). However, only 54 % of energy needs can be provided during the coldest months (December, January and February).

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REFERENCES 1. D.T.R. C 3-2, Réglementation thermique des bâtiments d'habitation, « Règles de calcul des déperditions calorifiques des bâtiments », fascicule 1, 2ème Edition (1998). 2. GIOL (L.). - Une expérimentation de douze maisons solaires passives et actives à Bassens en Gironde. C.E.T.E. Bordeaux, 1983. 3. MANDINEAU (D.). - Maison individuelle à chauffage par plancher solaire direct (villa Morant). Colloque A.F.M.E. Performances expérimentales des installations solaires à capteurs plans, Marseille, 1985. 4. GIOL (L.). - La performance énergétique grâce au soleil dans l’habitat social. C.E.T.E. Bordeaux, 1988. 5. A.J.E.N.A. - Suivi d’une installation de chauffage par Plancher Solaire Direct. A.J.E.N.A., Lons-le- Saulnier, 1991. 6. M. CAPDEROU, ATLAS SOLAIRE de l’ALGÉRIE, Tome 3 « Aspect géométrique, Synthèse géographique », Volume 1, Alger 1986.

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[Abstract:0079][Comfort Cooling] EXPERIMENTAL INVESTIGATION ON PROTOTYPE CELLS USING WALLS

INCORPORATING PCM

Lotfi DERRADJI1,2, Farid BOUDALI ERREBAI1, and Mohamed AMARA1 1 National Center of Studies and Integrated Research on Building Engineering (CNERIB), Cité Nouvelle El Mokrani,

Souidania, Algiers, Algeria. 2 Department of Mechanical Engineering, University of Blida, BP 270 route de Soumma, Blida, Algeria.

(Corresponding Email: lotfi.derradji @ yahoo.fr)

SUMMARY This paper presents an experimental investigation of a new use of phase change materials as coating material on concrete and hollow bricks walls. Its aim is to study the influence of the integration of PCM on the thermal behaviour of cells and on the improvement of thermal comfort in the algerian climate. In situ measurements of air and walls temperatures are made in three cells located in the Algiers region. The measurement results show that the use of a gypsum plaster incorporating 30 % PCM contributes to a reduction of the amplitude of indoor temperature of the cell by 4 °C in the summer period. The use of PCM has also improved thermal comfort and increased the cell maximum temperature of 2 °C during winter. INTRODUCTION The use of phase change materials (PCM) is a solution to improve the energy efficiency of buildings. The phase change materials (PCM) can store heat to compensate for the mismatch between supply and hourly heat demand in a building, such as the storage of solar thermal energy for heating during the evening [1−6]. Several investigations have been conducted to study the thermal behaviour of cells with walls incorporating phase change materials. Among these studies, Ahmad et al. [7,8] studied the thermal performances of a test-cell with a new structure of light wallboards containing PCMs submitted to the outside climate in France and a comparison was made with a test-cell without PCMs. The evolution of the indoor temperature of cells, the flux densities and temperatures at the surface of the walls was measured. It was found that the walls containing the PCM properly play their role of "thermal dampers" and that the fluctuations of the internal temperatures were considerably reduced. The concept of coupling PCM with a super insulation material proves to be a promising solution for light envelopes of low thickness having a good insulation and a significant inertia. Kuznik et al. [9] conducted an experimental study of the thermal performance of a PCM copolymer composite wallboard in a full scale test room. The effect of the PCM is studied by comparing the results obtained with and without composite wallboards for three cases: a summer day, a winter day and a mid-season day. The results show that the presence of PCM reduces the air temperature of the room up to 4.2 °C. The decrement factor observed for the cases with PCM wallboard and with regular wall is about 0.7 for all the seasons tested. Fluctuations in the surface temperature of the walls are also reduced. The walls incorporating PCM improve thermal comfort by returning the heat when the room temperature drops. Cabeza et al. [6] studied the thermal aspect of a new innovative concrete incorporating phase change materials (PCM) to develop a product that provides a significant energy savings on heating and cooling buildings. Two real size concrete cubicles have been made to study the effect of PCM with a melting point of 26 °C. The two prototypes were built in the location of Puigverd of Lleida (Spain). The results of this study show that the energy storage in the walls by encapsulating PCMs and the comparison with conventional concrete without PCMs, leads to an improved thermal inertia and internal temperature with lower amplitudes. These results demonstrate that the incorporation of PCM with concrete provides a real opportunity to achieve energy savings for buildings. The main objective of this paper is to demonstrate the possibility of using microencapsulated PCM with plaster to improve the thermal comfort in the Algerian climate and to achieve energy savings for buildings. Three cells were performed with the walls of the conventional type in Algeria using an interior coating plaster/PCM. The cells are located in the Algiers region, more precisely in the village of Souidania. Measurement instruments have been installed to determine the influence of PCM on the thermal behaviour of cells in Algeria during summer without cooling and during winter with heating. Thermocouples were installed within the cells for measuring the temperature of the inner face of the walls and the air temperature. The outdoor temperature and the direct solar radiation were also measured. CELLS DESCRIPTION In order to study the influence of the Phase Change Materials on the energy efficiency of buildings in Algeria, three cells of 1 cubic meter were constructed in the same manner except that the inside mortar is different (Figure 1). These cells are located in the Algiers region, more precisely in the village of Souidania. This region belongs to the climate zone (A) (Latitude 36.70 N, Longitude 03.20 E), which is characterized by a cool winter and hot and humid summer (DTRC 3.2) [10].

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Figure 1. The three cells of 1 cubic meter

The first cell is considered as a reference element. It is made with conventional type walls in Algeria. The floors are made with reinforced concrete and the walls are made with a single hollow brick wall. The top floor and the walls are coated on the inside with pure plaster. A PVC double glazed window (60x60 cm²) is installed on the south side of the cell. In order to increase the thermal inertia of the walls, 30% of a Phase Change Material (Micronal DS 5001 provided by BASF), in powder form, was introduced with coating plaster in the second cell C2. The phase change temperature of the PCM is 26 °C. According to the product PCM description, the Micronal DS 5001 has a capacity of latent heat of 110 kJ/kg. In cell C3, the roof and the walls are plastered on the inside with a coating composed of 70% plaster and 30% phase change material (DS 5008). The PCM, DS 5008, used in this case (cell C3), has a phase change temperature of 23 °C and a latent heat capacity of 100 kJ/kg .

IN SITU MEASUREMENTS

In order to study the thermal behaviour of cells with Phase Change Material walls, two T-type thermocouples are installed within each cell to measure the temperature of the ceiling and of the south wall. A thermocouple is installed in the outside in order to determine the outdoor temperature. The different thermocouples are connected to the KEITHLEY 2700 device, which is a data acquisition system of 40 measuring channels. Measurements are recorded every half-hour. The thermocouples were calibrated using a calibration furnace (CONTROLAB, CALIBRATED), to check the response of each thermocouple and detect anomalies in the measurements. The measurements of the air temperature in the cells are made using the thermo-hygrometers TESTO-H1 175 installed in the cells. During the summer period, the thermal measurements were performed in the cells which were not conditioned. In the winter period, the cells were heated by lamps, of a power equal to 200 W during the day (9 am to 3 pm), the windows were covered by a film of aluminium in order to make sure that the luminous flux will heat the cell and to prevent it from spreading outside through the glazing. Figure 2 illustrates a sketch of the acquisition system used to study the thermal behaviour of cells.

Figure 2. Sketch of the acquisition system

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RESULTS OF THERMAL MEASUREMENTS IN SITU

THERMAL MEASUREMENTS IN SUMMER Figure 3 illustrates the evolution over the time of the outside air temperature and that of the inside air of the three cells (C1 cell without PCM, C2 cell with PCM (26 °C) and C3 cell with PCM (23 °C)) for the period from 05 to 09 September 2012. The results show that during the day, the maximum temperature of the interior environment of the C2 and C3 reaches 32.5 °C whereas that of the C1 cell exceeds 35 °C. The heat absorbed by the PCM during its phase change resulted in a slow increase in the air temperature of the C2 and C3 cells compared to that to the cell C1. During night period, the temperature of the C2 and C3 cells drops to 22 °C, whereas that of the C1 cell decreases up to 21 °C. It is noted that the cooling of the C2 and C3 cells with heat stored in the PCM is slower compared to that of the C1 cell. The amplitude of the variation of the air temperature in the cell C2 of PCM (26 °C) and the cell C3 with the PCM (23 °C) is almost the same, the two cells temperature varies between 22 and 32 °C. The results also show that the presence of 30 % PCM material with plaster can reduce the amplitude (daytime/night) of the temperature of the air by 4 °C.

Figure 3. Evolution over the time of ambient temperature (7−9 September 2012)

Figure 4 shows the evolution over time of the wall temperature of the C1, C2 and C3 cells for the period from 05 to 09 September 2012. The analysis of experimental results in the first three days, shows that the temperature of the wall of the cell C1 varies from 20 to 34 °C with an amplitude of 14 °C. While the temperature of the wall of the cells C2 and C3 varies between 22 and 30 °C with an amplitude of 8 °C. In the last two days, the temperature of the wall of the cell C1 varies between 21 and 38 °C with an amplitude of 17 °C, while that of the cells C2 and C3 is between 22 and 34 °C with an amplitude of 12 °C. The lowest temperature of the wall of the cell C2 with PCM 26 °C is higher than that of the cell C3 with PCM 23 °C. The C2 cell starts to release its stored heat before the C3 cell since its melting temperature is greater than that of the C3 cell. The results revealed that the coating of a hollow brick wall with a PCM has increased its thermal storage capacity, resulting in a reduction of the amplitude of the surface temperature of 6 °C. The incorporation of a PCM material brings an increase in the thermal mass of the bricks that already has a high level of thermal insulation compared to other materials.

Figure 4. Evolution over the time of wall temperature (7−9 September 2012)

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THERMAL MEASUREMENTS IN WINTER Figure 5 illustrates the evolution of the outdoor temperature and the indoor temperature of the cells for the period from 24 to 28 February 2013. The cells were heated by lamps of 200 W power, during the day 9 am to 3 pm. The results show that with the same heating power, the maximum air temperature of the C2 and C3 cells are higher than that of the C1 cell. It is found that in the first three days, the PCM improves thermal comfort and increases the maximum temperature by 2 °C. In the last day, the temperatures in all the cells increase and become close to 26 °C. When the heating of cells is shut down, the cooling of the C2 and C3 cells is slower compared with that of the cell C1, this is due to the release of the heat stored in the PCM. When the cells are left overnight without heating, their temperatures are identical, due to the negligible effect of the PCM.

Figure 5. Evolution over the time of ambient temperature (24−28 February 2013)

Figure 6 shows the evolution over time of the temperature of the inner face of the wall of the cell C1 without PCM and that of the C2 cell with PCM (26°C) and C3 cell with PCM (23°C) for the period from 24 to 28 February 2013. The results show that for the cells heated with a lamp of 200 W power, the temperature of the PCM incorporating walls rises above 20 °C, during the day. It is noticed that in the first four days, the maximum wall temperature of the C2 and C3 cells vary between 20 and 24 °C, and the maximum temperature of the wall without PCM (cell C1) varies between 17 and 21 °C. The wall temperature of the C3 cell is slightly higher than of the C2 cell. In the last day, the maximum temperature of the wall without PCM is 29 °C , while the temperature of the C2 wall reaches 31 °C and that of the C3 wall reached 33 °C. The minimum wall temperatures without heating vary in the same way between 6 °C and 3 °C during the first four days. For the last day, the minimum temperatures rise to 10 °C. The results revealed that the coating of a hollow brick wall with a PCM material improves thermal comfort and increase the maximum wall temperature by 4 °C.

Figure 6. Evolution over the time of wall temperature (24−28 February 2013)

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CONCLUSION

An experimental investigation was conducted to study the influence of the integration of PCM on the thermal behaviour of three cells and on improving thermal comfort in the algerian climate. Cell C1 is composed of construction materials conventionally used in Algeria, cell C2 is made with walls and a roof incorporating a PCM (26 °C) and C3 cell is made with walls and a roof incorporating a PCM (23 °C). The results of in situ measurements in summer have shown that during the sunny period, the incorporation of PCM in the walls and ceiling of the cell decreases by 2.5 °C the maximum indoor temperature and reduces the day-night amplitude of the cell's indoor temperature by 4 °C. The winter measurements results show that the PCM improve the thermal comfort and increase the maximum cell temperature by 2 °C. This demonstrates that the use of plaster/PCM composites as coatings provides a real opportunity to achieve energy savings for buildings in Algeria. REFERENCES 1. Zalba, B., Marın, J.M., Cabeza, L.F., & Mehling, H. 2003. Review on thermal energy storage with phase change:

materials, heat transfer analysis and applications. Applied Thermal Engineering, 23, 251-283. 2. Tyagi, V.V., Buddhi, D. 2007. PCM thermal storage in buildings: A state of art. Renewable and Sustainable Energy

Reviews, 11, 1146-1166. 3. Khudhair, A.M., Farid, M.M. 2004. A review on energy conservation in building applications with thermal storage by

latent heat using phase change materials. Energy Conversion and Management, 45, 263-275. 4. Hawes, D.W., Feldman, D., & Banu, D. 1993. Latent heat storage in building materials. Energy and Buildings, 20, 77-

86. 5. Feldman, D., Banu, D., Hawes, D., & Ghanbari, E. 1991. Obtaining an energy storing building material by direct

incorporation of an organic phase change material in gypsum wallboard. Solar Energy Materials, 22, 231-242. 6. Cabeza, L.F., Castello, C., Nogue´s, M., Medrano, M., Zubillaga, O. 2007. Use of microencapsulated PCM in concrete

walls for energy savings. Energy and Buildings, 39, 113–119. 7. Ahmad, M., Bontemps, A., Salle, H., & Quenard, D., 2006. Experimental investigation and computer simulation of

thermal behaviour of wallboards containing a phase change material. Energy and Buildings, 38, 357-366. 8. Ahmad, M., Bontemps, A., Salle, H., & Quenard, D., 2006. Thermal testing and numerical simulation of a prototype cell

using light wallboards coupling vacuumisolation panels and phase change material. Energy and Buildings, 38, 673-681. 9. Kuznik, F., & Virgone, J. (2009). Experimental investigation of wallboard containing phase change material: Data for

validation of numerical modeling. Energy and Buildings, 41, 561–570. 10. Document Technique Réglementaire (DTR C3-2). (1997). Réglementation thermique des bâtiments d’habitation. Règle

de calcul des déperditions calorifiques. National Center of Studies and Integrated Research on Building Engineering (CNERIB). Algeria.

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[Abstract:0080][Energy Efficient Buildings]

ENERGY EFFICIENT ROOM AUTOMATION PRACTICES BASED ON EN15232 AND VDI3813

Cristóbal Fernández 1, Mustafa Değirmenci 2 1 SAUTER Building Control International Ltd., Freiburg, Germany

2 SBC Otomasyon Sistemleri San.ve Tic. A.Ş., İstanbul, Türkiye CRISTOBAL.FERNANDEZ@DE.SAUTER-BC.COM

SUMMARY

The role of the room automation is providing a pleasant environment with the right amount of light, good air quality and a comfortable temperature. Integrated room automation is defined as the optimized interaction of the controlling of the room air conditioning, the lighting and the solar protection. The condition for an optimal operation of the building according to the BA efficiency classes is a good choice of the room automation functions. The Energy Performance of Buildings Directive (EPBD), which was adopted by the EU in 2002 and revised in 2010, is an important step towards the improvement of the energy efficiency of the large stock of existing buildings in Europe. As a result, more than forty EN standards have been developed with the aim of harmonizing the methods of calculating the energy consumption of buildings in Europe. EN 15232 is a summary of 20 years of experience in field of energy efficiency in buildings, and it shows how building automation can be used to minimize energy consumption, and it helps to quantify the measures. One of the central requirements of EN15232 is the demand led operation of the heating, cooling ventilation and lighting. A smart balance of the energy sources based on the current demand will improve the energy efficiency significantly without compromising the comfort. A key to the implementation is the communication between the energy consumers, the energy distribution and the energy sources. INTRODUCTION

The primary purpose of a building is providing a space for working, living, presentations, pleasure, etc. Each building must satisfy the user requirements. The most favorable conditions possible for humans and furnishings must be created in the building. The building automation ensures that, depending on the use requirements of the building, optimal climatic conditions prevail. It links all existing systems in a building and optimizes the interaction in terms of the best ambient conditions with maximum energy efficiency. The prerequisite for energy efficiency at a desired level of comfort is an intelligent building automation system that provides the required energy in the right quantity at the right time and at the right place. A large proportion of the heating, cooling, air conditioning and lighting provided in the building is not used, so we can say that it is wasted. This means they are provided even when this results in no benefit at all. A continuous process of the consumption recording - analysis - improvement can reduce the consumption of energy in the phase of utilization of a building drastically as well as lower the expenses and spare environment. A modern building automation system can support the users in this endeavor. Some practice examples show that with reasonable investments, e.g., a saving of 24% of heating energy is realistic. Effective measures for preventing waste are described in the European standard EN 15232. The principle or central theme in this set of regulations is demand-based operation of the systems. Energy in the form of heat, cold, conditioned air and lighting should only be provided when there is demand on the user side. Usage and thus, in large part, the users dictate when and how much energy must be consumed. Systems with consistent, automatic identification of needs with, for example, occupancy detectors and a room control which requests the energy requirement in a targeted manner during energy processing, achieve the highest energy efficiency class. By implementing the functions explained in the standard, the BMS industry is playing a major part in meeting the requirements for energy demand and climate protection.

Demand control according to EN15232 1

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This is achieved by a substantial, long-term reduction in energy demand and CO2 emissions for buildings equipped with appropriate regulation and control technology. Building automation measures feature shorter amortization periods than those based on building physics. The basic principle is: Use energy only when it results in a benefit. Eliminate waste! METHODS

EN15232 provides a framework for a systematic rating of the energy efficiency of a building at planning time, defining the requirements to obtain a rating from A (best) through D (worst). The building blocks that have to be implemented in the room automation are listed below.

A checklist defines the score that can be obtained depending on the implemented functionality eu.bac System key performance indicators (KPI) are key figures from the operating data in a building automation system with information on:

Energy efficiency of a building automation system / components Daily value in relation to utilization variables / standard variables Diagnosis through calibration of the plant Manual operation over long periods of time Adherence to comfort conditions (temperature, humidity, indoor air quality)

These KPIs are calculated on the automation level, summarized to daily values and shown each day using a “traffic light” system. Thanks to the key performance indicators, operators can detect and optimize settings; energy consultants will find a reliable data source and a basis of energy optimization. The basic principle here is: The greater the level of instrumentation and the finer the resolution, the more accurate the information FUNCTIONS FOR THE ROOM CONDITIONING

Functions for controlling the room conditioning (temperature, air quality, air humidity) with the room automation unit (heating, cooling, ventilation) Heating

For the heating, the controller is installed on the room level (individual room, room segment or multiple rooms as an area). For energy-optimized controlling of the room temperature, the room controllers must be in communication with the building management system. Occupancy-dependent, demand-led controlling

increases the efficiency. A controller with a time program enables intermittent operation for fixed occupancy patterns, flexible switching (optimised switching) or demand-led usage (Comfort, Precomfort, Economy, Protection). Cooling

For the cooling, the controller is installed on the room level (individual room, room segment or multiple rooms as an area). For energy-optimized controlling of the room temperature, the room controllers should be in

A High energy performance

B Advanced BACS

C Standard BACS

(reference)

D Not energy efficient

The energy efficiency is ratedfrom A to D

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communication with the building management system. Occupancy-dependent, demand-led controlling increases the efficiency. A controller with a time program enables intermittent operation for fixed occupancy patterns, flexible switching (optimized switching) or demand-led usage (Comfort, Precomfort, Economy, Protection). Heating and cooling – interlocked

Heating and cooling never occur simultaneously in the room, therefore this is automatically interlocked. Complete interlocking also guarantees the highest level of energy efficiency. Should dehumidification be required, it can technically involve “simultaneous” heating and cooling, and is therefore not to be seen as a function of

interlocking heating and cooling. Ventilation and air-conditioning

The air-conditioning and/or air quality on the room level can be controlled by controlling the ventilation or air volume respectively on the room level or by means of free, automatic cooling. Energy-optimized controlling in the room is by means of a time program for occupancy or with an occupancy detector. The most optimal

installation is demand-led controlling based on the room air quality (CO2, VOC…). Both free, automatic cooling (night cooling, free cooling) and H,x-led controlling enable energy-saving potential for refrigeration energy preparation. Free night cooling

During periods in which the room is unoccupied, the cool outside air is used for free night-time cooling, e.g. via automatically opening windows. The most cooling energy can be saved when the cool outside air is used to adjust the automatic cooling during the entire period.

Air quality control – occupancy-dependent

Occupancy-dependent air quality control enables optimised room conditioning for occupied and used rooms. Depending on whether the room is occupied, increasing the fan speed can help raise the proportion of fresh air. Occupancy switches, occupancy detectors and/or time programs for the room occupancy can define the

occupancy. The ventilation then usually reacts on different fan levels. Air quality control – air-quality-dependent

Air-quality-dependent room control optimises the room conditioning depending on the actually measured room air quality (CO2, VOC…) and creates fresh air by means of room fans, or controls air dampers for the supply air with a fresh air portion. The room fans are usually controlled with continuous control signals.

Room air humidity control

The humidity control in the room and the monitoring of the humidity in the supply air are implemented with humidifying and dehumidifying devices (or reheating of the supply air, dew point control). For optimal room conditioning, the controlling is structured within a comfort zone (temperature, humidity = enthalpy).

Energy level selection and setpoint determination

With energy level selection or a time program for occupancy, the demand- or occupancy-based controlling determines the suitable setpoints for the integrated room automation and where applicable, pre-conditions the air-conditioning suitably

FUNCTIONS FOR THE LIGHTING

Manual lighting control Manual lighting control is based on manual switching on/off. In this case, the energy optimisation fully depends on the room users present and their knowledge of energy saving. Additionally, the switching off can also occur automatically, e.g. using a timer.

Lighting control – with switch-off delay

Lighting control with switch-off delay is switched on and off manually by means of a switch. In addition, the light is switched off automatically at least once a day.

Lighting control – occupancy-dependent Occupancy-dependent lighting control can be performed in different ways and to meet different requirements.

Automatic switching on / automatic dimming Automatic switching on / automatic switching off Manual switching on / manual dimming Manual switching on / automatic switching off

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Lighting control – dependent on occupancy, daylight; switched Occupancy- and daylight-dependent lighting control switches on the lights automatically depending on the proportion of daylight. A light sensor in the room captures the daylight, and if the brightness is below a pre-defined setpoint, the light switches on automatically if the room is occupied.

Lighting control – dependent on occupancy, daylight; dimmed

Occupancy- and daylight-dependent lighting control dims the lights automatically depending on the proportion of daylight.

FUNCTIONS FOR THE SOLAR PROTECTION

Solar protection control – manual Manual controlling of the solar protection devices can prevent overheating or avoid glare. In this case, the energy optimisation fully depends on the room users present and their knowledge of the energy-saving effects of solar protection devices.

Glare protection – light-dependent

Light-dependent glare protection - that is, automatically controlled reduction of the incoming light - also reduces the cooling energy in the summer, aside from protecting from glare.

Glare protection – dependent on the sun’s position Glare protection dependent on the sun’s position ensures optimal slat adjustment depending on the date/time and the current position of the sun, and on the location and direction of the window blinds.

FUNCTIONS OF INTEGRATED ROOM AUTOMATION

Integrated room automation is defined as the optimized interaction of the controlling of the room air conditioning, the lighting and the solar protection. When the optimal room automation functions are selected, the building can be operated optimally in accordance with the BA efficiency classes. BA efficiency classes for room automation

The functions of the integrated room automation are defined by BA efficiency classes and are selected in such a way that the most energy-efficient building automation possible can be achieved.

Automatic thermal control Automatic thermal control uses the solar protection to support the heating and cooling processes in unoccupied rooms. In winter when the solar protection is open, the incoming solar radiation lessens the heating required, and in summer when it is closed it prevents overheating (reduction in the cooling energy used).

CENTRAL ROOM MANAGEMENT TASKS

Along with the local operating and display functions with the standard room operating units (outer, light- grey triangle), the central functions on the management level of the building automation are summarized for the various requirements of the room automation and for the room management (inner, dark-grey triangle).

Environment/weather Environmental factors affect the integrated room automation to the extent that weather conditions particularly affect, according to

priority, the regulation and controlling of the solar protection device. A central weather station on the building performs this task. For slow heating/cooling reservoirs (TABS: thermoactive building elements), weather forecast data can also be used to predictively influence the room automation. Optimisation

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Technical house and building management and communicative control facilities enable continuous, central monitoring and optimization of the setpoints and control parameters (coefficients for PI-controllers). Central, automatic correction and optimization of setpoints, as well as the set control parameters, affect the energy-efficient and therefore the cost-optimised operation of the building. This central optimization fulfils the specified

BA efficiency class. Energy data

The central acquisition of energy consumption data, separately for every building section or even every room, and the displaying and logging of this energy data, can contribute to better, more energy-efficient operation of the building.

An energy management system with energy monitoring (energy data acquisition and monitoring), energy data logging, energy billing or benchmark, can be used by the operator to minimise the energy costs. With virtual acquisitions of consumption data in each room (virtual counters), the costs for metering devices is minimised while the room users are still informed of their energy consumption Time programs

The building management system also centrally manages the time programs and (operating) calendars for the overall operation of the room and building automation. As the time profiles and calendars are stored locally in the respective automation stations, the building management system only has to manage and synchronize the time

programs and calendars. Alarm

Notification is important for the safe operation of the room and building automation. Alarm monitoring and forwarding, but also alarm confirmation by the user, with or without an audit trail, and alarm logs, are the tasks of a notification system integrated into the building management system. The various prioritized events and alarms are displayed clearly in alarm and event lists. Important alarms can also be forwarded to defined persons (alarm

dispatching). Trend

To ensure the continuous, high-quality operation of the building and rooms, the states, events and measurement and positioning values of the MCR devices are logged. This trend and event data logging is performed using current and historical databases and, optimised visually, supports the data monitoring, both for current values (live

data) and for long-term historical values (data archiving). RESULTS

Many buildings have been audited and certified in Germany and Switzerland. The lowest results come mainly from existing buildings. The methodology allows the potential for subsequent installations and optimizations to be demonstrated using concrete examples. The high ratings are found in buildings where special attention was paid to sustainability during planning and construction. Most of these buildings are also certified according to LEED or DGNB and achieve the best results under those systems.

Thermal energy factor Electrical energy factor

Class D C B A D C B A

Offices +51% 0% -20% -30% +10% 0% -7% -13%

Lecture hall +24% 0% -25% -50% +6% 0% -6% -11% Potential energy savings depending on room type and rating DISCUSSION

The Internet of Things embeds the devices and systems inside and outside the building, regardless of the vendor. The room automation stations read out the ambient conditions of smart sensors, command the actors and ask the distribution network for the right amount of heat, chill or air. The network balances the loads of the consumers and energy sources and embedded in the smart grid. Weather forecasts are used to activate thermal elements and load energy storage only as much as necessary. The management and analysis functions are moving more and more out of the building into the cloud. Open protocols grant the interoperability of the different systems. Internet Protocol IP is the undisputed standard in the internet. In the domain of the building automation BACnet has become the most common protocol.

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BACnet has its strength in the large set of objects and services specific to the building automation that make it not only suitable for the communication between controllers and field devices, but also with management services. The way of handling alarms, historical data, schedules, I/O and lighting objects is standardized, making it easy to embed the systems of the different vendors. The adherence of the devices to the standard may be certified by the BACnet Laboratory. A programmable room automation station, powerful engineering tools and well proved libraries are the ideal combination to implement the optimal control strategy, no matter how complex the task is. REFERENCES

EN15232:2012: Energy performance of buildings — Impact of Building Automation, Controls and Building Management.

Directive 2002/91/EC of the European Parliament and of the Council of 16 December 2002 on the energy performance of buildings

eu.bac: www.eubac.org: European Building Automation Controls Association

BACnet57%

LonWorks14%

Modbus5%

KNX5%

Other1%

Proprietary18%

Market share of communication protocols in thebuilding automation (source: BSRIA)

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[Abstract:0081][Energy Efficient Buildings]

FLEXIBILITY OF THE CEN AND ISO STANDARDS ON ENERGY PERFORMANCE OF BUILDINGS ASSESMENT PROCEDURES SUPPORTING THE SUSTAINABLE AND

SMART COMMUNITIES

Jaap Hogeling ISSO, Manager international projects and standards at the Netherlands Building Services Knowledge Centre

Chair of CEN TC 371 Program Committee on EPBD; Chair of JWG of ISO TC 163 & 205 advisory group on coordination of ISO and CEN Work on EPB

Fellow of ASHRAE, Fellow of REHVA j.hogeling@isso.nl

Abstract The Recast-EPBD1 requires an update of the current (2007/2008) set of CEN-EPB standards. This update work started in 2012 and will result in a new set of CEN-EPB standards.. Where possible this work will be done parallel with ISO. This project is based on EU-Mandate 480. This mandate accepted by CEN, requires a really out of the box thinking approach of the standard developers. This project is coordinated by CENTC371 the “Program Committee on EPBD” and is considered to be a step forward in progressing towards European Energy Codes for Buildings. This second generation of EPB standards aims on more comprehensive standards, a clear split between informative text in Technical Reports and normative text in Standards, attached excel files to illustrate the calculation procedures etc.. The EPB2 set of standards and technical reports will support the holistic approach needed for the Nearly Zero Energy Buildings (nZEB) and high performance energy renovation of the existing building stock. The modular structure of EPB standards is flexible in order to take into account national, regional and regional choices. An approach has been introduced, via the so-called Annex A and B in all EPB standards. Annex B is an informative Annex and includes all default values, choices and options needed to use the standard. Normative Annex A includes empty tables for these needed values, choices and options, this empty template shall be used by National Standard Bodies (NSB) (or recognized local, regional or national authorities) to declare these values, choices and options to be followed under their jurisdiction. This approach allows maximal flexibility and transparency in applying the EPB standards. If published by the NSB’s These filled in Annexes conform Annex A are indicated as National Annexes. The flexible approach included in these EPB standards, sometimes criticized, but allowing maximal freedom in innovative design approaches , able to demonstrate the impact of smart energy infrastructures as expected in future smart communities. Formal Voting drafts of all EPB standards are expected to be ready by April 2016. After the EPB standards are accepted the publication by the end of 2016 seems possible.

Keywords: EU Energy Performance Buildings Directive; CEN ISO Standards; EPBD; smart energy infrastructures

1. Introduction Analyses regarding the use of the in 2007/2008 published set of CEN-EPB3 standards and the requirements set out in the recast-EPBD showed the clear need for a second EU mandate to CEN in order to improve these standards. The revision will improve the accessibility, transparency, comparability and objectivity of the energy performance assessment in the Member States, as mentioned in the EPBD. The "first generation" CEN-EPB standards were implemented in many EU Member States "in a practical way". Typically: partly copied in "all in one" national standards or national legal documents, mixed with national procedures, boundary conditions and input data. For a more direct implementation of the EPB standards in the national and regional building regulations, it is necessary to reformulate the content of these standards so that they become unambiguous (the actual harmonized procedures), with a clear and explicit overview of the choices, boundary conditions and input data that can or needs to be defined at national or regional level. This implies that the current set of CEN-EPB standards is improved and expanded on the basis of the recast-EPBD.

1 EPBD: DIRECTIVE 2002/91/EC OF THE EUROPEAN PARLIAMENT AND OF THE COUNCIL of 16 December 2002 on the energy performance of buildings. Recast-EPBD: DIRECTIVE 2010/31/EU OF THE EUROPEAN PARLIAMENT AND OF THE COUNCIL of 19 May 2010 on the energy performance of buildings; (recast). 2 In this paper EPB stands for “Energy Performance of Buildings” the D for the EU-Directive is intentional deleted in relation to the standards. The EU-directive is of great importance for the EU-member states however these CEN standards could become ISO standards as well and it is more appropriate to use just EPB.

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The standards shall be flexible enough to allow for necessary national and regional differentiation to facilitate Member States implementation. Such national or regional choices remain necessary, due to differences in climate, culture & building tradition and building typologies, policy and/or legal frameworks.

2. Work in progress, the last phase of the on-going work on the EPB standards The EPB standards have been developed by the following CEN/TC’s:

- TC 089 Thermal performance of buildings and building components; - TC 156 Ventilation for buildings; - TC 169 Light and lighting systems; - TC 228 Heating systems for buildings; - TC 247 Building automation, control and building management; - TC 371 Project Committee on Energy Performance of Buildings.

These TC’s are responsible for the technical content of EPB standards to be revised. CEN/TC 371, the overall responsible coordinating committee, also ensuring that the timetable will be met and that the basic principles and rules, the modular approach and the foreseen improvements of the current set of EPB standards, are in line with the targets indicated and meeting the expectations of the end users. CEN/TC 371 formulated common Basic Principles (CEN/TS 16628:2014) on the required quality, accuracy, usability and consistency and a common format for EPB standards, including a systematic, hierarchic and procedural description of options, input/output variables and relations with other standards and elaborated a unique hierarchic system for the EPB standards. CEN/TC 3714 prepared the Basic Principles (BP) and the supporting Detailed Technical Rules (DTR) (CEN/TS 16629:2014), as basis and guidance for the total set.

Figure 1 Current status

3. The Principles The mandate M/480 explicitly requests for identification and prioritisation of items for revision and gaps in the current set of standards in consultation with the EU member states (MS). The following, general principles are valid for the set of EPB standards : 1. The complexity of the building energy performance calculation requires a good documentation and justification of the

procedures. Informative text is required but it will be separated from actual normative procedures to avoid confusion and unpractical heavy documents. Therefore, each EPB standard ( or sometimes a close connected set of) shall be accompanied by a Technical Report where all related informative material will be concentrated.5

2. The complexity of the building energy performance calculation requires also a very good coordination and testing of each calculation module. Therefore, each EPB standard shall be accompanied by a spread-sheet where the proposed calculation algorithms and data input/output are tested and proved to be consistent. A Software Tool team checks the calculation modules of the total set of EPB standards. With this excel based software it will be possible to assure that the in/output files of the various connected EPB standards are valid.

4CEN/TS 16628:2014 Energy Performance of Buildings - Basic Principles for the set of EPBD standards CEN/TS 16629:2014 Energy Performance of Buildings - Detailed Technical Rules for the set of EPB-standards 5 Either as a separate TR or if very limited as an informative annex to the standard. It is also possible that a TR will cover more standards.

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4. The deliverables of CENTC371

4.1 CEN/TS Basic Principles CEN/TS 16628:2014 Energy Performance of Buildings - Basic Principles for the set of EPBD standards. This TS provides a record of the rationale, background information and all choices made in designing the EPB package. These basic principles are based on the analysis of the weak points within the first generation EPB package and on an evaluation of requirements by the Regulating Authorities and the outcome of the IEE-project CENSE (see http://www.buildup.eu and http://www.iee-cense.eu/ ). The TS Basic Principles provides guidance on the required quality, accuracy, usability and consistency of each standard and the rationalisation of different options given in the standards, providing a balance between the accuracy and level of detail, on one hand, and the simplicity and availability of input data, on the other. 4.2 CEN/TS Detailed Technical Rules CEN/TS 16629:2014 Energy Performance of Buildings - Detailed Technical Rules for the set of EPB-standards. This TS is based on the CEN/TS BP and provides mandatory detailed technical rules to be followed in the preparation of each individual EPB standard. This is in addition to the CEN drafting rules and complementary to the Overarching Standard (former prEN15603 and current draft-ISO 52000-1) in this article indicated as OAS. The OAS, containing the common terms, definitions and symbols and the overall modular structure for the set of EPB standards. The DTR gives a common format for each standard, including a systematic and hierarchic structure to pinpoint the position of the standard within the framework of EPB standards and procedural description of options, input/output variables. The CEN/TS DTR includes guidance for:

-a clear separation of the procedures, options and default data to be provided as default CEN option in an annex B but also allowing for national or regional choices conform the normative annex A of each of the EPB standard (where appropriate); -a specification of the input data, also indicating the source of the data if this is the output calculated according to another EPB standard or related product standard; -a specification of the intended output that is intended to provide the energy performance assessment results, the related data necessary for their proper interpretation and use, and all relevant information documenting the relevant boundary conditions and calculation or measurement steps. -an informative CEN Technical Report, accompanying each standard6, according to a common structure, comprising at least the results of internal validation tests (such as spread sheet calculations for testing and demonstrating the procedures), examples and background information. Almost all informative parts of EPB standards will be in these technical reports.

4.3 Energy performance of buildings-Overarching standard EPB; the former FprEN 15603: 2014 and current prEN-ISO 52000-1 (expected out for Formal Vote at ISO and CEN level around October 2016) This standard (OAS) specifies a general framework for the assessment of the overall energy use of a building, and the calculation of energy ratings in terms of primary energy, using data from other EPB standards, providing methods for calculating the energy use of services within a building (heating, cooling, humidification, dehumidification, domestic hot water, ventilation, and lighting). This assessment is not limited to the building alone, but takes into account the wider environmental impact of the energy supply chain. The OAS handles the framework of the overall energy performance of a building, covering inter alia:

1. common terms, definitions and symbols; 2. building and system boundaries; 3. building partitioning; 4. unambiguous set of overall equations on energy used, delivered, produced and/or exported at the building site, near-

by and distant; 5. unambiguous set of overall equations and input-output relations, linking the various elements relevant for the

assessment of the overall energy performance of buildings which are treated in separate standards; 6. general requirements to standards dealing with partial calculation periods; 7. general rules in setting out alternative calculation routes according to the calculation scope and requirements; 8. rules for the combination of different partitioning.

The OAS provides a systematic, clear and comprehensive, continuous and modular overall structure on the integrated energy performance of buildings, unlocking all standards related to the energy performance of buildings. The overall framework provided by the OAS will work as the “Backbone” (see figure 2) of the set of EPB standards, it facilitates a step-by-step implementation by the user, taking also into account the nature of each procedure identifying the

6 This to significantly reduce the length of the standards and strengthen their focus, thus facilitating the adoption (including translation) in national/regional regulations.

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typical type of user. More information is given in a Technical Report accompanying the OAS. The justification for the CEN defaults and options are provided in this TR (draft ISO TR 52000-2). Current (February 2016) status: this prEN ISO 52000-1 is passed the enquiry. The enquiry results are currently be processed. The Formal Voting is expected at the latest by October 2016. After the standard is accepted publication by the end of 2016/ beginning 2017 seems possible .

Figure 2 The OAS as backbone for the set of EPB standards This flexible approach is supporting the sustainable and smart communities. The way modules can be handled and the unambiguous set of overall equations on energy used, delivered, produced and/or exported at the building site, near-by and distant allow the flexibility needed to support the development of smart grids and local energy communities. 4.4 Draft ISO TR 52000-2 (former prCEN/TR 15615:2014) Energy Performance of buildings - Accompanying Technical Report. This draft-TR contains information to support the correct understanding, use and national implementation of this standard. This draft is expected to be published at the same time as the OAS.

5. Hierarchic numbering system - Modular structure The setup of a coherent and hierarchically numbered system of EPB standards is a requirement. Given the fact that not all standards will be ready for parallel ISO enquiry or publication and that standard numbering system in CEN doesn’t allow this, a modular structure has been developed, allowing for addressing documents given hierarchic positioning in that structure. By adding the identification code of a specific cell of the modular structure (see Figure 3 & 4) the purpose of a standard (and/or specific clauses of the standard) can be identified easily.

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Figure 3 Overarching Modular structure

Overarching Building (as such) Technical Building Systems

Des

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Des

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Des

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Coo

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Ven

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sub M1

sub M2

sub M

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M5

M6

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M8

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M 11

1 General 1 General 1 General

2

Common terms and definitions; symbols, units and subscripts

2 Building Energy Needs 2 Needs

3 Applications 3

(Free) Indoor Conditions without Systems

3

Maximum Load and Power

4

Ways to Express Energy Performance

4

Ways to Express Energy Performance

4

Ways to Express Energy Performance

5

Building Functions and Building

5

Heat Transfer by Transmiss

5 Emission & control

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Boundaries

ion

6

Building Occupancy and Operating Conditions

6

Heat Transfer by Infiltration and Ventilation

6

Distribution & control

7

Aggregation of Energy Services and Energy Carriers

7 Internal Heat Gains 7

Storage & control

8 Building Partitioning 8

Solar Heat Gains 8

Generation & control

9

Calculated Energy Performance

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Building Dynamics (thermal mass)

9

Load dispatching and operating conditions

10

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10

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10

Measured Energy Performance

11 Inspection 11 Inspection 11 Inspection

12

Ways to Express Indoor Comfort

12 BMS

13

External Environment Conditions

14 Economic Calculation

Fig. 4 – The overarching modular structure of EPB standards

6. Calculation tool and Module description The complexity of the building energy performance calculation requires also a very good coordination and testing of each calculation module to ensure coherence and the software-proof of the set of EPB standards. Therefore, each EN EPB standard shall be accompanied by a spread sheet in which the proposed calculation algorithms and data input/output are tested and proved coherent.

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Figure 5 Software check of the excel sheets of the EPB standards

7. How the EPB standards interacts with the relevant product standards Saving energy in the build environment requires not only that products consuming electricity and fuels are designed to be intrinsically more energy efficient. The interaction of a product with the rest of the system or installation in a building into which it is fitted plays an important role. This appears obvious for a number of product categories such as building equipment for ventilation, heating, cooling, lighting and control and automation. With the increasing application of electronic and communication technologies, this is also increasingly true for many other products, used in buildings but not considered as EPB related, that become ‘smart’ and ‘networked’, and can be controlled through wider systems. When EU-policies such as the Ecodesign Directive use a too narrow product-based view, products are considered irrespective if their surroundings and tested in standard conditions. If only their technical efficiency is considered, this approach may look straightforward but misses the savings that can be expected from ensuring that the product is also correctly sized, fitted and controlled to render its service optimally in a well-designed building installation. While it may be difficult to reach an EU regulation of systems under product policies, it may be possible to find creative ways for tackling at least a part of the energy savings. On one hand we have the Ecodesign Directive requiring through EU regulation minimal energy performances of energy using products. On the other side we have the EPBD where the EU Member States are obliged to require minimal target values for the energy performance of buildings, also having specific requirements for the overall thermal performance and the energy performances of the heating, ventilation lighting and cooling systems The CEN expert teams working on the different EPB-system standards have to check if the product data available on basis of product standards and/or related EU regulations are sufficient as input for their system standards. At the same moment the CEN and ISO product Technical Committees and/or experts have to be convinced that using the EPB system approach, to describe and test the products, is the most efficient way to ensure effective energy performance targets for products, systems and finally the buildings (figure 6).

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Figure 6 Products not longer evaluated as products but as part of the system.

8. Co-operation with ISO An active process of interaction for the overarching type of standards through the JWG of ISO TC 163 & 205, for the other CEN-EPB standards via the different WG’s of ISO TC 163 and ISO TC 205. Experts in the ISO and CEN teams are working on these standards, with the ultimate goal to agree on EN-ISO standards. A challenge given the geographic and other differences in the building sector. For several EPB standards under some of the CEN TC’s the cooperation with ISO is still informal. This means that for these standards no parallel voting is expected before 2017. Current parallel voting on EN-ISO EPB standards is expected for the OAS and the building thermal performance related standards as developed under ISO/TC 163. These ISO standards are indicated like EN-ISO 520xx-1 and the connected Technical Reports as EN-ISO TR 520xx-2. Several (11 of the 42) first generation EBP standards are already EN-ISO standards. They have been developed under the Vienna Agreement. Revision of these standards requires co-operation with the responsible ISO/TC. The central co-ordination of the preparation of a set of international standards on the energy performance of buildings at the ISO level is in the hands of ISO /TC 163/WG 4, Joint Working Group of ISO TC 163 and TC 205 on energy performance of buildings using a holistic approach. The main leading and active experts in CEN and ISO are among the main leading and active members of this ISO Joint Working Group. This co-operation with ISO aims to avoid serious duplication of work, to avoid incompatibilities in input) product data, procedures and (output) energy performance data.

Figure 7 Schematic operational structure.

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[Abstract:0083][Indoor Air Quality and Comfort Conditions] THE INFLUENCE OF WIND TOWER ON NATURAL COOLING AND THERMAL

COMFORT IN ARID REGIONS

Louafi Bellara Samira1 and Louafi Fadila2 1Laboratory ABE, University of Constantine 3, ALGERIA

2University Constantine 1, ALGERIA

Corresponding email: samiralouafi@yahoo.fr SUMMARY In hot and arid climates, evaporative cooling appears to have significant potential to provide thermal comfort and to reduce the energy loads of summer season. In these areas the air is hot, dry and often contaminated with dust and sand. The most used modern technique is air conditioning, which is high consumption energy, damage the environment, and create high discomfort. The proposed wind towers are installed on top of the individual houses, in the direction of the maximum wind speed in Touggourt town Algeria. This study aims at assessing the thermal performance of a bioclimatic housing using wind towers. The analysis confirms the advantage of the application of this passive cooling strategy. A simulation using TAS logical to heat and mass transfer balances to compare temperature with or without wind tower. Results shown that presence of wind towers provide thermal comfort during the warm months of the year. Wind towers maintain a constant indoor air temperature. INTRODUCTION Effective integration of passive features into the building can reduce the air-conditioning demand in buildings while maintaining thermal comfort inside the living space [1] Natural ventilation is recommended in many towns with hot and arid region, were Wind is a powerful force that should be harnessed and harvested to maximize its potential positive contribution to the built environment. Wind towers, or wind catchers, or Baud-Geers (as they are called in Persian), have been used in Iran and the neighboring countries for natural ventilation and passive cooling of buildings for centuries [2-4]. The function of this tower is to catch cooler breeze that prevail at a higher level above the ground and to direct it into the interior of the buildings.[1] Wind tower come in various designs, such as the uni-directional, bi-directional, and multi-directional (see Figure1).

Figure1: Various design of wind tower

Kalantar, 2009 investigated the performances of conventional wind tower operated in the hot and arid regions of Yazd. This study presents also a numerical technique to simulate the influence of evaporative cooling systems on the wind tower performances and they prove its potential [5] Wind towers, as a cooling system, provide ventilation and passive cooling in the hot, arid regions of Iran and neighboring countries for centuries. As wind towers work with the renewable energy of the wind, they do not use electricity and

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therefore they do not produce any greenhouse gases. Some of the defining characteristics of wind towers are related to their esthetic values while others have important roles in their thermal behavior [2-4 and 6]. The aim of the study work is to verify the hypothesis of important of role of the wind rower on thermal comfort and on evaporation cooling inside the individual houses, testing house with or without presence of wind tower and its impact on the ground and first floor. Also Comparing a the thermal performance of two different wind towers One had a wet column made of curtains suspended in the tower and the other with wet surfaces was equipped with evaporative cooling pads at the inlet under such harsh climatic conditions architectural solution. So the objective of the study is to appreciate the role of wind tower on thermal comfort. METHODOLOGY Touggourt is a town situated in the south-Est of Algeria (Figure), Built next to an oasis in the Sahara. Its geographical coordinates are: latitude 33.06°North and longitude 6.04°Est. The altitude is about 67 m over of the sea level. Its climate is dry and very hot in summer and very cold in winter. The hot period of the year is too long; it spreads from April to October. For the present study, July is chosen as a period of simulation where the demand on the air-conditioning reaches its maximum. In this work, it is presented a survey led on a new system of passive cooling which consists of modeling house with presence to a wind tower operated under the climatic conditions of Algerian desert. This survey permits us to examine, at the first time, the influence of presence of the wind tower and in the second time, the effect of the presence of water like a humidification of air for the ventilation system as the air outlet temperature, the daily cooling potential and the air velocity inside the buried pipe and inside the space. CLIMATIC AND BIOCLIMATIC ANALYSIS Climatic analysis The climatic of Touggourt city is mush contrasted; the aridity is expressed by very high temperatures in summer, very weak precipitations and by importance of evaporation due to dryness of the air. The index of aridity calculated according to the formula of Martonne is about I = 4.34, that confirms the aridity of this city.

Figure2: Variation of air temperature and air humidity at Touggourt city

The diurnal amplitude can exceed the value of 14°C; In July and august air temperature value reaches 41,8 °C (Figure2). Precipitations are very irregular. The relative humidity remains very low 29% in July and a maximum in January about 52%. In summer winds come from North-Est direction and sandy winds come from the North-East and South-West direction [7]. The climatic of Touggourt present than an extremely hot, dry and arid summer, and cold winter. Bioclimatic analysis On the psychrométrique diagram figure3, of Givoni we can deduct what follows: · The major part of April and October as well as 1/3 of May are situated in the comfort zone. · All of January, February, March, November and December are situated in the zone of the passive solar heating. The comfort can be reached (affected) only when the solar radiations for the passive heating are spread (displayed, deployed) by means of an active heating and especially in December and January · The hottest and driest months spread out 2/3 of May as well as the June to September, for it the solar control, the effect of thermal mass (thermal slowness), and especially a evaporative cooling and a night-ventilation are the strategies recommended for this period to insure indoor comfort in summer. The months of great heat require a passive Refresh: plan compact, Mass Effect and thermal inertia, night ventilation and cooling by evaporation, these passive techniques alone are not enough to maintain a temperature of acceptable comfort throughout the day, it is therefore necessary to call the air conditioning to topped up with an appropriate architecture which would take into account the recommendations previously cited. It could happen to a gain of considerable energy on the basis of proceeds purely passive, and have recourse in a reasonable manner to the man's assets.

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Figure3: Givoni bioclimatic chart applied for Touggourt

Figure4 : Szokolay bioclimatic chart applied for Touggourt

The tower to wind is a means of passive evaporative cooling very used in the warmer regions and arid areas across the world. The use of this means of refreshment liabilities in Algeria will surely have beneficial effects especially the significant gain in energy and the reduction of emissions of greenhouse gases which participates and creates the climate change [9] MODELISATION All of the components of the home, and the two specific areas (living room, kitchen and the hall) and the areas of wind towers are automatically transferred to pile Building Simulator with their volumes and their respective surfaces.

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Figure5: ground plan of the city

Figure6 : Monitored house (Ground floor and first floor) The wind tower proposed is subject to the following parameters: 1. The introduction of a system of evaporative cooling permanent of the inside walls of the wind tower and the ducts linked to a reservoir which is located in the laundry; 2. The head of the sensor to wind is multidirectional, in order to adapt to the different branches of prevailing winds; 3. The base of the sensor is composed of a basin of water and a too full with a water duct dug under ground and directly linked to a pool of water which is located in the court in order to recycle the water in case there was a surplus; 4. The base of the sensor to wind is composed of three openings in order to offer multidirectional ventilation in the home.

Figure7: zones to simulate in TAS Building simulator. (TAS 2015)

RESULTS DISCUSSION

The model is divided into 3 zones; each zone represents a space to simulate: Zone 1: Ground floor. Zone 2: First floor. Zone 3: wind tower

1 3

2

Ground floor

First floor

Wind tower

System of humidification

Water basin

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It is here to express the whole of results relating to the behavior thermo-fluid mechanics from simulations by useful criteria reflecting the conditions of interior environments. It is necessary to note that these data correspond to the situation where the house is subject to an outdoor temperature extreme. The results of temperatures and relative humidity presented are intended to analyze the thermal comfort inside of the House without and with the introduction of the wind tower t in a single day representative (21 July 2005) by an identification of areas at different temperatures and humidity. The study of the thermal behavior of the house is made through the comparison between the results of three stations: S1: The external environment, S2: Ground floor, S3: First floor. After the introduction of the information and scenarios necessary for the software "TAS" The House knows significant overheating during summer heat waves. The three stations record high temperatures. The external temperature (S1) is outside of the zone of comfort. For the (S2) and (S3) we save an overheating of the House which is spread of 10h in the morning until 22h in the evening. The house is more at least in a situation of comfort of the 23h of the evening up to 9h of the morning. An amplitude Very important is recorded: (26°C) in the (S1), (21°C) in the (S2) and (19°C) in the (S3) which implies a great sensation of discomfort of the occupants.

Figure8: Air Temperature before the introduction of the wind tower

The three stations are outside the comfort zone for the summer Figure 8. For the (S2) and (S3) There is a strong drought to the interior of the House which is spread out throughout the day. The house is in a situation of discomfort because of the humidity relatively very low, this last can cause of the sensations of choking and difficulty in breathing.

Figure 9: Air humidity before the introduction of the wind tower

The study of the effect of the sensor to wind on the refresh of the internal atmosphere of the house is made through the comparison between the results of four stations: S1: The external environment, S2: Ground floor, S3: First floor ,S4: output of the wind tower. The internal temperature of the output of wind tower at the level of the mouth of blowing (air outlet) reaches its minimum (21°C) to 5h00 of the morning and its maximum (27°C) at 16h00 in the afternoon. The temperatures of the air measured at the level of the mouth of blowing remained low throughout the day.

S1

S3

S2

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Figure 10: The air Temperature after introduction of wind tower

There is a decline very interesting temperature in the House at the level of the ground floor (S3) with a significant gap of (9.2 to 13.5 °C) in comparison with the outside (S1) whose temperature reaches its minimum (23.8°C) to 5h of the morning and its maximum (33.5) to 16h of the afternoon. The introduction of a mezzanine at the level of the first floor has allows for the exchange of air by thermal circulation between the DRC and the 1st floor, after the curves of the evolution of the temperature of the air at the top level of the house there is a slight increase in temperature with a low clearance of (0.5 to 2.3°C) between the DRC and the 1st floor, and an important deviation from (8 to 11.2 °C) between the outside and the 1st floor, including the temperature reaches its minimum (25°C) to 5h of the morning and its maximum (35.8) to 16h of the afternoon.

Figure 11: Air humidity after the introduction of the wind tower

The indoor moisture of the output of the wind tower at the level of the mouth of blowing (air outlet) reached a very high rate with a maximum of (75%) at 5h00 of the morning, and a minimum of (62%) at 17h00 in the afternoon. By comparing the relative humidity recorded at the outside in the (S1), to those of the (S3) at the level of the ground floor, we find that there is a considerable increase in relative humidity of the air with a very large gap (23 to 33%) of which the humidity reaches its maximum (61%) to 5h of the morning and its minimum (45%) to 17h of the afternoon. According to the curves of the evolution of the relative humidity at the top level of the House the (S3) saves a lowering of moisture with a gap of (7 to 10%) between the DRC and the 1erétage, and an important deviation from (13 to 26%) between the outside and the 1st floor, including the humidity reaches its maximum (51%) to 5h of the morning and its minimum (38%) to 17h of the afternoon. By comparing the measured temperatures inside before and after the integration of the sensor to wind, we note that there is a gap of (6.2 to 7°C) during the heavy heatwaves, which proves the ability of the sensor to wind of decrease indoor temperatures [8]. We note that there is a gap of (36 to 16%) during the heavy heatwaves, which proves another time the ability of the sensor to increase the relative humidity of the air for a cooling. CONCLUSION The wind tower has been able to show its effectiveness to bring closer together the internal atmosphere of the situations of comfort by reducing the temperature in the House. the temperature of the air seen that after the integration of the sensor to wind, the temperatures of the air are domestic in or very close to the comfort zone of been the city of Touggourt (28-32°C). Indeed, the effectiveness of the tower to wind is strictly linked to its role and evaporative to its power refreshing [10]. The relative humidity of the air recorded in the interior of the house (38-61%), included in or very close to the comfort zone of summer of the city. Given the conditions unfavorable to the outside and to the inside before the integration of the sensor to wind, the temperatures of the air and the relative humidity recorded during this day of simulation are considered as comfortable, there must be no more recourse to the conventional air conditioning to ensure conditions of comfort to the interior of the House.

Resim görüntülenemiyor.

S2

S2 S3

S4

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REFERENCES [1] Alemu AT, Saman W, Belusko M. A model for integrating passive and low energy airflow components into low rise buildings. Energy Build 2012;49:148 57. [2] M.N. Bahadori, Natural cooling in hot arid region, in: A.A.M. Sayigh (Ed.), Solar Energy Application in Buildings, Academic Press Inc., New York, 1977, pp.195–225. [3] M.N. Bahadori, An improved design of wind towers for natural ventilation and passive cooling, Solar Energy 35 (2) (1985) 119–129. [4] C. Karakatsanis, M.N. Bahadori, B.J. Vickery, Evaluation of pressure coefficients and estimation of airflow rates in buildings employing wind towers, Solar Energy 37 (5) (1986) 347–363. [5] Kalantar V. Numerical simulation of cooling performance of wind tower (Baud-Geer) in hot and arid region. Renew Energy 2009;34:246–54. [6] M.N. Bahadori, M. Mazidi, A.R. Dehghani, Experimental investigation of new design of wind towers, Renewable Energy 33 (2008) 2273–2281. [7]ONM meteo station of Ouargla (2014) [8]Y Bouchahm, F Bourbia, A Belhamri . Performance analysis and improvement of the use of wind tower in hot dry climate, Renewable Energy, Volume 36, Issue 3, March 2011, Pages 898-906 [9] S Outtas, S Bellara Louafi. Traditional architecture in hot dry climate of Algeria: Alesson for a sustainable architecture [10] W Cunningham, T Thompson. Passive cooling with natural draft cooling towers in combination with solar chimneys. In: Proc. of PLEA 86, Passive and low energy architecture, Pecs, Hungary; September 1e5, 1986.

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[Abstract:0087][Energy Efficient Buildings] ENERGY AND COST ANALYSIS OF THE ALTERNATIVE HVAC SYSTEM

APPLICATIONS FOR A COLLEGE BUILDING IN DOHA, QATAR

Ünal Sınar1, Hale Tuğçin Kırant2, Assoc. Prof. Dr. Ebru Mançuhan3, Assist. Prof. Dr. Barış Yılmaz1 and Dr. Mustafa Kemal Sevindir4

1Marmara University, Department of Mechanical Engineering, Goztepe Campus, Kadıkoy 34722 - Istanbul/Turkey 2Istanbul Technical University, Energy Institute, Ayazağa Campus, 34469 Maslak – Istanbul/Turkey

3Marmara University, Department of Chemical Engineering, Goztepe Campus, Kadıkoy 34722 - Istanbul/Turkey 4Yildiz Technical University, Department of Mechanical Engineering, Yildiz Campus, 34349 Beşiktaş - Istanbul/Turkey

Corresponding email: unalsinar@hotmail.com SUMMARY The objective of this study is to design an energy efficient HVAC system with long life cycle, and excellent indoor air quality while considering the cost effectiveness for a college building which was located in Doha, Qatar. The main leading consideration for whole design includes energy efficiency, health and safety, occupant comfort, functionality and etc. referring the ASHRAE standards. Before selecting a proper HVAC system, HAP (Hourly Analysis Program) was utilized to calculate heating and cooling loads and ventilation requirements. Due to the fact that the air and space design conditions are different for the classrooms, the offices, the conference hall, and the workshop of the college building, HVAC system components were selected to meet these conditions. An energy efficient has been designed for the building by comparing the various alternative HVAC systems. The first investment, operating and maintenance cost analysis were taken into account for each alternative HVAC systems as well. For final system decision, the air and other conditions like available water sources in Doha were also considered. A photovoltaic (PV) system was designed and recommended to meet some of total electricity consumption of selected HVAC system equipment.

INTRODUCTION The contents of this article summarizes the processes that were used to provide a feasible and energy-efficient HVAC system design for a new 3-story junior college classroom building, which will also house support spaces for administration, faculty offices and information technology. It is decided to build in Doha, Qatar in accordance with 2015 ASHRAE competition HVAC Design Calculations category.

The main leading consideration for whole design is Owner’s Project Requirements (OPR) which includes energy efficiency, health and safety, occupant comfort, functionality and etc. referring the ASHRAE standards such as 55, 62.1, 90.1 and so on. In addition to ASHRAE Standards Compliance requirements, ASHRAE Handbooks were also utilized for some assumptions like heat loads of electrical equipment, lighting, occupant density and etc. which are not specifically defined in OPR. Other factors in the design were location and environmental conditions, design criteria and applicability and feasibility of the system taking into, the remarks and suggestions of some consulting engineers and experts in HVAC industry, account.

The system that is compulsory to be used according to OPR is Variable Air Volume (VAV) Air Handling Unit (AHU). Due to the fact that the air and space design conditions are different for the classroom building and the workshop, system equipment selection was made to meet these conditions. While doing these, the air and other conditions like water sufficiency in Doha were also considered. The main decision that the team made for this selection was to use a VRF outdoor unit to feed the cooling and heating coil of VAV AHU instead of using a chiller or a boiler. Why we made this kind of decision will be mentioned in following sections of the article by making some comparison and demonstrating how the selection is well-advised. The conditioned air in AHU was used to satisfy heating and cooling requirements of entire the building in addition to ventilation requirement so that our system became all-air system. In addition to this decision, as specified in OPR, photovoltaic (PV) system was designed to meet the total electricity consumption of selected HVAC system equipment, in this way, green building approach was tried to be accomplished.

In order to model the building and calculate heating, cooling and ventilation loads, Hourly Analysis Program (HAP) was used. HAP is such kind of software that we were able to enter required values to create basic components of building like walls, windows, doors and etc. The most important advantage that HAP contributed us is that it uses ASHRAE Standards. However, we needed to do some correction for the values that HAP uses by updating with current ones. Load calculations of the building using HAP are resulted in 86 MBH for heating and 957 MBH for cooling. Due to the fact that selected system is all-air one, the total air volume flow rate that needs to be supplied entire the building including workshop was required to be 53390 CFM which is relatively much. Because it is not a good choice to select one AHU to meet all these requirements especially the air volume flow rate, the team decided to use four AHU units.

In order to make the final decision for system selection, several calculations and analyses such as NPV and cost analysis were done by comparing conventional chiller system and adiabatic cooling integrated one with the final system decision.

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Design Conditions In order to prepare a good project work, the team was restricted by several factors. While several design conditions that should be implemented are specified by Owner Project Requirements (OPR) for HVAC Design Calculation pillar of the competition, the judging criteria outlines how the teams’ project will be evaluated. On the other hand, the unspecified points in OPR, such as the number of the students that the classroom building houses, are determined by using the ASHRAE Standards.

The OPR summarily stipulates the following design conditions;

Temperature and Humidity

Indoor design conditions, required by the owner, are given in Table 1.

Table 1. Indoor Design Condition

Office & Administrative Support Spaces, Classroom & Study Spaces and Library

Special Instruction Spaces

Summer 73.4oF (23oC) DB, 50% RH 78.8 F (26oC) DB, 55% RH Winter 70oF (21oC) DB, 50% RH 73.4 F DB (23oC), 55% RH

The specified design requirements for the noise criteria levels are given in Table 2. This criterion leads to make an appropriate duct design, grill calculation and selection.

Table 2. Noise Criteria

Office & Administrative Support Spaces

Classroom & Study Spaces Library

Noise Levels NC 35 NC30 NC30

The system that will be implemented in the building must be variable air volume (VAV) air handling unit (AHU). Because the selection of the source of the cooling and heating for the VAV AHU is up to the team meaning that there is no restriction for this, the decision of DX AHU that must include VRF outdoor unit instead of chiller or boiler was made by the team.

To provide adequate space to allow the service of designed HVAC equipment. In order to do so, the spaces left for mechanical equipment were effectively utilized.

Selected HVAC system energy conservation requirements should be satisfied in compliance with ASHRAE 189.1 and ASHRAE 90.1.

In order to provide an excellent indoor air quality, ASHRAE 62.1 is used.

Building Properties The project involves building a new 3-story building in Doha, Qatar. It has two different structures. The main structure classroom building has 3 floors that contain break area, computer room, conference hall, classrooms, Mech./Elect room, offices and library. The other one is workshop for special instruction areas include woodworking/carpentry shop and welding shop.

Typical building operating hours are 7am-6pm Monday through Friday and 8-1pm on Saturday. There is typically full time staff, students, and instructors in the building Monday- Friday and only students and instructors (typical per default classroom counts) on Saturday. The building is closed on Sundays.

The building shall facilitate functions that accommodate the operations of the administrative staff including some individual private offices, cubicle spaces, meeting rooms, and shared spaces consisting of a reception area, storage and copier areas, small meeting and break areas, and larger meeting areas.

Envelope constructions of the building according to OPR were assumed as follows.

Window Data: Double glazed, fixed windows, ½” air space, low emissivity coating on third surface, bronze tint. Exterior Wall Data: Masonry mass wall construction, light tan color limestone. Roof Data: Insulation over concrete deck. Floor Data: Concrete poured in place mass.

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SYSTEM MODELIng AND SIZING

Modeling Software Carrier Hourly Analysis Program (HAP) Hourly Analysis Program 4.61 (HAP 4.61) was used to model HVAC system design and to obtain design calculations. It directly uses the ASHRAE standards for the amount of ventilation air, heating and cooling loads that can be standardly assumed for every space defined.

Heating and cooling loads, supply and ventilation air requirements based on spaces and zones were calculated by HAP. It could also size the heating and cooling coils used in HVAC system. After the building was modeled, very detailed reports were obtained which is related to heating, cooling and ventilation loads.

Zoning Zoning is the one of the important energy saving and comfort enhancing approaches for HVAC systems in building. Because each space in a building cannot have the same air condition depending on the variation of thermal loads, it is a necessity that air condition must be considered for the spaces in the same zone.

According to this-year ASHRAE competition, the selected building has two different structures. The main structure classroom building has 3 floors that contain break area, computer room, conference hall, classrooms, Mech./Elect room, offices and library. The other one is workshop for instruction of wood-metal works.

Modeling Method and Approach In order to make a good and proper air system design, peak zone sensible load method was used for zone airflow sizing method because each space may have its own peak load in different hours within a day.

The sources of nonresidential thermal loads or internal loads are specified by the OPR. However, what amount of load they generate and source quantities are not given. Therefore, these and other undefined loads, which are lighting and occupant density, in OPR were determined according to ASHRAE Standards 55-2013 and 62.1-2013 and 2009 ASHRAE Handbook, Fundamentals was also utilized.

Heating, Cooling Load Calculations The heating -or cooling- load is the amount of heat energy that would need to be added to a space –or removed from a space- to maintain the temperature in an acceptable range. In order to make HVAC design calculations with minimum error –even impeccable-, make a proper system selection for the building and keep the capital cost minimum while providing the maximum comfort for occupants, spaces constituting the building must be modeled well enough in used program.

Thermal loads can be considered residential and nonresidential heating and cooling loads. Residential loads are caused by the dwelling’s construction and insulation including floors, walls, ceilings, windows, roof, doors and, if any, skylights and shades. Nonresidential loads arise from dwelling’s internal stuffs such as people, lighting, electrical equipment and any other miscellaneous loads.

DESIGN AND ANALYSIS

System Description: Variable Air Volume (VAV) System According to the Owner Project Requirement (OPR) of this year ASHRAE’s competition, each team must make their design calculations by selecting VAV (Variable Air Volume) AHU (Air Handling Unit) for the HVAC system. Because heating, cooling and ventilation needs must be met by using all-air system there is no terminal unit operating with water (Fan Coil) or gas (VRF) in spaces. This selection has some conveniences and difficulties. The task of the team was to get rid of the difficulties as much as possible and increase the effectiveness of the conveniences as well.

The advantages of VAV systems can be summarized as more precise temperature control, lower energy consumption by system fans, less fan noise, relatively low initial cost. A schematic of the general VAV system is shown in Figure 1.

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Figure 1. VAV System Operating Principle

System Selection Components of the AHU system were determined according to the conditioned type (heating cooling) for the zones which have special properties. Mixed Air DX AHU systems were selected for classrooms, offices and conference hall in first, second and third floors. Fresh Air DX AHU system was chosen for workshop. Split air conditioner was selected for server room because it is required to work for 24 hours.

Cassroom Building

DX AHUs operating with VRF outdoor units were designed to use for classroom building (three floors of the building). In these AHUs, rotary heat exchangers were thought to provide heat recovery. Following sections describe these components in detail.

DX AHU with VRF Outdoor Units;

The team, according to literature researches and utilization of experienced people working in this area, determined to design the HVAC system of the building by the AHU units with DX heating and cooling coils instead of conventional ones. DX AHU system operates with VRF outdoor unit that can heat and cool the coils, in this way; there is no need to use –water or air cooled- chiller and boiler system.

The aim of the system selection is that DX AHU is very energy-efficient one, which has higher COP values, compared to AHU with air cooled chiller. There is less energy loss in the fluid pipes going through from AHU to outdoor VRF unit and no need to use the secondary fluid such as water to cool or heat the coil within the AHU. The outdoor VRF unit is modular. Therefore, more than one outdoor unit can be used or required number of units can be operated or not for required heating or cooling demand building. Consequently, the system can give the quick response to the partial loads; also meet the variable capacity by using inverter technology for compressors and fans. Although the first investment cost is a little bit higher, the operating and maintenance cost is really lower than conventional systems. While it is a possibility for conventional water coil of AHU to have a freezing risk, there is no this kind of risk for these systems. Because the AHU and VRF units were planned to be located on the roof of the building, the noise criterion did not affect the design noticeably. The important thing that the team would suffer to design the system was the outdoor temperature air being too high that can be seen as 106oF in Doha, Qatar. However, VRF units can seamlessly operate up to 109.4oF.

Workshop Building

AHU system with 100% fresh air was selected for the workshop. The workshop is negatively pressurized with respect to neighboring areas. This is accomplished by exhaustion of indoor air at a higher cfm than supplying air. Therefore, any potential of dust particulates from equipment, any air pollutants such as welding fume are prevented from moving workshop to other area of the building.

Workshop was separated into two departments to prevent the fire that will cause from the meet of welding fume and sawdust. As seen, the north side was used for carpentry, south for welding area, also the exhaust connections are separated to avoid an explosion inside the exhaust system. AHU unit supplying air to workshop and exhaust system are completely separate from each other, any type of heat recovery unit was not possible to use.

Photovoltaic System Analysis The solar energy part of the system was applied with the photovoltaic solar panels to meet the %5 total electricity demand of HVAC system. The CIGS Thin-Film Solar Cells were chosen so that to minimize the system losses.

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Thin-film solar cells working depend on traditional silicon-wafer cells. Thin-film PV solar modules employ a very thin layer of semiconductor and the most common semiconductors used in thin-film solar module production are: amorphous silicon (a-Si). cadmium telluride (CdTe). and copper indium gallium di-selenide (CIGS).

The solar radiation average daily per square meter in every month; energy and price per month was calculated. The results were presented in Table 3.

Table 3. Solar Radiation

Month Solar Radiation (kWh/m2/ month)

Energy (kWh)

Price ($)

January 5.53 5890 235.6 February 6.41 5780 231.2 March 7.07 6720 268.8 April 6.91 6370 254.8 May 7.37 6510 260.4 June 7.49 6420 256.8 July 7.17 6240 249.6 August 7.27 6150 246 September 7.44 6200 248 October 7.19 6750 270 November 5.97 5750 230 December 5.50 6000 240 Annual Total 81.32 74.780 2991.2

Average 74.780 kWh of energy was obtained from the Thin-Film PVs and the price was 2991.2$ per year. The energy produced by the system was designed so that to satisfy the 5% energy of components of the system. During the setup of the solar system 228 pieces of Solar Frontier 165 W were used. Installation cost of this PV system was 56.430$.

The cost of integration of PV system into entire the system was only calculated and recommended. It was not added to total cost at the end. This means that the decision of installation of PV system would be up to the investor.

ENERGY AND ECONOMIC ANALYSIS

Evaluation of System Alternatives The VAV system design is a required criterion according to OPR. However, the resource specification for AHUs was up to the team. There could be several types of resource for this purpose. Three of them became prominent after assessment stage and the team determined to use DX AHUs operating with VRF outdoor unit considering several factors such as energy efficiency, cost analysis, sustainability and also feasibility. Other system alternatives were briefly explained below and their detailed comparison depending on electricity consumption, water use, maintenance and first investment cost in cost analysis section.

AHU with Air Cooled Water Chiller

AHU with air cooled chiller system is a compact system; a breakdown on it affects the entire HVAC system of the building and heating or cooling processes can stop. Air-cooled chillers are refrigeration devices of a sort. They utilize a process of evaporation and condensation within a closed system to chill the surrounding air.

AHU with Adiabatic Cooling Integrated Air Cooled Water Chiller

AHU with adiabatic cooling integrated air cooled water chiller is shown in this figure. The typical evaporative pad cooling system draws outside air into the building through wet vertical pads. In this system outside air is taken by the air cooler through wet pads to cool the refrigerant gas. As air flows past the moist pad surfaces, some of the moisture evaporates into the air stream. Heat is withdrawn from the air during this process and the air leaves the pad at a lower temperature with higher moisture content.

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Cost Analysis While making first investment cost and annual total cost analysis, three systems, which are DX AHUs operating with VRF outdoor units, conventional AHUs operating with air cooled water chiller and conventional AHUs operating with adiabatic cooling integrated air cooled water chiller, were compared. Electricity consumption, maintenance and water usage costs were taken into account while calculating the annual total operating cost. Electricity consumption price per kWh and potable water price per m3 were taken according to Qatar General Electricity & Water Corporation related tariffs [7]. The results are shown in Figure 2 and 3. The outline of water cost calculation during system operation is given below. Also, adiabatic cooling process in Doha, Qatar air conditions is shown with the psycometric chart.

P1 P2 t (dry bulb) 41,1°C t (dry bulb) 29,6°C t (wet bulb) 22,9°C t (wet bulb) 22,5°C t (dew point) 14,5°C t (dew point) 19,6°C Pressure 101207,7Pa Pressure 101207,7Pa RH 21,0% RH 55,0% g 0,0103kg/kg g 0,0143kg/kg Enthalpy 67,8kJ/kg Enthalpy 66,4kJ/kg Density 1,115kg/m³ Density 1,155kg/m³ Airflow 110,0l/s Airflow 110,0l/s

Absolute Humidity Difference of dry air, w= 0.0143 – 0.0103 = 0.004 kg H2O / kg dry air Average Air Requirement to Cool Adiabatic Cooling Integrated Water Chiller, Q = 50 m3 / s Q = (50 m3 / s) * (3600 s / hr) = 180000 m3 / hr M = ρ * Q = 1.2 (kg / m3) * 180000 m3 / hr = 216000 kg air/ hr Water Requirement in mass flow rate = m* w = 216000 kg air/ hr * 0.004 kg H2O / kg dry air = 864 kg H2O / hr ρ water= 1000 kg/m3 Water Requirement in volume flow rate= (864 kg / hr) / (1000 kg / m3) = 0.864 m3 / hr Annually Working Time for Adiabatic Cooling Integrated Water Chiller = 2448 hr/year Annual Water Requirement = (0.864 m3 / hr) * (2448 hr/year) = 2115 m3/year Water tariff = 1.88 $ / m3 Annual Water Cost = (2115 m3/year) * (1.88 $ / m3) Annual Water Cost = 3976 $/year It can be seen that adiabatic cooling integrated chiller has this much more annual water cost. However, the annual electricity consumption has lower than conventional one as shown in table 4.

Table 4. Electricity Consumption Comparision between Chiller Systems

System Electricity Consumption, kW/h Air Cooled Water Chiller 248,19 Adiabatic Cooling Integrated Water Chiller 150,5

10

20

30

405060708090100Relative humidity (%)

-20Enthalpy (kJ/kg) -100

1020

30

40

50

60

70

80

90

100

110

120

130

140

150

P1

P2

-40 -35 -30 -25 -20 -15 -10 -5 0 5 10 15 20 25 30 35 40 45 50 55 60Dry bulb temperature (°C)

0,0000

0,0050

0,0100

0,0150

0,0200

0,0250

0,0300

0,0350

0,0400

0,0450

Absolute humidity (kg/kg)

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Figure 2. First Investment Cost Comparison Figure 3. Annual Total Operating Cost Comparison

Although the first investment cost of conventional air cooled chiller has pretty much the same one of the adiabatic cooling integrated air cooled chiller, because its operation cost is the maximum among three, it is not evaluated to select the final system. Because it was a must to select an VAV system in accordance with the rule of competition, the selection of AHUs was unavoidable and DX VRF outdoor units were determined by considering its long-term return. However, if the team had to determine a system by comparing conventional and adiabatic cooling integrated chiller, the adiabatic one would be more optimum choice.

Final decision was made by comparing the DX AHU+VRF outdoor unit and Water AHU+ adiabatic cooling integrated air cooled chiller. This was carried out by NPV analysis and life cost analysis was obtained after NPV (Net Present Value) calculations. This analysis was given in Table 4.

NPV and Life Cycle Cost Analysis

NPV method is the difference between the present value of cash inflows and the present value of cash outflows. NPV is used in capital budgeting to analyze the profitability of an investment or project. In order to accomplish this analysis, first investment cost difference between systems was calculated as 62837$ which shows that DX system is more expensive that much. However, considering electricity consumption, maintenance cost difference and water usage costs, it was seen that DX AHU system has 6713$ lower annual total operating cost. These differences were arisen from the selection of outdoor units that operates of the AHUs. Water usage amount of adiabatic cooling integrated chiller was determined by hand calculations. Another important thing is that inflation rate was taken as 3% according to OPR and analysis was done for 50 years. Table 5 briefly shows the results for only 15 years.

Table 5. NPV Analysis for DX AHU

According to NPV analysis, the system constituting of DX AHUs and VRF outdoor units can amortize itself in 12 years and starts to bring in after this year. When the life of the building was thought as 50 years, amortizing time is quite short. Figure 4 shows the life cost analysis result based on outdoor units which are VRF and adiabatic cooling integrated air cooled chiller.

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Figure 4. Life Cycle Cost Analysis

CONCLUSION

The aim was to design the HVAC System with desired requirements declared at OPR. While doing this, we consider the ASHRAE 55. 62.1 and 90.1 Standards. To obtain the accordance of ASHRAE Standards and software data, Hourly Analysis Program-Carrier (HAP 4.62) was used. Obtained data guide us about ductwork requirements and to select the appropriate equipment. The shaft area was defined as mechanical room and the duct design was provided with a different method. By using this method a duct never jump over another duct and pressure drop was avoided. The team decided to compare three systems that have some advantages according to each other. Their system comparison and cost analysis showed as the most available option for final decision.

The instruction department for welding and woodworking was examined specially. AHU with 100% fresh air was selected for this department. Area usage for equipment and the local exhaust systems were determined according to the needs and fire possibility. So, the equipment with sensors was selected. Also general ventilation was designed and workshop was provided as negative pressure to avoid the negative effect of poisoned welding fume and sawdust.

The system is selected as DX AHU system operates with VRF (Variable Refrigerant Flow) outdoor unit(s) that can heat and cool the coils. In this way; there is no need water or air cooled- chiller and boiler system. DX AHU is very energy-efficient compared to air cooled water chiller. Because of VRF units being modular and heating or cooling demand of the building is variable. More than one outdoor unit can be used and required number of the units can be operated and other can be off .Finally, the team performed the design by providing the needs of the building with energy efficient criteria, ASHRAE standards and supplied budget.

REFERENCES 1. 2009 ASHRAE Handbook - Fundamentals (I-P Edition) (2009). American Society of Heating. Refrigerating and

Air-Conditioning Engineers. Inc. 2. ASHRAE 55-2004 (2004) Thermal Environmental Conditions for Human Occupancy. American Society of

Heating. Refrigerating and Air-Conditioning Engineers. Inc. 3. ASHRAE 62.1-2004. (2005). Ventilation for Acceptable Indoor Air Quality. American Society of Heating.

Refrigerating and Air-Conditioning Engineers. Inc. 4. ASHRAE 90.1-2004. (2004) Energy Standard for Buildings Except Low-Rise Residential Buildings. American

Society of Heating. Refrigerating and Air-Conditioning Engineers. Inc. 5. www.nederman.com 6. http://airware.systemair.se/ (to design AHUs) 7. https://www.km.com.qa/CustomerService/Pages/Tariff.aspx

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[Abstract:0091][Project Management in Structure Design and Applications] NUMERICAL ANALYSIS OF THERMAL CONDUCTIVITY OF AUTOCLAVED AERATED

CONCRETE ON DIFFERENT POROSITY RATIO

Battal Doğan, Hüsamettin Tan, bttldgn@gmail.com, husamettintan@kku.edu.tr,

Department of Mechanical Engineering,Kırıkkale University,Kırıkkale Turkey

ABSTRACT The autoclaved aerated material that is used as building envelope has a very important place in construction industry. In this work, the effect of different porosity ratio on thermal conductivity has been examined as numerically in one dimensions. When numerical analysis was processed, geometric model of autoclaved aerated concrete was formed and handled as two dimensional planes. Thermal conductivity values, in which quartzite does not change while air bubbles in the aerated concrete change depending on heat, have been used. When heat transfer was calculated with a program, which is based on finite element. Calculated thermal conductivity values were given as comparatively with TS 825 standards. Key Words: Autoclaved aerated concrete, Thermal conductivity, Porosity 1.INTRODUCTION The depletion of energy sources in the world with day after day increases the importance of using energy resources more efficiently. Energy sources are spent mainly to meet heating and cooling requirements of houses in our country as the situation is the same all around the world. The energy performance of a building highly depends on the thermal conductivity of the building materials1. The highest rate of energy loss is in the outer wall which is a construction element. Therefore, construction elements which have low thermal conductivity are preferred in the external wall. There are many physical, chemical, and geometrical factors influencing thermal properties of porous building materials, including density, temperature, moisture content, microstructure, and composition as well2. Autoclaved aerated concrete is a high porous construction material and the proportion of macro and micro level porous in all construction has been changing between 69-88 %3-4. Therefore; it has low thermal conductivity value and density.Autoclaved aerated concrete with certain density value reaches its lowest thermal conductivity value in its dry situation but while the moist value in material is increasing, the thermal conductivity of the material starts to increase, too5. Moreover, autoclaved aerated concrete has been recognized for its superior performance in thermal insulation and sound insulation characteristics due to its porous structure6-7. The relationship between porosity and thermal conductivity in building materials was determined both experimentally and numerically8. When internal structure images of autoclaved aerated concrete were investigated, air bubbles, which compose of porosity, accepted in spherical geometry9. Autoclaved aerated concrete is relatively homogeneous when compared to normal concrete10.A great number of experimental studies were made in literature that usually identified thermal properties of autoclaved aerated concrete11-20. Moreover, effect of thermal conductivity was also measured by making various applications to outer side of autoclaved aerated concrete specimen21-22. In this study, various electron microscope images of autoclaved aerated concrete are investigated. When designing the internal structure of autoclaved aerated concrete (100mm x 100 mm) in size, porosity ratio is obtained by placing air bubbles, which are in circular geometry, randomly. One dimensional numerical solutions were made in a program,which is based on finite element, that was used to internal construction design. In literature, it was mainly made experimental studies, numerical solutions of various porosity ratio were not encountered. At the end of this study, comparison of thermal conductivity value is made with TS 825 standard values. 2. METHOD Autoclaved aerated concrete is composite, which is composed of air and quartzite. Porosity of autoclaved aerated concrete is a measure of the amount of air23.If porosity increases, thermal conductivity of composite will decrease because the thermal conductivity of air is lower than quartzite. In the study, a lot of electron micrographs have been examined for gas concrete material, and it has been identified that the general view is similar to that of internal structure photograph given in Fig. 1. Air bubbles in autoclaved aerated concrete usually were seen to be about circular geometry in the investigated images.

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Fig. 1. Internal structure image of aerated concrete2

Porosity for a composite is voids ratio in unit volume of the material. The escalation of porosity affects density of the material. As porosity of autoclaved aerated concrete increases, density of the material decreases. The variation of density depending on the porosity is given in the Table-1.

TABLE-1 CHANGE OF DENSITY DEPENDING ON THE POROSITY RATIO

In this work, 4 different models have been created by considering the values in the Table-2. While models are being created, random placement is made and rate is determined in accordance with values in the literature.

TABLE-2

THE MODELS USED IN THIS STUDY

Models Mass Per Unit Volume

(kg/m3)

Porosity Ratio

(%)

Model-1 300 88,09

Model-2 400 85,01

Model-3 500 81,05

Model-4 700 73,07

Mass Per Unit Volume (kg/m3)

Porosity Ratio (%)

310 - 400 88 – 85 410 - 500 85 – 81 510 - 600 81 – 77 610 - 700 77 - 73 710 - 800 73 - 69

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Considering porosity ratio that corresponds to mass per unit volume given in the Table-2 for 4 different models, internal structure images are given in the Fig. 2. Random placement was made in the internal structure design and effect of different distribution was taken into account for the same porosity ratio. In the case where the shape of air bubbles are accepted as circular, location in the construction will not affect thermal conductivity in the one dimensional condition. But in the same solution plane, when air bubbles were created as square, triangular and circular separate models, large deviations from experimental measurement were determined in the results of square and triangular geometry.

a) Model-1 b) Model-2 c) Model-3 d) Model-4

Fig. 2. Internal structure image of model Fourier heat

conduction law, which is given in the eqn. 1, has been used in the numerical determination of thermal conductivity. Solution was implemented as one-dimensional. While numerical solutions were made, constant surface temperature boundary conditions, which is given in the Fig. 3, were used.

. . (1)

= 0 303o C 273o C y

= 0 x Fig. 3. Demonstration of boundary conditions

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While numerical solution was made, thermal conductivity value of autoclaved aerated concrete was calculated by using flow diagram that is given in the Fig. 4.

Fig. 4. Flow diagram for numerical solution

In this work, by using the representative solution plane geometry of autoclaved aerated concrete specimen drawn in solidworks, which is given in the Fig. 5a. Random circles, which are in various dimensions, were located to interior of square element with a view to obtain porosity ratio. Porosity ratio was determined by calculating percentage of summation of circle areas in square area. Circles, which are in the interior of geometric draw in the Fig. 5b, were defined as air and and the rest was defined as quartzite.

a) Drawing of geometric model b) Image of areas in

ansys

Fig. 5. Image of model in the ansys

Calculation of thermal conductivity value

Making of numerical solution

Defining of boundary conditions

Mesh

Entering of thermal conductivity values by defining areas

The transferring of drawn geometry to the Ansys

Creation of geometry in the Solidworks

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While material properties were defined in Ansys, values in the Table-3 which gives the change in thermal conductivity value depending on the temperature for air bubbles, which forms porosity, were used. The remaining area of the space by assuming that it represented quartzite was taken as 3 W/m K, which is an average value.

TABLE-3 CHANGE OF THERMAL CONDUCTIVITY OF AIR DEPENDING

ON THE TEMPERTAURE

After defining properties of autoclaved aerated concrete material to the program, mesh operation was made by dividing it into triangular structure elements. By using different mesh density in certain areas, which is given in the Fig. 6, the accuracy of the analysis is increased. The change in the number of nodes were considered in the mesh operation.

Fig. 6. The demonstration of mesh structure

After mesh operation and defining boundary condition, which is given in the Fig. 3, solution was made and the amount of one-dimensional heat transfer was obtained on nodal points. Heat flux and amount of heat per unit time in the solution plane was calculated using the finite elements in the plane perpendicular directions. The amount of average heat transfer was determined by taking average of these values. Similar temperature curves that belongs to 4 models, was obtained from solution results, which are given in the Fig. 7.

a) Model-1 b) Model-2

Temperature (K)

Thermal Conductivity Value(Air) (W/m.K)

100 0,00934 200 0,0181 400 0,0338 500 0,0407

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c) Model-3 d) Model-4 Fig. 7. Temperature distrubition for models 3. RESULTS Geometry of internal structure is one of the most important parameters affecting the thermal conductivity. In this study, circular geometry which entertains similarities with electron micrographs showing the internal view of aerated concrete material has been preferred. As it has been observed during the analysis, gas concrete’s thermal conductivity value changes depending on its porosity ratio. If porosity ratio of autoclaved aerated concrete increases, value of density and thermal conductivity will decrease. Results of the thermal conductivity value obtained in this work were given in the Table-4. When Table-4 was examined, as dry bulk density is increasing, thermal conductivity value is increasing. Because when the bulk density increases, the porosity decreases. The thermal conductivity value of air representing the porosity is lower than the aerated concrete material.

TABLE-4

THE THERMAL CONDUCTIVITY VALUES OBTAINED IN THIS STUDY

As the

numerical solutions were carried out, the effect of nodal point numbers is given in the Table-5. For the analyzed Model (1) and Model (2), heat transmission has been calculated by using nodal points between 10.000-40.000. Since the change has reached to a negligible point after the value of 40.000 for the models analyzed, solutions have been carried out by using the approximate value of 40.000 node number for all the models. TABLE-5

THE EFFECT OF NODAL POINT NUMBERS

Models qort (W) Calculated Effective Thermal Conductivity Value(W/m.K)

Model-1 0,22246 0,07 Model-2 0,34219 0,11 Model-3 0,49943 0,17 Model-4 0,61708 0,20

Numerical Models Nodal Point Number Calculated ‘q’ (W)

Model-1

14600 0,22241

28900 0,22244

39970 0,22246

Model-2

10800 0,34213

23460 0,34216

39980 0,34219

Model-3 39560 0,49943

Model-4 40200 0,61708

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The experimental assessment based values of TS 825 standards with which numerical results gathered at the end of the study will be compared, have been given in the Table-6. The value of porosity ratio of the autoclaved aerated concrete material is between 70-80 %, which is used in the market. Therefore, thermal conductivity value is (0.22), which correspond to the value of 73.07% porosity in the TS825 standards. Percentage error is about 10 % for (0.20) values obtained from this study.

TABLE-6 TS 825 AND TS EN 1745 STANDARD VALUES

Thermal conductivity values, which were calculated in the study, have been given in the Fig. 8 in comparison with values of TS 825 standards. It is seen that values obtained from studies and the standards based on assesment coincide with a certain error rate.

Fig. 8. The comparative chart of thermal conductivity values

0

0,05

0,1

0,15

0,2

0,25

300 400 500 700

Ther

mal

Con

duct

ivity

(W/m

K)

Density (kg/m3)

Calculated values

TS 825

Dry Bulk Density (kg/m3)

Thermal Conductivity Value

(W/m.K) (TS 825) 300 0,08 400 0,13 500 0,16 700 0,22

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4.REFERENCES

1. I. Budaiwi, A. Abdou, M. Al-,homoud, Variations of thermal conductivity of insulation materials under diffirent operating temperatures:impact on envolope-induced cooling load, Journal of Archaelogicial Engineering 8 (4) (2002) 125-132.

2. Hong-Qing Jin, Xiao-Li Yao, Li-Wu Fan, Xu Xuc, Zi-Tao Yu,Experimental determination and fractal modeling of the effective thermal conductivity of autoclaved aerated concrete: Effects of moisture content,International Journal of Heat and Mass Transfer 92 (2016) 589–602

3. Y.E. Cicek, Ph.D. Thesis, Istanbul Technical University, Turkey (2002). 4. S. Kartal, Ph.D. Thesis, Trakya University, Turkey (2001). 5. Z. Pehlivanli, R. Calin, I. Uzun, Effect of moisture and temperature on thermal conductivity of G2/04 class

autoclaved aerate concrete, Asian J. Chem. 22 (2010) 4104–4110. 6. A, Kılıç, C.D.Atis, E.Yaşar, Özcan, High-strenght ligtweight concrete made with scoria aggregate containing

mineral admixtures, Cement Concrete Research 33 (10) (2003) 1595-1599. 7. G.M. Glenn, G.M. Gray, W.J. Orts, D.W. Wood, Starch-based lightweight concrete: effect of starch source

processing method and aggregate geometry, Industrial Crops and Products 9 (2) (1999)133-144. 8. A. Bouguerra, A. Ait-Mokhtar, 0. Amiri, M. B. Diop, “Measurement of thermal conductivity, thermal diffusivity

and heat capacity of highly porous building materials using transient plane source technique” Heat Mass Transfer, 28, 1065-1078, (2001).

9. Kadashevich, H.-J. Schneider, D. Stoyan, “Statistical modeling of the geometrical structure of the system of artificial air pores in autoclaved aerated concrete”, Cement and Concrete Research, 35, 1495, (2005)

10. N. Narayanan, K. Ramamurthy, Structure and properties of aerated concrete: a review, Cement Concrete Composite 22 (2000) 321–329.

11. A.D. Stuckes, A. Simpson, The effect of moisture on the thermal conductivity of aerated concrete, Build. Serv. Eng. Res. Technol. 6 (1985) 49–53.

12. J.P. Laurent, An estimation model for the dry thermal-conductivity of autoclaved aerated concrete, Mater. Struct. 24 (1991) 221–226.

13. J.P. Laurent, C. Guerrechaley, Influence of water-content and temperature on the thermal-conductivity of autoclaved aerated concrete, Mater. Struct. 28 (1995) 464–472.

14. C. Boutin, Thermal conductivity of autoclaved aerated concrete: modelling by the self-consistent method, Mater. Struct. 29 (1996) 609–615.

15. M.S. Goual, A. Bali, M. Quéneudec, Effective thermal conductivity of clayed aerated concrete in the dry state: experimental results and modelling, J. Phys.D Appl. Phys. 32 (1999) 3041–3046.

16. D. Gawin, J. Kosny, A. Desjarlais, Effect of moisture on thermal performance and energy efficiency of buildings with lightweight concrete walls, in: 2000 ACEEE Summer Study on Energy Efficiency in Buildings Efficiency, Panel 3:Commercial Buildings: Technologies, Design, and Performance Analysis,August 20–25, 2000, Pacific Grove, California, USA, 2000.

17. B. Bhattacharjee, S. Krishnamoorthy, Permeable porosity and thermal conductivity of construction materials, J. Mater. Civ. Eng. 16 (2004) 322–330.

18. M. Albayrak, A. Yörükog˘lu, S. Karahan, S. Atlıhan, H.Y. Aruntas_, _I. Girgin, Influence of zeolite additive on properties of autoclaved aerated concrete,Build. Environ. 42 (2007) 3161–3165.

19. M. Jerman, M. Keppert, J. Vy´ borny´, R. Cˇerny´ , Hygric, thermal and durability properties of autoclaved aerated concrete, Constr. Build. Mater. 41 (2013) 352–359.

20. M. Campanale, M. Deganello, L. Moro, Effect of moisture movement on tested thermal conductivity of moist aerated autoclaved concrete, Transp. Porous Media 98 (2013) 125–146.

21. Soon-Ching Ng, Kaw-Sai Low, Thermal conductivity of newspaper sandwiched aerated lightweight concrete panel, Energy and Buildings 42(2010) 2452-2456

22. Kaw-Sai Low, Soon-Ching Ng, and Ngee-Heng Tioh, Thermal conductivity of soil-based aerated lightweight concrete, KSCE Journal of Civil Engineering (2014) 18(1):220-225.

23. N. Narayanan, K. Ramamurthy,Microstructural investigations on aerated concrete,Cement and Concrete Research 30 (2000) 457± 464

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[Abstract:0095][Heating, Climatization and Air-conditioning Applications in Buildings] EFFECT OF ICE THERMAL ENERGY STORAGE SYSTEM ON COOLING COST

IN A SHOPPING CENTER

Dogan Erdemir1,* and Necdet Altuntop1

1Erciyes University, Engineering Faculty, Department of Mechanical Engineering, 38039, Kayseri, Turkey *Corresponding email: erdemir@erciyes.edu.tr

SUMMARY This paper investigates the thermodynamic and economic analysis of ice thermal energy storage system (ITES) in a shopping center. ITES can be used to shift electric demand from high demand period to low demand period. The electric load of cooling can be shift to peak-off period. Thus, cooling cost can be reduced. In present study, ITES usage in a shopping center is investigated in the terms of cooling cost. The shopping center is located in Ankara, Turkey. Also it has 15,000 m2 closed-area. The total cooling load per day for the shopping center 6,589 kW. Two different ITES usage scenario (full storage and partial storage) have been tested. At the end of the study, electric consumption cost saving has been presented. In normal cooling mode, without ITES, cooling cost is 867 USD. When ITES is used, cooling cost for partial storage and full storage are 560 and 394 USD, respectively. INTRODUCTION Thermal energy storage (TES) is to store energy for later usage. TES is an improved technology, which is receiving renewed consideration today in commercial and institutional building applications. Advantages of the TES are [1];

Reduced energy cost, Reduced energy consumption, Improved indoor air quality, Increased flexibility of operation, Low initial and maintenance costs, Reduced equipment size, Efficient and effective utilization of equipment, Conservation of fossil fuels, Reduced pollutant emissions, Shifting peak load to off-peak periods.

TES has the advantage of an efficient use of energy by reducing the imbalance of an electric demand between daytime and nighttime in summer. It is classified as sensible and latent thermal storage systems. The latent thermal storage energy system is superior to the sensible thermal energy storage system since the former reduces the instillation area and the expense due to a large thermal storage capacity and constant phase change temperature during freezing and melting processes. Many cities throughout the world are faced with increasingly high energy cost. Often, these costs are not directly incurred due to the energy used, but rather due to the costs associated with overall demand for electricity. Many electrical utility companies are having difficulties in maintaining sufficient capacity to serve the peak demand of their customers, while at the same time supplying reasonably priced electricity. One significant method for leveling the electrical demand in commercial buildings is to shift the electrical loads incurred by heating, ventilating and air conditioning equipment to periods lower electric usage. This load shifting can be supplied by TES. The most common method to use the low electricity rates during off-peak periods is using ice thermal energy storage (ITES) systems. The basic of ITES is in-expensive off-peak electricity is utilized during the off-peak hours to produce ice, and during the day this ice store is melted by absorbing the heat from buildings needing cooling [2]. Ismail and Henriquez [3], investigated spherical encapsulated ITES. They used 1-D numerical model and validated their result with experimental study. They used different capsule materials. At the end of their study PVC is the most suitable material for capsule in the terms of heat transfer and production cost. Kousksou et al. [4] researched tank position. They emphasized that tank should be vertical for better performance. MacPhee et al. [5] researched different capsule geometry. Sphere capsule supplied better performance. Chaichana et al. [6] investigated the economic performance of ice-on-coil systems. ITES system saved 55 % in full storage, 15 % in partial storage. Cho and Choi [7], used paraffin instead of water inside the capsule. Erek and Dincer [8], investigated temperature changing through the flow direction. They emphasized that heat transfer coefficient decreased through the flow direction. Ezan et al. [9], developed thermal resistant mesh model for heat transfer analysis. Ismail et al. [10], used finite difference model for spherical encapsulated ITES. Chen and Yue [11,12] applied porous medium equation into encapsulated ITES. Ryu et al. [13] investigated super-cooling effect in ITES. Super-cooling was seen in inlet and outlet of the tank. Calvet et al. [14] added graphite inside the capsule. Amin et al. [15] developed a new model consists buoyancy forces and density changing with the temperature.

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There are many studies about ITES systems. Most of these studies are about enhancing storage capacity or shorten ice-making time. There is no clear study about ITES usage in shopping center. In present study, effect of the ITES usage on cooling cost in a shopping center has been investigated. Thermodynamic and economic analyses have been performed. MATERIALS AND METHODS Systems description The shopping center investigated in present study is located in Ankara, Turkey. Ankara has arid climate. Outer view of the shopping center is seen in Fig. 1. The shopping center has 15,000 m2 closed-area. Metrological information for Ankara between 1950 and 2014 is given in Table 1[16]. As seen from Table 1, cooling is a need for Ankara between 4-10th months in a year. Electric load of the shopping center is seen in Fig. 2. As seen from Fig. 2, Peak time cooling load has been shifted off-peak time. Chiller usage is also given in Table 2. Cooling cost has been decreased by ITES. Three different ITES usage scenarios have been theoretically applied to the shopping center cooling system. These scenarios are

1. Partial storage for peak time cooling: in this case, just cooling load in peak time is stored, 2. Partial storage for entire cooling load: in this case, a portion of all cooling load is stored, 3. Fully storage: in this case, all cooling load is stored by ITES system.

Figure 1. Outer view of the shopping center

Electric load of the shopping center is seen in Fig. 2. As seen from Fig. 2, Peak time cooling load has been shifted off-peak time. Chiller usage is also given in Table 2. Cooling cost has been decreased by ITES. Three different ITES usage scenarios have been theoretically applied to the shopping center cooling system. These scenarios are

4. Partial storage for peak time cooling: in this case, just cooling load in peak time is stored, 5. Partial storage for entire cooling load: in this case, a portion of all cooling load is stored, 6. Fully storage: in this case, all cooling load is stored by ITES system.

Table 1. Metrological data for Ankara between 1950 and 2014 Months 1 2 3 4 5 6 7 8 9 10 11 12 Mean

temperature 0.4 1.8 6.0 11.4 16.0 20.1 23.5 23.3 18.7 12.9 7.0 2.6

Mean the highest temp. 4.4 6.5 11.6 17.3 22.2 26.6 30.2 30.3 26.0 19.8 12.8 6.6

Mean the lowest temp. -3.0 -2.2 0.9 5.6 9.6 13.0 15.9 16.0 11.8 7.2 2.4 -0.7

Mean hours of sunshine 2.5 3.5 5.2 6.4 8.4 10.2 11.3 11.6 9.2 6.5 4.4 2.3

Mean rainy day 12.3 10.9 11.0 11.7 12.6 8.8 3.8 2.7 3.9 6.9 8.5 11.6

The highest temp. 16.6 20.4 27.8 31.1 33.0 37.0 41.0 40.4 36.0 33.3 24.4 20.4

The lowest temperature -22.4 -22.2 -19.2 -6.7 -1.6 3.8 4.5 6.3 2.5 -5.3 -13.4 -18.0

Schematic view of the ITES system is seen in Fig. 3. All operating temperatures have been given in Fig. 3. In Fig. 3, the green line is charging (ice-making) line, the red line is discharging (ice-melting) line and the blue line is normal cooling line (storing and normal cooling with cooling system). HTF is circulated with 130 m3/h flowrate. Thermodynamics analysis A thermodynamic assessment is described of an encapsulated ITES system in this section. Some assumptions are used to simplify analysis. There are;

Storage medium temperature stays constant at the melting point. Fluids are frictionless and pumping power is zero.

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The net effect of these assumptions is to increase the apparent exergy efficiencies, but probably not their relative values. Thus, the result should be valid in ordering the performances of competitive systems in an optimization process.

Figure 2. Cooling load and chiller usage

Figure 3. Schematic view of the ITES system in charging, discharging and storing

Energy Balance of Storage Capsule: Schematic view of the charging, discharging and storing periods is given in Fig. 4. Energy balance for entire cycle of a cold capacity storage can be written as;

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Figure 4. Schematic view of the charging, discharging and storing periods

, (1)

Where “cold input” is gathering cold energy from HTF during charging period, “cold recovered” is gathering cold energy from storage medium during discharging period, “cold loss” is heat gain from environment to the storage medium during all periods and “cold accumulation” is the resultant change in energy of the storage fluid. For the systems given in Fig. 4, overall energy balance is,

, (2) Where Ha, Hb, Hc and Hd are enthalpies of the flows at points a, b, c and d given in Fig. 4. Ql is the heat gain during charging, discharging and storing periods. ΔE is the change in energy of the storage fluid. If ITES undergoes a complete cycle, ΔE=0. So, the energy changing is also defined as:

, (3)

The energy content of the solid and liquid portions of the storage fluid can be evaluated separately and summed as follows:

1 , (4)

Here, us and ul are specific internal energies of the solid and liquid portions of the storage fluid, respectively; uo is specific enthalpy at environmental conditions; F is the melted fraction and m is the mass of storage fluid. Cold input can also be expressed as:

, (5)

Here, “ ” is the mass flow rate for HTF, “C” is the specific heat for HTF at the mean temperature. Cold output can be written as:

, (6) Heat loss is calculated as follows:

, (7)

Where “Rt” is overall thermal resistance, A is total surface area for the storage tank and ΔT is temperature difference between mean HTF temperature and environment. Exergy Balance of Storage Capsule: Exergy balance for a cold capacity storage undergoing a complete cycle of charging, storing and discharging can be written as: . . . . . (8)

Here, Ex. is abbreviation of the exergy. Exergy loss is the quantity of exergy dissipated from the storage fluid through heat gain from the environment. Exergy consumption is the exergy loss due to irreversibility. The overall exergy balance according to Fig. 4 is

, (9)

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Where “εa”, “εb”, “εc” and “εd” are the exergy flows at points a, b, c and d in Fig. 4, “I” is the exergy consumption due to irreversibility and “ ” is the change in non-flow exergy of the storage fluid. If the ITES is a complete cycle, =0. The exergy change can also be expressed as:

(10)

The exergy content of the storage fluid having solid and liquid phases can be determined as follows:

1 (11) Where To denotes the temperature of the reference environment; ss and st are the specific entropies of the solid and liquid portions of the storage fluid, respectively; so is the specific entropy of the storage fluid at environmental conditions. The exergy transfer associated with the charging heat transfer fluid can be expressed as:

ln (12) Energy and Exergy Efficiencies: Energy efficiencies for ITES can be calculated as:

∑

∑ (13)

(14)

1 (15)

(16) Exergy efficiencies for ITES system can be calculated as:

∑

∑ (17)

(18)

1 (19)

(20) RESULTS AND DISCUSSIONS ITES systems defined in previous section has been investigated in the terms of thermodynamic and economical. The shopping center cooling load and chiller usage are given in Table 2. All analyses have been performed according to these values. Hourly heat gain for storage tank is 37 kW. This heat gain amount should be charged into the storage tank. Also 20% fresh air cooling load has been considered. Green color represents off-peak, yellow color represents day-time and red color represents peak-time periods in Table 2. In full storage mode, all cooling load and heat gain are charged in off-peak time. Since off-peak period is short, this system requires larger-capacity chiller. This situation increases initial investment cost. In peak load storage mode, the cooling load and the heat gains of storage in peak-time is stored in off-peak period. In partial storage mode, a portion of the cooling load and the heat gains of storage tank is stored in off-peak time. In peak load and partial storage modes, initial investment cost is lower than fully storage mode and without ITES system case in the terms of chiller cost.

Table 2. Hourly cooling load and chiller usage for three different ITES usage scenarios

Hours Cooling

load (kW)

Normal cooling without

ITES (kW)

Full Storage (kW)

Peak Load Storing (kW)

Partial Storage (kW)

00:00 - 01:00 0 0 1079 547 621 01:00 - 02:00 0 0 1079 547 621 02:00 - 03:00 0 0 1079 547 621 03:00 - 04:00 0 0 1079 547 621

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04:00 - 05:00 0 0 1079 547 621 05:00 - 06:00 0 0 1079 547 621 06:00 - 07:00 0 0 0 0 0 07:00 - 08:00 0 0 0 0 0 08:00 - 09:00 16 16 0 16 16 09:00 - 10:00 128 128 0 128 128 10:00 - 11:00 329 329 0 329 329 11:00 - 12:00 485 485 0 485 485 12:00 - 13:00 512 512 0 512 560 13:00 - 14:00 624 624 0 624 600 14:00 - 15:00 742 742 0 742 600 15:00 - 16:00 791 791 0 791 600 16:00 - 17:00 830 830 0 763 600 17:00 - 18:00 831 831 0 0 0 18:00 - 19:00 847 847 0 0 0 19:00 - 20:00 678 678 0 0 0 20:00 - 21:00 650 650 0 0 0 21:00 - 22:00 498 498 0 0 0 22:00 - 23:00 372 372 1079 372 0 23:00 - 24:00 0 0 1079 547 621

Energy efficiencies for charging, discharging and storing periods have been calculated as 97.46 %, 97,40 % and 96.7 %, respectively. Also exergy efficiencies for charging, discharging and storing periods have been found as 72.47 %, 62.49 % and 62.02 %, respectively. As seen from these results, energy analysis hasn’t give a significant information about the system. Because all efficiencies are higher than 96 % and these values are close each other. Besides, exergy analyses have given important information about the system operating. The lowest exergy efficiencies are seen in discharging and storing mode. To enhance the ITES system, these periods should be improved. This situation is same with Rosen et al.’s [1] study. In the terms of cooling cost, daily electric usage cost for normal cooling mode, without ITES is 867 USD. When ITES is used, cooling cost for partial storage and full storage are 560 and 394 USD, respectively. Besides this cost value, initial investment cost should be considered. Initial investment cost and ITES usage scenarios have great impact on payback period. CONCLUSION This study presents the effect of ITES usage on cooling cost in the shopping center. Three different ITES usage scenarios have been applied. Two different ITES usage scenario (full storage and partial storage) have been tested. At the end of the study, electric consumption cost saving has been presented. In normal cooling mode, without ITES, cooling cost is 867 USD. When ITES is used, cooling cost for partial storage and full storage are 560 and 394 USD, respectively. Thus, ITES has supplied 35-45 % saving for cooling cost.

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REFERENCES 1. Rosen, M.A., Dincer I., Pedinelli N., 2000. Thermodynamics Performance of Ice Thermal Energy Storage

Systems. Journal of Energy Resources Technology, 122: 205-211. 2. Beggs C., 1991. The economics of ice thermal storage, Buildings Research and Information, Vol. 19, No. 6, pp.

342-355. 3. Ismail, K.A.R., Henriquez, J.R., 2002. Numerical and Experimental Study if Spherical Capsules Packed Bed

Latent Heat Storage System. Applied Thermal Engineering, 22: 1705-1716. 4. Kousksou, T., Bedecarrats, J.P., Dumas, J.P., Mimet, A., 2005. Dynamic Modelling of the Storage of an

Encapsulated Ice Tank, 25: 1534-1548. 5. MacPhee, D., Dincer, I., Beyene, A., 2012. Numerical Simulation and Exergetic Performance Assessment of

Charging Process in Encapsulated Ice Thermal Energy Storage System. Energy, 41: 491 – 498. 6. Chaichana, C., Charters W.W.S., Aye, L., 2001. An Ice Thermal Energy Storage Computer Model. Applied

Thermal Engineering, 21: 1769 – 1778. 7. Cho, K., Choi, S.H., 2000. Thermal Characterictics of Paraffin in a Shperical Capsule During Freezing and

Melting Processes. International Journal of Heat and Mass Transfer, 43: 3183 – 3196. 8. Erek, A., Dincer I., 2009. Numerical Heat Transfer Analysis of Encapsulated Ice Thermal Energy Storage Systems

with Variable Heat Transfer Coefficient in Downstream. International Journal of Heat and Mass Transfer, 52: 851 – 859.

9. Ezan, M.A., Erek, A., Dincer, I., 2011. Energy and Exergy Analyses of an Ice-on-Coil Thermal Energy Storage System. Energy 36: 6375 - 6386.

10. Ismail, K.A.R., Henquez, J.R., da Silva, T.M., 2003. A parametric study on ice formation inside a spherical capsule. International Journal of Thermal Sciences 42: 881-887.

11. Chen, S.L., Yue, J.S., 1991. Thermal performance of cool storage in packed capsuled for air conditioning. Heat Recovery Systems and CHP, 11: 551-531.

12. Chen, S.L., Yue, J.S., 1991. A simplified analysis for cold storage in porous capsules with solidification. Energy Resources, 113: 108-116.

13. Ryu, H.W., Hong, S.A., Shin, B.C., Kim, A.D., 1991. Heat transfer characterictics of cool-thermal storage systems. Energy 16: 727-737.

14. Calvet, N., Py, X., Olives, R., Bedecarrats, J.P., Dumas, J.P., Jay, F., 2013. Enhanced performances of macro-encapsulated phase change materials (PCMs) by intensification of the internal effective thermal conductivity. Energy 55: 956-964.

15. Amin, N.A.M., Bruno, F., Belusko, M., 2014. Effective thermal conductivity for melting in PCM encapsulated in a sphere. Applied Energy 122: 280-287.

16. http://www.mgm.gov.tr/veridegerlendirme/il-ve-ilceler-istatistik.aspx?m=ANKARA#sfB

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[Abstract:0096][Renewable Energy Systems and Applications] PREDICTION OF SOIL TEMPERATURES FOR UNDERGROUND HEAT EXCHANGER

APPLICATIONS IN IZMIR TURKEY

Deniz Yenera, Onder Ozgener b, Leyla Ozgenerc,*

a Graduate School of Natural and Applied Sciences, Solar Energy Science Branch, Ege University, Bornova, Izmir, Turkey

bSolar Energy Institute, Ege University, Bornova, Izmir, Turkey

c,*Department of Mechanical Engineering Faculty of Engineering, Celal Bayar University Muradiye, Manisa, Turkey Corresponding e mail:leyla.ozgener@gmail.com

SUMMARY This investigation deals with the prediction of soil temperature’s dependence with depth and time for underground heat exchanger applications, specially earth to air heat exchangers and ground source heat pump systems. Plenty of parameters can affect ground thermal behaviors, for instance short-term climatic effects , conductivity and moisture density, especially near the earth surface. The main drawback is that although these values are significantly important, they cannot be reached easily. Thus, we developed a mathematical model in regards to predict the soil temperature for Izmir. Measured data that was taken from the Izmir State Meteorological Station, and predicted soil temperatures were compared for 2014 data. Additionally, at depths of 5cm, 10cm, 20cm, 50cm and 100cm, the maximum average percentage errors were 14% , 12.7% , 12.6% , 11.3% , 5.6% respectively. Consequently, this paper gives an insight in the relationship between air and soil temperatures in terms of depth from 5cm to 300cm. INTRODUCTION Having correct soil temperature values provide an advantage for researchers. Despite it is importance, in today's world, accessibility is a real problem. Because of the fact that studies related to soil temperature estimations had increased throughout the world by researchers. A. O. Ogunlela determined soil temperature values as a function of time and depth [1,2] . O. Ozgener et al. focused the prediction of soil temperatures for geothermal heat exchanger applications. Several measuring data was taken from author's experimental earth to air heat exchanger system in Ege University - Izmir. According to these experimental study, authors validated the relation between soil temperatures at different depths and ambient air temperature [3] . M. Badache et al. investigated the periodic variation of solar radiation and ambient air temperature in order to estimate ground temperature profiles [4] . Even though soil temperature values are design parameters in ground or soil applications, still, data could be incomplete ,therefore studies may lose accuracy. Due to this reason, the objective of this paper is to access easily to necessary values. METHODS Detailed explanation and formulas of this theoretical model was indicated by the authors in earlier studies [3] . According to this investigation, z is the depth below the ground surface with a maximum value of 1000cm and T is the soil temperature. A sinusoidal temperature model was derived by Hillel (1982) [4] solving previous equations subject to initial and appropriate boundary conditions to yield the following: , (1) where T(z, t) is the soil temperature as a function of time t and depth z, Tm is the average ambient air temperature and t0 is the time lap needed for the surface soil temperature to reach Tm. In Eq. (2) Az is the amplitude of the temperature wave at depth z, time t and decays exponentially with depth. Is defined as:

. (2) The amplitude of soil temperatures symbolized by Ao or Az represent the temperature fluctuations. It is very important to calculate soil temperature values at different depths in terms of applications and studies. Ao is the difference between maximum air temperature and average air temperature in the investigated year. The amplitude of monthly temperatures was predicted as Ao =10. In addition, γ is the damping depth and was determined as 300cm for Izmir. This mathematical model was applied to depths from 5cm to 300cm. Another necessary design parameter is the time parameter to. This value takes an active role in the developed model for more accurate results. to of the annual cycle represents point zero. to is the time value of junction's average mean temperature with temperature distribution. A one-year period (P) had been taken in hours as

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8760h. t is the daily hourly period. On the other hand, some values are assumed constant in this study, for example the thermal diffusivity. Under these conditions, the average air temperature was measured as 18.8°C and the soil temperature fluctuation can be state based on Eq. (1) with the mentioned parameters for Izmir.

, 18.88 10exp 0.0189 0.0189 (3) RESULTS In this study, the main goal is to compare the measured and calculated soil temperatures at different depths. Soil temperature estimations as a function of time and depth were obtained from the 3rd equation. As it can be seen from Fig. 1. soil and air temperatures indicate the same tendency.

Figure 1. Soil and air temperatures according to months

These estimations were acquired for Izmir at depths of 5cm, 10cm, 20cm, 50cm, 100cm and 300 cm. The highest error rates are observed near the surface. This is because the soil temperature at shallow depths is affected by short-term conditions such as rain and snow. Therefore, as shown in Fig. 2. percentage errors reduce when depth increases.

Figure 2. Percentage errors according to depths

Figure 3 gives a comparison of the measured and calculated soil temperatures at 5-100cm

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Figure 3. Comparison of the measured and calculated soil temperatures at 5-100cm

On the contrary, the variety of soil temperatures at different depths as 300 cm, the soil temperature is calculated equally to the mean air temperature, demonstrating the validity of the improved theoretical model. CONCLUSİONS The improved theoretical model involves the prediction of soil temperature from monthly average air temperatures. The estimation of soil temperatures related to air temperatures provide to reach the necessary values for engineering applications, especially those in relation with low geothermal enthalpy resources. We reached following result,

According to depths of 5cm, 10cm, 20cm, 50cm and 100cm , maximum average percentage errors are approximately 14% , 12.7%, 12.6%, 11.3%, 5.6% respectively for Izmir.

ACKNOWLEDGEMENTS The authors are grateful for the meteorological data provided by the Izmir State Meteorological Station. REFERENCES [1]Izmir State Meteorological Station, 1960–2014, Records For Monthly and Annual Mean Air and Soil Temperatures, Izmir, Turkey (in Turkish). [2] A.O. Ogunlela, Modeling soil temperature variations, J. Agric. Res. Dev. 2 (2003). Faculty of Agriculture, University of Ilorin. [3] O. Ozgener, L. Ozgener , J. W. Tester, '' A practical approach to predict soil temperature variations for geothermal (ground) heat exchangers applications'' , International Journal of Heat and Mass Transfer, Volume 62, July 2013, Pages 473-480 [4]M. Badache , P. E. Nejad, M. Ouzzane, Z. Aidoun et al. '' A new modeling approach for improved ground temperature profile determination'' , Renewable Energy, Volume 85, January 2016, Pages 436–444 [5] D. Hillel, Fundamentals of Soil Physics, Academic Press Inc., San Diego, CA, 1980.

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[Abstract:0097][Renewable Energy Systems and Applications] EFFECT OF THE HFC AND HC REFRIGERANTS AS SECONDARY WORKING FLUID ON

PERFORMANCE OF BINARY GEOTHERMAL POWER PLANT

Anıl Başaran 1, Leyla Özgener2 1 Graduate School of Natural and Applied Sciences, Department of Mechanical Engineering Branch, Manisa Celal Bayar

University, Muradiye, TR 45140, Manisa, Turkey 2 Department of Mechanical Engineering, Faculty of Engineering, Manisa Celal Bayar University, Muradiye, TR 45140,

Manisa, Turkey Corresponding email: leyla.ozgener@cbu.edu.tr

SUMMARY Exergy destruction rates of both overall plant and various components of the plant were determined for screening refrigerants. It has been determined that energy and exergy efficiencies of the geothermal binary power plant decrease with increasing geothermal source temperature. It has been obviously seen that lower ΔTpp values result in high energy and exergy efficiencies. The low ΔTpp value is due to the high boiling point at evaporator pressure. The heat exchanger (evaporator) is of great importance in terms of system performance. As the exergy destruction which has been caused by the refrigerants in the HE increase, energy and exergy efficiencies decrease. INTRODUCTION In binary geothermal power plants, thermal energy obtained from geothermal sources is transferred to second working fluid. Therefore, selection of second working fluid plays a key role on the cycle performance. In this study, a sample geothermal binary cycle was modelled, and HC (Hidrocarbon) and HFC (Hidrofluorocarbon) refrigerants that have the least environmental impacts were selected as working fluid. HC and HFC refrigerants have better environmental characteristics compared with CFCs (chlorofluorocarbons) and HCFCs (hydrochlorofluorocarbons). Among the selected refrigerants, R 227ea, R 236ea R 245ca, R 245fa are HFC refrigerants; R 600, R 600a, R601 and R 601a are HC refrigerants. The effect of geothermal source temperature on energy and exergy efficiencies of the geothermal binary power plant was explored by the way of comparison with previous study [1]. Consumption of fossil fuels leads to some environmental problems such as pollution, global warming, and ozone depletion. Renewable energy sources have currently gained great importance because they avoid environmental problems and refresh themselves in the natural process. New energy conversion technologies are required to generate power from low-grade energy resources without causing environmental pollution. Solar heat, waste heat, and geothermal energy are typical examples for low-grade heat sources with their available temperatures ranging between 60 and 200 oC [2]. Geothermal energy utilization is generally divided into two categories, i.e., electric energy production and direct uses for space heating and cooling, industrial processes [3–6]. Most of the world’s geothermal power plants were built in the 1970s and 1980s following the 1973 oil crisis [7]. Chammaro et al. claim that the capacity factor of geothermal power plants is highest of all renewable energies and also higher than capacity factor of other type power plants [8]. Bertani prepared a report in the matter of geothermal power generation in the world 2005–2010 update. According to this report, the 2010 worldwide installed geothermal power plant capacity is 10.9 GW, and the average capacity per unit for all the 536 units in operation is 20MW [9]. Binary power plantsare appropriate for electricity production from low and medium temperature geothermal sources. The plants working with principles of Organic Rankine Cycles (ORC) and proprietary systems known as Kalina cycles are common binary power plants [10]. Yıldırım and Ozgener examined Aydın-Salavatlı geothermal field’s features and working principle of the 2 power plants (DORA I and DORA II). They calculated energy and exergy efficiencies of the power plants with emphasis on the effects of thermal fluid and developed some correlations [11]. Makhall et al. made exergy-topological analysis and optimization of a binary power plant ultilizing medium-grade geothermal energy. This study shows that medium-grade geothermal heat-to-power conversion in a well-designed secondary regenerative Rankine cycle can achieve a high degree of thermodynamic perfection and exergy efficiency, without the use of high-cost advance powerfluids [12]. In geothermal binary power plants, thermal energy obtained from geothermal fluid is transferred to second working fluid which is generally refrigerants, and this working fluid circulates through the Rankine cycle. Stability, safety, compatibility, cost, availability of refrigerant and environmental impacts are among the considerations while selection and utilization second working fluid for these systems. There are some restrictions in using of some refrigerants especially CFC and HFCF, due to their environmental effects. The Montreal Protocol seriously stress on the issue of reduction in the use and production of substances derived from the CFCs and HCFCs, which are regarded as ozone depleting substances [13]. Compared with CFCs and HCFCs, HFCs have better environmental characteristics. HFC refrigerants such as R 227ea, R 236ea contain no chlorine atoms, so their ODP (Ozone Depletion Potential) is zero. Hydrocarbons such as propane, n-butane (R 600), isobutane (R 600a) are generally considered as environmentally benign and consist only hydrogen and carbon atoms. They have negligible GWP and zero ODP [1]. Mohanraj et al. presented a review on environment friendly alternatives to halogenated refrigerants [14]. Calm made e review study on the next generation of refrigerants which includes historical review, considerations and outlook [15].

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Energy and exergy efficiencies are considered by many to be useful for the assessment of energy conversion and other systems and for efficiency improvement. Yari studied on exergetic analysis of various types of geothermal power plants and investigated the exergy destruction of each component of binary geothermal power plants [16]. Heberle and Brüggemann made an investigation in the matter of exergy based fluid selection for a geothermal organic Rankine cycle for combined heat and power generation. According to conclusion of this investigation, exergy analysis is appropriate to evaluate various concept of energy conversion in geothermal power plants [17]. In this study,the effects of geothermal source temperature and using different HC and HFC refrigerants as working fluid on performance of geothermal binary power plants were investigated. Mass, energy, and exergy equalities were applied. First and second law efficiencies of thermodynamic were calculated for cycle equipments. Refrigerants can be working fluids for binary power cycle were determined and analyzed. METHODS A binary geothermal power plant shown in Fig 1 was modeled and studied in this study. The following working conditions were assumed in this plant with the aim of selecting and screening of optimum working fluids based on characteristic operation conditions of binary power cycle where secondary working fluid is used. These working conditions were specified considering reality plant operation conditions and they show parallelism with the reality binary power plants examined by e.g., [10,11] with respect to being same thermodynamic phase at the same status in the cycle.

Figure 1. Shematic of analyzed geothermal binary power cycle.

There is no environmental discharge in the considered plant. Geothermal fluid is completely reinjected to reservoir. The power plant operates on a liquid dominated and low temperature resource and geothermal fluid remains as a liquid throughout the plant. Geothermal water enters the plant at T5=408 K with a mass flow rate m5=75 kgs-1 and pressure of geothermal water is P5=0.5 MPa at the entrance of heat exchanger (evaporator). Geothermal water leaves the heat exchanger at T6=334 K. The working fluid circulates in closed cycle that is based on Rankine cycle. The working fluid enter the pump as saturated liquid at T1=283 K and is pumped to P2=1 MPa which is the heat exchanger pressure. The isentropic efficiency of the pump is assumed to be 75%. The working fluid enters the heat exchanger at P2=1 MPa and leaves after it is evaporated and superheated to T3=401 K. The superheated working fluid passes through the turbine. The isentropic efficiency of turbine is assumed to be 85%. After the working fluid expands at the turbine, it passes through the water-cooled condenser. It leaves the water-cooled condenser as saturated liquid at T1=283 K. The cooling water enters the water-cooled condenser at T7=280 K and leaves at T8=288 K. Selection of the working fluid affects the performance of the binary geothermal power plant to a great extent. For the purpose of this study, 8 different refrigerants were selected and analyzed as working fluid in the binary geothermal power plant given in Fig 1. Investigated refrigerants are R 227ea, R 236ea, R 600 (butane), R 600a (isobutane), R245ca, R 245fa, R601 (pentane) and R 601a (isopantane). Among these refrigerants, R 227ea, R 236ea, R 600 and R 600a were selected to explore effect of geothermal source temperature on energy and exergy efficiencies of binary geothermal power plant via comparison of study made by Basaran and Ozgener [1]. Refrigerants R245ca, R 245fa, R601 and R 601a were selected to increase energy and exergy efficiencies of the plant for 408 K geothermal source temperature. These refrigerants are considered that they are able to increase energy and exergy efficiencies of the plant as against R 227ea, R 236ea, R 600 and R 600a. Also, environmental properties of refrigerants were taken into account during the refrigerant selecting process. Among the investigated refrigerants, R 227ea, R 236ea R245ca, R 245fa are HFC refrigerants; R 600, R 600a, R601 and R 601a are HC refrigerants. Their environmental, thermophysical and safety properties are given in Table 1. These refrigerants chosen because they enable operating conditions that assumed for considered binary geothermal power plant. The freezing points of refrigerants are below the ambient temperature. Otherwise, the refrigerants may solidify during the system operation. In this study, restricted dead state temperature and pressure were taken to be T0=278 K and P0=1 atm (101.25 kPa). REFPROP 8.0 [18] software was used in this paper to calculate the properties of working fluids, geothermal water and cooling water.

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Table 1. Physical, safety and environmental datas of investigated refrigerants [20,21]

Number Name Safety ODP GWP1

00

Atmospehric

Lifetime

Molecular

Weight ()

Normal Boiling Point (K)

Critical Tempera

ture (K)

Critical Pressure

(kPa)

R 227ea

1,1,1,2,3,3,3-Heptafluoropropane

- 0 3220 34.2 170.03 256.81 374.9 2925

R 236ea

1,1,1,2,3,3-Hexafluoropropane

A1 0 1200 10 152.04 279.34 412.44 3502

R 245ca

1,1,2,2,3-Pentafluoropropan

- 0 693 6.2 134.05 298.28 447.57 3930

R245fa 1,1,1,3,3-pentafluoropropane

B1 0 1030 7.6 134.05 288.29 427.20 3640

R 600 Butane A3 0 ~20 0.018 58.122 272.66 425.13 3796

R 600a Isobutane A3 0 ~20 0.019 58.122 261.4 407.81 3629

R601 Pentane A3 0 ~20 0.01 72.15 309.21 469.70 3370

R 601a Isopentane A3 0 ~20 0.01 72.15 300.98 460.35 3378

The following assumptions are considered during this study:

The kinetic and potential energy changes are negligible. The friction losses are negligible. The system is assumed as a steady-state, steady-flow process. Thermophysical properties of geothermal water are considered same as thermophysical properties of water.

Chemical substances in geothermal fluid and non-condensable gases are neglected. The pressure drops throughout the heat exchangers (evaporator and condenser) and pipelines are neglected The kinetic, potential, and chemical exergies are neglected.

Mass, energy and exergy balance equations For a general steady-flow process, the three balance equations, namely mass, energy, and exergy balance equations, are employed to find the heat input, the rate of exergy decrease, the rate of irreversibility, and the energy and exergy efficiencies. In general, the mass balance equation can be expressed in the rate form as

in outm m (1) where m is the mass flow rate, the subscripts in and out stand for inlet and outlet. The general energy balance with negligible kinetic and potential energy changes can be expressed by,

in in out outQ W m h m h (2)

where Q and W are the net heat input and work output, and h is the specific enthalpy. The energy efficiency of the system can be defined as the ratio of total energy output to total energy input [19].

outputsystem

input

EE

(3)

where in most cases “output” refers to “useful” one. In this context, the specific flow exergy can be defined as,

0 0 0( ) ( )h h T s s (4) Multiplying specific flow exergy by the mass flow rate of the fluids gives the rate of total exergy:

.Ex m (5)

0 0 0[( ) ( )]Ex m h h T s s (6)

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where subscript 0 stands for the restricted dead state. 0T is the dead state temperature, 0h and 0s are enthalpy and entropy at

the restricted dead state of 0P and 0T . The general exergy rate balance may be expressed as follows:

, ,dest heat work mass in mass outEx Ex Ex Ex Ex (7) Exergy destruction (or irreversibility) of the system can be defined as follows:

dest in outEx Ex Ex (8)

Exergy destruction is also defined as following equation. In this equation, firstly, the entropy generation genS is calculated and used in the equation,

0.dest genI Ex T S (9) Generally, the exergy efficiency (also called second law and exergetic efficiency) maybe expressed as the ratio of total exergy output to total energy input.

output

input

ExEx

(10)

RESULTS In this study, the effect of different refrigerants on performance of binary geothermal power plants was investigated. Mass, energy and exergy balance equations were solved under same input conditions to validate the computations. The effect of geothermal source temperature on energy and exergy efficiencies of geothermal power plant was explored via comparison of study made by Basaran and Ozgener [1]. Alternation of energy and exergy efficiencies for screening refrigerants depending on geothermal source temperature is shown in Fig. 2. Results show that the energy and exergy efficiencies of the geothermal binary power plant decrease with increasing geothermal source temperature. According to Fig. 2 there is an acceptable decrease in energy efficiency of the overall plant whereas there is an intolerable falling off in overall plant exergy efficiency. For screening refrigerants which are R 227ea, R236ea, R 600 and R 600a, the decrease of energy efficiencies are less than 1%. The minimum decrease of exergy efficiency depending on increment of geothermal source temperature is 9.02% that is showed by R 227ea. On the other hand, the maximum decrease of exergy efficiency of the plant is 11.55% that is showed by R 236ea. The basic reason of decrease of exergy efficiencies is that the temperature differences of both geothermal fluid and working fluid between inlet and outlet of evaporator increase.

Figure 2. Alternation of energy and exergy efficiencies for screening refrigerants depending on geothermal source

temperature. The place in the heat exchanger where the brine and the working fluid experience the minimum temperature differences is called the pinch-point, and the value of that differences is designated the pinch-point difference, ΔTpp [20]. The pinch-point temperature is simply the difference between brine pinch-point temperature and the vaporization temperature of working fluid [21]. The low ΔTpp value is indicator of better thermal match between geothermal and working fluids during heat transfer in evaporator. It is known that the high boiling point at evaporator pressure results in the low ΔTpp value. Therefore, R 245ca, R 245fa, R 601 (pentane), R 601a (isopentane) refrigerants, which have higher boiling point at 1 atm evaporator pressure than R 227ea, R236ea, R 600 and R 600a, were selected as working fluid to increase of the energy and exergy efficiencies of the plant for the 408 K geothermal source temperature. R 227ea, which has the lowest boiling point among the selected refrigerants at evaporator pressure 1 atm (326.58 K), shows the maximum ΔTpp value. Next-higher ΔTpp values are obtained from R 236ea, R 600a, R 600, one by one. On the other hand, R 601, which has the highest boiling point among the selected refrigerants at evaporator pressure 1 atm (398.07 K), exhibits the minimum ΔTpp value.

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Fig. 3a demonstrates that the exergy destruction percentages for all the refrigerants. The exergy destruction percentages were calculated via getting percentage of total exergy destruction of the system that calculated for all investigated refrigerants. According to Fig 3a, R 227ea has maximum exergy destruction percentage with 21% and it is followed by R 600a with 16%, R 236ea with 14%, R 600 with 13%. On the other hand, R 601 has minimum exergy destruction percentage with 7% and it is followed by R 601a with 8%, R 245ca with 9%. It can be also seen in Fig. 3b that R 601 and R 601a give rise to low exergy losses, relatively. It is expected that R 601 has the best performance, due to its minimum exergy destruction. Fig. 4 shows the exergy destructions for the various components of the plant. In compliance with this figure, higher exergy destruction occurs in the heat exchanger for investigated refrigerant except R 601 and R 601a. Due to lower exergy destruction in heat exchanger (evaporator), R 601 and R 601a lead to minimum exergy destruction and maximum performance for overall plant among the investigated refrigerants. Therefore, the heat exchanger plays a key role on the performance of the plant.

(a) (b)

Figure 3. (a) Exergy destruction percentage of refrigerants for overall plant, (b) Exergy destruction rates of overall plants

Figure 4. Exergy destruction of investigated refrigerants at various components of plant.

First and second law efficiencies of the overall cycle were calculated as in Fig. 5. It follows from this figure that R 601 exhibits the best first and second law efficiencies. It has 18.53% first law efficiency and 74.57% second law efficiency. It is followed by R 601a, which has 17.46% first law efficiency and 70.29% second law efficiency, and by R 245ca, which has 16.48% first law efficiency and 66.33% second law efficiency. On the other hand, R 227ea shows the worst first law efficiency (8.5%) and second law efficiency (34.19%). R 245ca, R 245fa, R 601 and R 601a illustrate better performance than R 227ea, R 236ea, R 600 and R 600a. It is mentioned previously that low ΔTpp value causes better performance. Therefore, it is expected that R 245ca, R 245fa, R 601 and R 601a have the best performance, due to their lower ΔTpp values.

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Figure 5. Energy and exergy efficiencies of investigated refrigerants.

DISCUSSION The effects of geothermal source temperature and using different HC and HFC refrigerants as working fluid on performance of geothermal binary power plants were investigated. A binary geothermal power plant was modeled and working conditions of this plant were specified considering reality plant operation conditions. This result is consistent with the literature [10].The results show that:

1- In the considered geothermal binary power plant, energy (first law) efficiencies are obtained in the range of 8.5-18.53% and exergy (second law) efficiencies are gained in the range of 34.19-74.57%. The considered geothermal binary power plant which uses low-temperature geothermal resource as heat source exhibits higher exergy efficiency than energy efficiency. It has been determined that energy and exergy efficiencies of the geothermal binary power plant decrease with increasing geothermal source temperature. Decrease of energy efficiency is less than 1% while decrease of exergy efficiency decrease is between 9.02% and 11.55%.

2- It has been obviously seen that ΔTpp values influence the system performance. It has been determined that lower ΔTpp values result in high energy and exergy efficiencies. Refrigerants with low ΔTpp value has illustrated higher performance. The low ΔTpp value is due to the high boiling point at evaporator pressure. Therefore, refrigerants presented lower ΔTpp value that is to say refrigerants have high boiling point at evaporator pressure increase performance of geothermal binary power plants.

3- The heat exchanger (evaporator) is of great importance in terms of system performance. As the exergy destruction which has been caused by the refrigerants in the HE increase, energy and exergy efficiencies decrease. It is obvious that effort of decrease exergy destruction in evaporator increase the system performance. R 601a has caused the minimum exergy destruction in the evaporator and exhibited the highest energy and exergy efficiencies. In contrast, R 227ea has caused the maximum exergy destruction in the evaporator has shown the lowest energy and exergy efficiencies. Therefore, exergy destruction in the HE should be taken into consideration during the selection of the working fluid.

4- HFCs and HCs have better environmental characteristics, compared with CFCs and HCFCs, so they are good alternative to CFCs and HCFCs. Besides this, toxicity and flammability of refrigerants should be low as much as possible. While selecting the working fluid for the geothermal binary power cycle, refrigerants have optimum thermodynamics, thermophysical and environmental properties should be chosen, but it is certainly difficult to find ideal working fluid which shows high energy and exergy efficiencies, reasonable operating conditions, low ODP and GWP values and is non-toxic and non-flammable.

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REFERENCES [1] Basaran A, Ozgener L. Investigation of the effect of different refrigerants on performances of binary geothermal

power plants. Energy Convers Manag 2013;76:483–98. doi:10.1016/j.enconman.2013.07.058. [2] Yamamoto T, Furuhata T, Arai N, Mori K. Design and testing of the organic rankine cycle. Energy 2001;26:239–

51. doi:10.1016/S0360-5442(00)00063-3. [3] Ozgener L, Hepbasli A, Dincer I. Performance investigation of two geothermal district heating systems for building

applications: Energy analysis. Energy Build 2006;38:286–92. doi:10.1016/j.enbuild.2005.06.021. [4] Ozgener L, Ozgener O. Monitoring of energy exergy efficiencies and exergoeconomic parameters of geothermal

district heating systems (GDHSs). Appl Energy 2009;86:1704–11. doi:10.1016/j.apenergy.2008.11.017. [5] Ozgener O, Ozgener L. Exergoeconomic analysis of an underground air tunnel system for greenhouse cooling

system. Int J Refrig 2010;33:995–1005. doi:10.1016/j.ijrefrig.2010.02.008. [6] Ozgener O, Ozgener L, Goswami DY. Experimental prediction of total thermal resistance of a closed loop EAHE

for greenhouse cooling system. Int Commun Heat Mass Transf 2011;38:711–6. doi:10.1016/j.icheatmasstransfer.2011.03.009.

[7] Ozgener L, Hepbasli A, Dincer I. Energy and exergy analysis of geothermal district heating systems: An application. Build Environ 2005;40:1309–22. doi:10.1016/j.buildenv.2004.11.001.

[8] Chamorro CR, Mondéjar ME, Ramos R, Segovia JJ, Martín MC, Villamañán M a. World geothermal power production status: Energy, environmental and economic study of high enthalpy technologies. Energy 2012;42:10–8. doi:10.1016/j.energy.2011.06.005.

[9] Bertani R. Geothermal power generation in the world 2005-2010 update report. Geothermics 2012;41:1–29. doi:10.1016/j.geothermics.2011.10.001.

[10] DiPippo R. Second Law assessment of binary plants generating power from low-temperature geothermal fluids. Geothermics 2004;33:565–86. doi:10.1016/j.geothermics.2003.10.003.

[11] Yildirim D, Ozgener L. Thermodynamics and exergoeconomic analysis of geothermal power plants. Renew Sustain Energy Rev 2012;16:6438–54. doi:10.1016/j.rser.2012.07.024.

[12] Makhanlall D, Zhang F, Xu R, Jiang P. Exergy-topological analysis and optimization of a binary power plant utilizing medium-grade geothermal energy. Appl Therm Eng 2014. doi:10.1016/j.applthermaleng.2014.09.017.

[13] Aljundi IH. Effect of dry hydrocarbons and critical point temperature on the efficiencies of organic Rankine cycle. Renew Energy 2011;36:1196–202. doi:10.1016/j.renene.2010.09.022.

[14] Mohanraj M, Jayaraj S, Muraleedharan C. Environment friendly alternatives to halogenated refrigerants-A review. Int J Greenh Gas Control 2009;3:108–19. doi:10.1016/j.ijggc.2008.07.003.

[15] Calm JM. The next generation of refrigerants - Historical review, considerations, and outlook. Int J Refrig 2008;31:1123–33. doi:10.1016/j.ijrefrig.2008.01.013.

[16] Yari M. Exergetic analysis of various types of geothermal power plants. Renew Energy 2010;35:112–21. doi:10.1016/j.renene.2009.07.023.

[17] Heberle F, Brüggemann D. Exergy based fluid selection for a geothermal Organic Rankine Cycle for combined heat and power generation. Appl Therm Eng 2010;30:1326–32. doi:10.1016/j.applthermaleng.2010.02.012.

[18] NIST standart reference database 23. REFPROP 2007. [19] Özgener L. Exergoeconomic analysis of geothermal district heating systems (in Turkish). Ege University, 2005. [20] DiPippo R. Geothermal Power Plants: Principles, Applications, Case Studies and Environmental Impact. 2nd ed.

Oxford: UK: Elsevier Ltd; 2008. [21] Kanoglu M. Exergy analysis of a dual-level binary geothermal power plant. Geothermics 2002;31:709–24.

doi:10.1016/S0375-6505(02)00032-9.

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[Abstract:0098][Energy Efficient Buildings] ANALYZİNG AUTOMATİON SYSTEMS WHİCH İS USED FOR HEATİNG-COOLİNG THE

BULDİNGS, İN TERMS OF COMFORT AND ENERGY EFFİCİENCY

Semiha Öztuna1, Seyhan Özkan2 1Trakya Üniversitesi, Edirne, Turkey 2Bakırcı Otomativ, İstanbul, Turkey

Corresponding email: semihae@trakya.edu.tr

SUMMARY In this study, Differences in the structures of building due to the aim of their usage were investigated by considering the effects of heating -cooling, lighting and security systems, and comparisons were made in terms of cost, energy efficiency and comfort. 5 different building types were investigated in this study. The first building type is a mall (Avm) saving 22% in energy in heating-cooling with the usage only of time-dependent programming and population density control. With this saving, the amortization time is approximately 36 months. Secondly, a hotel is analyzed. In this case, a card automation is applied. The results show that a 25% energy saving is obtained. The amortization time in this case is found to be 45 months. Then, a surgery room is handled. Sustained hygienic air-conditioning and lighting are important at the surgery rooms. With the automation, a 24% energy saving is obtained. Amortization time is approximately 47 months. The other building type analyzed is congress center. This can be thought as an example of state buildings. With the control of water temperature and time programming, an approximately 25% energy consumption can be obtained. The amortization time in this case is 62 months. Lastly, automation systems are used on modern reinforced concrete buildings, in which the 70% of country population lives. The results show that a 23% energy saving The amortization time is 68 months. With only lighting automation, an annual 158 kw electricity saving is provided. In this case amortization time is 109 months. INTRODUCTION In our day issues regarding the energy consumption, energy saving and energy efficiency constitute some of the most important problems in all around the world. The building automation offers wide opportunities to the operator and a functional building to the user since this system can control the device and armatures of heating-cooling-ventilation, air conditioning, lighting, fire call and security Systems. The user can operate all systems in a building from one center according to preprepared and adaptable programs, can receive malfunction warning, can make capacity control, can archive incoming and outgoing commands and can realize a well designed automation system capacity in parallel with the facilities whose number increase according to the need in the building. The aim of building automation is to provide central supervision and operation, energy saving, comfort and security control. With its developing skills, building automation systems become the most important tools in energy management in buildings [1]. Providing to proceed according to the dynamic changes in a building, building automation systems led to the improvement of European norm EN15232 in 2007 and has become such a system that can offer live realization of the optimum solutions to the possible changes in behavior of a building (load, fullness, synchronization, malfunction of equipment etc.) [2]. Automation is one of the most important applications in timely determination of the malfunctions of equipment in heating, ventilation and air conditioning systems and in energy management and optimization [3]. Developments in building automation systems have provided various systems to operate integrated to each other and the information collected further provided buildings to operate in an unexpected high efficiency with instant or previous record assessments. In the energy usage in the buildings, smart buildings, building automation, automation systems in heating-cooling and energy efficiency stand out. In literature there are many researches on automation systems and energy efficiency. Yumurtacı and Keçebaş [4], Küçükbakırcı [5], Baysal [6] and Göksu [7] investigated smart buildings and automation applications in smart buildings in their works. Lin et al. [8], Maro [9], Ercan[2], Çimen [10], Atabaş et al. [11], Yakut et al. [12], Bayer (13] and Çilingir [14] studied building automation and automation applications in heating-cooling systems. In this work the buildings with 5 different usage purposes have been taken into consideration and automation (automatic control) systems used for energy observation and management in these buildings have been investigated in terms of energy efficiency, comfort and cost. The influence of the differences from the usage purpose of the buildings on heating-cooling, lighting and security systems have been evaluated and comparisons have been made. In this work, the important factors for the selection of the system have also been reviewed in order to make a comparison between different systems in 5 analyzed buildings and to obtain an opinion for an efficient building automation system.

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AUTOMATION SYSTEMS AND BUILDING AUTOMATION Automation is system activities in relation with its environment within the scope of the information supplied to it. The activities are realized according to the inputted or calculated information or according to a certain procedure with previously determined parameters. The usage areas of the automation systems can be divided into two groups: industrial automation systems and building automation systems. Building automation systems are systems that enable to monitor, command and control the systems in buildings by computers [15]. 1. Influence of Automation Systems on Energy Efficiency The most important problem in energy efficiency applications is the sustainability of the system. Energy efficiency systems that are efficient within a certain time frame but inefficient in general can be seen in practice. When reviewing the steps to be traced for the sustainable energy efficiency, the energy measurement should be made firstly, then application or process devices should be examined and the basic factors should be determined. After the determination of basic factors, control-management systems (automation) are installed. Finally the operation data of the installed system is monitored and the application is supervised. The steps to be followed are given in Figure 1.

Figure 1. Necessary steps for a sustainable energy efficiency [14] It is without question that automatic control and management systems are an

important factors in energy saving. The energy consumed in buildings is electrical or fuel energy. The highest ratio in consumption of electrical energy belongs to lighting, cooling and ventilation. The highest rates belong to the heating in usage of fuel energy. Savings are obtained up to rates reaching 30% thanks to HVAC control, lighting control and building management system control which are provided with automation systems in buildings which consume 20% of the energy in the world. It is possible to obtain a saving between 10% and 40% with the control of heating-cooling system, control of lighting system, control of security system [14] 2. Systems in buildings which can be applied automation Application areas of building automation systems are as the following: fire (function control, detector control, fire and life safety), security (doors, PIR (movement), integrity work), access (doors, buildings, fullness ratio), energy (consumption (electricity, water, gas, fuel), customer consumption, air/water, lighting and the usage of alternative energy sources), lighting (timeline, occupation), elevators (malfunction, maintenance), communication (voice, video, data), 24/8 monitoring (failure detection, settings of devices, conditioned observation, parking lot usage), HVAC (air conditioning units, boilers, pumps, fans, energy control, variable air flow rate, air quality). Air conditioning is the conditioning of the air in a pace to hold the temperature, moisture, cleanness and air movement at optimum level for human health and comfort or for the industrial process. Security systems are generally used in environmental safety. They are applied to the surrounding environment of the residence or business office. They mostly function with security detectors. Besides, camera systems, windows and doors monitoring detectors, barrier detectors functioning with invisible laser rays or pressure sensible fiber cable systems mounted at the garden or neighboring of the house are used. Security systems can easily function integrated with building automation devices. Security measures are taken against gas, fire and water flood. The system is activated in cases of fire, gas leakage and water flood that might happen in a house via smoke sensor, heat sensor and water sensor. Lighting systems are systems that determine and regulate the light need inside a dwelling by a central computer as in heating-cooling system. The used equipment is with automatic control. 3. Energy monitoring and management systems Energy monitoring and management systems are necessary for efficient automation solutions. Energy data of the structure or process where automation will be applied should be analyzed carefully. It should be acted according to the results of this analysis. Automation systems applied by collecting multiple systems under a single roof provide easiness both in application and in management. With the integration provided by the software, tens of independent systems such as HVAC, lighting, energy

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monitoring and control, engine control, entrance-exit control, security, energy distribution can be controlled when requested. Besides control, an efficient and productive operation of all these systems should also be provided. ANALYSIS OF SAMPLE BUILDINGS WITH AUTOMATION SYSTEMS In the study 5 different buildings have been taken into consideration and analyzed. At the end of these analyses investment and operating costs, energy saving rate, amortization time have been obtained separately for each building. Comparisons of operating costs, first investment costs and energy saving rates of the heating-cooling, lighting and security automation systems have also been made. 1. Mall automation Malls are structures that host a lot of people at different hours of a day with a large closed areas. As the density of people is high, air conditioning, lighting and especially security automations carry high importance. The mall considered in this study has a 10.000m2 closed area and has 4 floors and it also has a closed parking garage. There are approximately 20 shops at each floor and there is a foyer area on the top of the building for general purposes. The analysis results for the mall are given in Table 1. Table1. A comparison of the systems with automation and without automation for a shopping centre.

Without Automation With Automation Heating Cooling Total Heating Cooling Total Diff.

Investment Cost (TL) 263.200 365.400 % 28

Operating Cost in Summer (TL) 10.584 52.920

8.316 41.580 Operating Cost in

Winter (TL

15.540 108.780 12.200

85.400

Energy Saving

Electricity 50.400 KW/Month 39.600 KW/Month % 22 Natural Gas 21.000 Nm3/Month 16.500 Nm3/Month

Amortization Time

Operating Cost in Summer

Operating Cost in Winter

Operating Cost in Summer

Operating Cost in Winter

10.584 TL/Month

15.540 TL/Month

8.316 TL/ Month

12.200 TL/Month

36 Months

(Note: Summer is assumed as 5 months and winter is assumed as 7 months) For the other analyzed buildings, the results for automated or non-automated buildings have been obtained in the same manner. 2. Hotel automation Hotels are complex structures that should be designed carefully in terms of security comfort and are expected to offer different comfort opportunities with many locations used for different purposes. The hotel considered in this study has a 8.000m2 closed area and has 8 floors. 3. Operating room automation Operating room automation should be applied in a more careful, detailed and precise way than all other building automations. Since the hygiene of the environment is very important for the patient health all details should be taken into consideration. 4. Convention center automation The convention center considered in this study has a 5135 m2 closed area of conference hall for 1.000 people. The management and other buildings in the convention center have a 7850m2 closed area. 5. Villa automation The villa automation have been examined in a more detailed manner as opposed to the other automation considered in this study as they are the most common structures. Heat loss is 63kW in the sample villa and it is assumed to be heated with a boiler. The heat gain is 161 Btu/h in the sample villa and air cooling is assumed to be provided with a VRF air conditioning unit.

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RESULTS AND DISCUSSION In the study, heating-cooling, illumination and security systems with and without automation have been taken into consideration for 5 different types of building. In the mall, VAV (variable air volume) system automation has been assumed to be used due to the fact that it keeps most of the zone at stable temperature range and it meets clean and fresh air need. As the mall is used intensely for seven days of the week, it is seen that a high rate of efficiency can be obtained with automation. Even though the cost of air conditioning unit is 28% of the total investment cost it provides amortization in approximately 36 months by increasing the efficiency at 23% level in electricity and natural gas consumption. Lighting system is assumed to be consist of LED armatures due to their durability. Security system consists of doorstep detectors, fire sensors and CCTV cameras. The results of cost, energy saving rate and amortization time analyses are given in Table 2 for 5 different buildings considered in this study. In hotel, four pipe fan coil system automation is assumed to be used due to its heating-cooling capability in high precision in different zones. There is no continuous usage in the hotel room. This fact is also taken into consideration in the automation. The automation is activated when the person enters the room and it operates in basal mode in predetermined set values when the person exits the room. The hotel air conditioning system automation includes unit automations in the number of rooms, Although its investment cost is 35% of the total investment cost, it provides amortization in approximately 49 months by the increase in the efficiency at 25% level in electricity and natural gas consumption. The usage in lighting and security systems is also important factors and they are activated when a person enters the room. Lighting armatures are checked with dimmers and lighting is provided as much as it is necessary. CCTV cameras cannot be used in hotel rooms. Another building type considered in this study is operating room. In operating rooms, mostly single zoned air conditioning system automation is used as the necessary air flow rate and positive pressure can easily be provided with this system, the other rooms aren’t influenced in any problem in the air conditioning unit and half volumetric flow rate working principle is easier to apply. Even though operating room air conditioning system automation consists of 43% of total investment cost, it provides amortization in approximately 42 months by the increase in the efficiency at 25% level in electricity and natural gas consumption. In operating rooms priority is given to the hygiene and comfort instead of the efficiency. Hygiene is achieved by a three phases filtering, and quality, gas, and particle sensors in the room and linear air flow achieved through positive pressure. The most important thing in the automation of this type of rooms is the quick and efficient control of very sensitive comfort conditions and hygiene from a single station. Convention center is one of the examples of a public building . Heating is provided by a four pipe fan coil system in this structure. There is also a air conditioning unit. As the usage hours are known in the public buildings, the frequency inverters used in the automation systems and pumps provide an efficiency increase. Even though convention center air conditioning automation consists of 24% of total investment cost, it increases efficiency at 34% level in electricity and natural gas consumption and provides amortization within approximately 43 months. In lighting system compact fluorescent bulbs have been used due to its low consumption. Closed circuit camera systems and sound systems are considered in security automation. The last building type considered in this study is villa. the building is assumed to be heated by a central heating boiler and to be cooled by an air conditioning unit. An automation system which can be integrated to heating-cooling system is assumed to be used. The first investment cost of automation system consists of 37% of total investment cost the amortization time is found as approximately 82 months. With the automation, energy efficiency shows an increase of 22- 24%. To assess these two systems in terms of comfort we get results in parallel with energy efficiency. House automation system provides its user a comfortable environment due to its positive features such as remote control capabilities, security means, precision, timing options, ability to control multiple devices from a single location. Table 2. Results of cost, energy saving rate and amortization time analyses for buildings with automation systems [16].

Investment Cost ( % ) Energy Saving Rate (%) Amortization Time

(Month)

MALL 28 22 36

HOTEL 35 25 45

OPERATING ROOM 43 24 47

CONVENTION CENTER 24 25 62

VILLA 20 23 68

In this study, an energy efficiency, comfort and cost comparison for the automation systems used in 5 different buildings has also been made. The results are given in Table 3.

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Table3. Comparison of energy efficiency rate, comfort and cost of automation systems used for five different buildings [16]. BUILDINGS ENERGY EFFICIENCY COMFORT COST SHOPPING CENTRE

►Instant control of heating, cooling and ventilating load by measuring temperature of inside and environment ►Improvement of efficiency by concentration control (or time scenario) ►Energy saving rate for %22 is provided by applying automation system

►Easy and quick use with central control panel ►Temperature and fault alerts on computers

►In shopping centre air-conditioning system, it has got %28 portion of total first investment costs ►Amortization is provided in 36 months from savings

HOTEL ►Providing a control to stop air-conditioning and lighting in rooms out of use and activating system in hired rooms if there is someone in the room ►%25 energy saving rate is succeeded by using automation system

► Providing versatile use according to personal comfort demand ► Sensitive temperature and humidity control and easy reach to control systems

► According to calculations for hotel air-conditioning system, it has got %35 portion of total first investment costs ►Amortization is provided in 45 months from savings in energy consumptions

OPERATING THEATRE

►Aır-conditioning simulation with qualified computer aided sensors ►%24 energy saving rate is provided in fuels by using automation system

► Control of systems like heating ,cooling , humidity , particle, gas, pressure ,lighting from a single centre by doctor or bearer ► Good and fast harmonizing to comfort and hygiene conditions

► According to calculations for air-conditioning system in operating theatre, it has got %43 portion of total investment costs ►Amortization is provided in about 47 months thanks to fuel saving with automation system

CONFERENCE CENTER

►Preventing unnecessary use of air-conditioning and lighting system by running in particular times ► Control of fluid conditioning units by inverters ►Saving from human effort with automation scenario ►%25 saving is provided thanks to improvement from automation system

►Efficiency in use and maintenance of air-conditioning system ►High performance in lighting and sound system of conference centre

►According to calculations for air-conditioning system of conference centre, it has got %24 portion of total first investment costs ►It provides amortization approximately in 62 months

VILLA ►Efficiency improvement by preventing losses from doors, windows and lighting caused by user’s fault ► Providing the optimum start-stop control of temperature equipment ►Estimating all automation systems in house, energy saving rate is %23

► Easy integration to widely used systems like heating, cooling and lighting ►Easy and efficient use with time and presence scenario

► Automation system has got portion for %20 ► It subsidizes the costs for only heating and cooling in 68 months and for only lighting in 109 months. In case of the use of the all systems, it subsidizes the first investment amount in 82 months

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CONCLUSION Energy efficiency ratios of the automation systems applied to the buildings according to their intended use has been determined to be as the following; energy saving ratio 22%-25%, share of automation within first investment costs, 20%-43%, amortisement periods, 36-68 months. Significant points in the application of automation to buildings:

Selection of building automation system should be made during the planning of architectural and fittings process if possible.

Reviewing the efficiency of the automation systems used in buildings, the following are the most important parameters: daily or monthly building air conditioning periods, heating-cooling-ventilation loads and the dimensions of air conditioned locations (automation zone impact area).

In the selection of each building automation system the most important priority of ours should be the meeting of our expectations regarding the system. The build might necessitate different automation applications required by its location, type of use and its structure.

Cost and efficiency balance should be well considered in the selection of automation systems. There is a connection necessitating a diligent examination between the technological scope of the automation system to be applied and the efficiency of energy and comfort

Maintenance repair activities are very important expense items for installed house automation systems.

Building automation system practice is as much important as the selection of the system.

Another important feature of building automation devices is that they enable the control of breakdowns and problems of the air conditioning elements from a single location. Reporting the whole system is important for the learning about the energy features of the structure.

The things building automation systems can contribute on projects which are not well isolated, not located well and whose fittings have not been planned or applied diligently are limited.

REFERENCES 1. ABB (2009) Smart Home and Intelligent Bulding Controls,Technical Handbooks ,Heidelberg Germany , Page 1-23. 2. Ercan S. M. (2009) Bina Otomasyon Sistemi ile Devreye Alma İşlemleri ve Diğer Sistemlerle Bilgi Alışverişi, IX. Ulusal Tesisat Mühendisliği Kongresi, Türkiye, pp. 1121-1127. 3. Siemens Belgium (2010) Realization and content of standart EN 15232 , Bulding Technologies. 4. Yumurtacı M., Keçebaş A., (2009) Akıllı Ev Teknolojileri ve Otomasyon Sistemleri, 5. Uluslararası İleri Teknolojiler Sempozyumu (IATS’09) ,Karabük, Türkiye. 5. Küçükbakırcı C. (2006) Bir Mikro Denetleyici Tabanlı Akıllı Ev Sıcaklık ve Aydınlatma Otomasyonu Uygulaması, Yüksek Lisans Tezi, Gebze İleri teknoloji Üniversitesi Mühendislik ve Fen Bilimleri Enstitüsü, Gebze, Türkiye 6. Baysal R. (2008) Akıllı Binalarda Enerji Yönetimi ve Kontrolü, Yüksek Lisans Tezi, Süleyman Demirel Üniversitesi Fen Bilimleri Enstitüsü, Isparta, Türkiye 7. Göksu S. (2010) Akıllı Bina Uygulamalarında Maliyet Artışının Geri Kazanım Süreç Analizi, Yüksek Lisans Tezi, Sakarya Üniversitesi Fen Bilimleri Enstitüsü, Sakarya, Türkiye. 8. Lin C. , C. Fiderspiel C. , Auslander D., (2007) Multi-Sensor Single-Actuator Control of HVAC Systems, University of California, Berkeley, National Science Foundation, Grant No: 0088648. 9. Maro O. (1995) Bina otomasyonu, II Ulusal Tesisat Mühendisleri Kongresi Bildirisi, İzmir Mazzacane S. ,Giaconiac. Costanzo S. ,Cusumano A. , Lupo G. , Ventilation Systems And Thermal Conditions in Operating Rooms ,Italy. 10. Çimen L. (2008) Bina Otomasyon Sistemleri, Türkiye Tesisat Mühendisleri Odası Dergisi, No: 33. 11. Atabaş İ., Arslan M., Uzun İ., (2007) Isıtma Sistemlerinin Otomasyonu ve İnternet Üzerinden Kontrolü, Akademik Bilişim, Kütahya, Türkiye. 12. Yakut K. ,Koru M., Şencan A.,(2008) Hvac Sistemlerinde Kontrol Sistemleri ve Enerji Tasarrufu, Tesisat Mühendisliği Dergisi, İstanbul, Türkiye. 13. Bayer S. (1993) Isıtma, Havalandırma Klima Sistemlerinde Genel Otomatik Kontrol Uygulamaları, Teskon Bildirisi, Ankara, Türkiye. 14. Çilingir Ü. (2011) Binalarda Enerji Verimliliği ve Otomasyon Çözümleri, Schneider Electric, İstanbul, Türkiye. 15. ASHRAE Fundamentals of HVAC Control Systems, (2007) ASHRAE Hong Kong Cheapter Technical Workshop, Hong Kong,China, pp :1-53. 16. Özkan S. (2011) Bina Isıtma- Soğutmasında Otomasyon Sistemlerinin Konfor ve Enerji Verimliliği Yönünden İncelenmesi, Yüksek Lisans Tezi, Trakya Üniversitesi, Fen Bilimleri Enstitüsü, Edirne-Türkiye

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[Abstract:0099][Energy Efficient Buildings] ESTIMATING OF THE EFFECT OF COOL ROOFS ON ENERGY SAVINGS IN VARIOUS

CLIMATE ZONES

Murat Bulut1,2, Gul Eroglu Bulut3 and Nedim Sozbir 1,2 1Sakarya University, Faculty of Engineering, Mechanical Engineering, Sakarya, 54187, Turkey

2TURKSAT A.S, Gazi University Golbasi Yerleskesi, Gazi Teknopark, Golbasi/Ankara, 06839, Turkey 3Gazi University, Graduate School of Natural and Applied Science, Environmental Sciences, Ankara, 06500, Turkey

Corresponding email: bulut44@gmail.com

SUMMARY DURING THE SUMMER SEASON DUE TO CLIMATE CHANGE, COOLING LOADS OF BUILDINGS INCREASE. Therefore, energy demands increase. The countries more focused on energy savings, the cool roof system is one of the technologies to reduce the energy demands and global warming. The use of cool roofs affects cooling and heating energy use in buildings. Cool roof is gaining more attention because roof of the building is most exposed to sunlight. In this study, energy savings based on heating and cooling loads are studied by using different cool roofs at 5 different climate zones with flat roof. The results of the impact of cool roofs on the heating and cooling loads of the buildings are explained. The results show that using cool roof system materials instead of non-cool roofing material in buildings decrease cooling-electricity use, cooling-power demand and save the energy throughout the year 10% and 20% and reduction in cooling energy costs in buildings between 10% and 40%. The effectiveness of cool roofs as the urban areas temperature control on energy savings and reduction of global warming was demonstrated in this study.

INTRODUCTION Over the recent decades the growth of the population in urban areas has been one of the most significant ever [1]. With increasing industrialization, the people living in urban areas increase compare to the people living in rural areas [1]. Today half of the world’s population lives in the cities while 100 years ago only 14% and again in the 1950’s less than 30% of the world’s population was urbanized [1]. According to estimation, by 2040, 70 % of the world’s population will live in cities [2]. Increasing the population in urban areas creates the urban heat island-uhi. The causes of the uhi effect include increased building density, the use of materials with inappropriate optical and thermal properties and lack of green spaces, increased anthropogenic heat and increased air pollution [3]. The urban heat island means that because of the buildings the temperature in urban areas is more than the temperature in rural areas. An important part of the phenomenon ‘’heat islands’’ are the roofs and road surfaces that are heated through solar radiation and reach very high temperatures. The main reason why the temperature of urban areas is more than rural areas is less green areas and more buildings. Another reason why the urban heat island is high in urban areas because buildings are close each other. The exhaust from the cars and exhaust from hvac systems create heat [5]. These causes the urban heat island too. The uhi was first monitored in london back to 19th century [4], but many studies were performed during the past decades [6-8]. Many measurement campaigns showed daytime average temperature difference between 2 and 5 oc in cities respect to close-by rural area [9]. The high temperature in urban areas because of uhi affect the quality of life, public healthy, and low income level people adversely. A city level, the temperature of air in the city centers is by several degrees higher compared to surrounding suburban and urban areas which are known as the urban heat island. The sketch of an urban heat profile is shown in figure 1

Figure 1. Sketch of an Urban Heat Island Profile (Source: Heat Island Group http://eetd.lbl.gov/HeatIsland) [10]

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METHODS A cool roof is a roofing system that delivers higher solar reflectance and higher thermal emittance. During the day the roof of the buildings take heat load from the sun. Cool roof system kept this heat loads as low as possible. The material used the roof of buildings and the material used for asphalt has high absorption values. These materials make increase temperature in urban areas. Energy balance on a roof and properties of cool roof are explained as follows. When a roof is exposed the sun, the following Figure 2 shows the physical process take place and then the temperature of roof is determined.

Figure 2. Energy balance of a roof (Source: http://garoofingandrepair.com/elastomeric-roofing/) [11]

Solar radiation reaches on the roof and part of it is reflected and part absorbed by the roof contributing to its heating [12]. The roof emits the radiation in the far infrared part of the spectrum as radiation exchange occurs two surfaces which are roof and sky when one is warmer than the other and they ‘’view’’ each other [12]. The roof exchanges energy by convection with the air above the roof [12]. The heat is conducted through the layers (insulation etc.) within the roof from the warmer side to the cooler side [12]. The conduction induced heat flow through the roof- qin, and hence the energy needed for heating or cooling, depends on the thermal resistance of the roof material and the difference in temperature between the outer and inner surface [12].

RTTq sc

in)(

(1)

In this equation, R is overall thermal resistance of the roof material (m2K/W). Ts is temperature of the outer surface of the roof (K) and Tc is temperature of the inner surface of the roof (K). The thermal balance of a horizontal surface (roof) exposed to the sun is the following:

(2) SR: solar reflectance or albedo of the surface I: insolation (W/m2) Ɛ: emittance of the surface σ: the Stefan-boltzmann constant (5.6685 x 10-8 W/m2K4) h: convection coefficient (W/m2K4) Tsky: sky temperature (K) Ta: air temperature (K) If the roof is insulated underneath, the main factor affecting the thermal performance of the surface are the solar reflectance and the infrared emittance [12]. During the day the dominant factor is solar reflectance and emissivity has a lower effect on the surface [12]. However, during night-time the surface temperature and the infrared emittance are strongly correlated which means that emissivity becomes the most important factor affecting the thermal performance [13, 14]. Properties of cool roof system [12] The following two properties is a roofing system characterized by 1.High solar reflectance (SR): The high solar reflectance is a measure of the ability of a surface material to reflect solar radiation. The solar reflectance (SR) designates the total reflectance of a surface, considering the hemispherical reflectance

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of radiation, integrated over the solar spectrum, including specular and diffuse reflectance. It is measured on a scale of 0 to 1 (or 0-100%). For example, a surface that reflects 60% of sunlight has a solar reflectance of 0.60. Most dark roof materials reflect 5 to 20 % of incoming sunlight, while light-colored materials typically reflect 55 to 90 %. Solar reflectance has the biggest effect on keeping the roof cool in the sun. 2.High infrared emittance (Ɛ): The high emittance is a measure of the ability of a surface to release, absorbed heat. It specifies how well a surface radiates energy away from itself as compared with a black body operating at the same temperature. Infrared emittance is measured on a scale from 0 to 1 (or 0-100%) where a value 1 indicates a perfectly efficient emitter. These two properties result in affecting the temperature of a surface [13, 15]. If a surface with high solar reflectance and infrared emittance is exposed to solar radiation it will have a lower surface temperature compared to a similar surface with lower SR and Ɛ values. If the cool surface is on the building envelope, this would result in decreasing the heat penetrating into the building and for a surface in the urban environment this would contribute to decrease the temperature of the ambient air as the heat convection intensity from a cooler surface is lower. The phenomenon is shown in Figure 3 and Figure 4.

Figure 3. Combined effects of solar reflectance and emittance on roof temperature [16]

Figure 4. Black roof and cool roof [17]

Comparing between standard roof and cool roof is shown in Figure 5. Cool roof looks like standard roof but have higher solar reflectance (R).

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Figure 5. Standard roof and cool roof (Image source: American Rooftile Coatings and Lawrance Berkeley National

Laboratory) Cool roof system analysis is based on calculating the ‘’solar reflectance index’’ or SRI value. SRI is another metric for comparing the ‘’coolness’’ of roof surfaces. The SRI is a value that incorporates both the solar reflectance and emittance in a single value to represent a material’s temperature in the sun. This index tells us how hot a surface would get relative to a standard black and standard white surface. The SRI would be a value between 0 (as hot as a black surface) and 100 (as cool as a white surface) found from:

)()(

100whiteblack

surfaceblack

TTTT

SRI

(3)

RESULTS

Comparing the Temperatures Between Cool Roof and Standard Roof Solar reflectance and thermal emittance have noticeable effects on surface temperature. The calculation was done by using the excel which was created by the Lawrence Berkeley National Laboratory Heat Island Group. The temperatures of standard and cool roof are shown in Table 1. ASTM E1980-11 standard used to calculate SRI values. Table 1 illustrates these differences using roof materials which are in Figure 5. It is seen from Table 1 that the temperature of cool roof is lower than the temperature of standard roof.

Table 1. Comparing the temperatures between cool roof and standard roof.

Energy Savings with Cool Roof In USA, the Department of Energy has developed an energy saving calculator (DOE Cool Roof Energy Calculator) that can help the people determine the potential savings associated with installing an energy saving roof or cool roof [18]. Five different cities within 5 different climatic conditions are selected. Table 2 shows energy savings cool roof at selected cities in five different climate zones by using DOE Cool Roof Energy Calculator. During the calculation, five selected cities at

Black Color Roof Standard Roof Cool Roof Terracotta Roof Standard Roof Cool RoofSolar Reflectance 0,04 0,41 Solar Reflectance 0,33 0,48Thermal Reflectance 0,8 0,8 Thermal Reflectance 0,8 0,8SRI -7 42 SRI 31 52Roof Temperature (oC) 85,2 66,4 Roof Temperature (oC) 70,6 62,7Blue Color Roof Standard Roof Cool Roof Green Color Roof Standard Roof Cool RoofSolar Reflectance 0,18 0,44 Solar Reflectance 0,17 0,46Thermal Reflectance 0,8 0,8 Thermal Reflectance 0,8 0,8SRI 11 47 SRI 10 49Roof Temperature (oC) 78,2 64,8 Roof Temperature (oC) 78,7 63,8Gray Color Roof Standard Roof Cool Roof Chocolate Color Roof Standard Roof Cool RoofSolar Reflectance 0,21 0,44 Solar Reflectance 0,12 0,41Thermal Reflectance 0,8 0,8 Thermal Reflectance 0,8 0,8SRI 15 47 SRI 3 42Roof Temperature (oC) 76,7 64,8 Roof Temperature (oC) 81,2 66,4

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different climate zones are considered in Figure 6. It assumed that residential house average roof is 130 m2 (1400 ft2) in USA. In Table 5 shows that Zone 5 is most energy savings area among all five zones. Zone 1 is the worst energy savings zone. Zone 3, Zone 4, and Zone 5 are three zones that energy savings per year has positive values. Zone 1 and Zone 2 has negative values that heating savings values are more than cooling savings values.

Table 2. Energy savings with cool roof at selected cities in five different climate zones in USA

Figure 6. Climate zones in USA [19]

DISCUSSION The high emittance material used for the cool roof system does not allow energy radiation into the atmosphere. Ceilings under roofs get cooler. Therefore, a cool roof can reduce air temperatures inside buildings with and without air conditionings. With cooler daytime temperatures, buildings use less air conditioning and save additional energy. A cool roof transfer less heat to the building below so the building stays cooler and uses less energy for air conditioning. In this study, cool roofs impact on energy saving were studied at various climate zones in USA showed that if heating degree days is less than 5,499, net savings is beneficial for residential houses. Therefore, human health and comfort get better with cool roofs. The use of cool roofs brings also many benefits such as lower energy use, reduced air pollution, heat islands and greenhouse gas emissions. Cool roof decreases the production of associated air pollution and greenhouse gas emissions (carbon dioxide (CO2), sulfur dioxide (SO2), nitrogen oxide (NOx) and mercury (Hg)) by lowering energy use. This results in a reduction in the CO2 emissions from electricity generating power plants.

Climate Zones City, StatesResidential(cents per

Kilowatthour)*Cooling savings ($/m2) per year

Heating savings (heating penalties if negatives) ($/m2

per year)Net savings

($/m2 per year) $/per yearZone 1 Rochester, New York 18.44 0,818 -1,861 -1,044 -135,80Zone 2 Cleveland, Ohio 12.55 0,829 -1,227 -0,398 -51,80Zone 3 Lexington, Kentucky 10.36 1,280 -0,893 0,387 50,40Zone 4 San Diego, California 18.38 1,237 -0,452 0,785 102,20Zone 5 Houston, Texas 11.65 2,195 -0,398 1,797 233,80

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REFERENCES 1. From America to Europe the ‘’Intelligent Cool roof’’ for energy saving and thermal comfort in building. Cool roof

system, Italiana Membrane, Green Building Council, Italia. 2. Akbari, H, Levinson, R, Rosenfeld, A. and Elliot, M. 2010. Global cooling: Policies to cool the worlds and offset

global warming from CO2 using reflective roofs and pavements. Lawrence Berkeley National Laboratory. 3. Oke, T.R, Johnson, D.G, Steyn, D.G. and Watson, I.D. 1991. Simulation of surface urban heat island under ‘ideal’

conditions at night- part 2: diagnosis and causation. Bound-Layer Meteor. 56 (339-358). 4. Howard, L.1883. The climate of London. vols. I-III, Harvey and Dorton, London. 5. Bulut, M. and Sozbir, N. 2012. Serin çatı sistemi ile ısıtma havalandırma ve iklimlendirme sistemlerinde enerji

tasarrufu. 1. Ulusal İklimlendirme Soğutma Eğitimi Sempozyumu, Balıkesir (in Turkish). 6. Kolotontroni, M, Giannitsaris, I, and Watkins, R. 2006. The effect of the London urban heat island on building

summer cooling demand and night ventilation strategies. Solar Energy, Vol.80 (4) pp 383-392. 7. Santamouris, M. 2007. Heat island research in Europe- the state of the art. Advanced Building Energy Research,

Vol.1, number 1, pp 123-150. 8. Taha, H, Chang, S.C, and Akbari, H. 2000. Meterological and air quality impacts on heat island mitigation measures

in the three U.S. cities. Lawrence Berkeley National Laboratory Report LBNL-44222, Berkeley, CA. 9. Santamouris, M, Synnefa, A, Zinzi, M, and Carnielo, E. 2007. The database of cool roof materials; products, testing

procedures and results. Report of Cool Roofs Project supported by IEE. Save. 10. Heat Island Group http://eetd.lbl.gov/HeatIsland 11. http://garoofingandrepair.com/elastomericroofing/). 12. Kolokotroni, M. and Warren, P. 2011. Cool roofs promotion of cool roofs in the EU. Cool Roofs Supported by

Intelligent Energy Europe. Contract No: EIE/07/475/S12.499428. WP3: Technical Aspect of Cool Roofs. 13. Bretz, S. and Akbari, H. 1997. Long-term performance of high albedo roof coatings. Energy and Buildings, 25, 159-

167. 14. Synnefa, A, Santamouris, M, and Livada, I. 2006. A Study of the thermal performance of reflective coatings for the

urban environment. Solar Energy Journal, Vol.40, pp.968-981. 15. Siegel, R, and Howell, J. 2002. Thermal radiation heat transfer. 4th edition, Taylor and Francis, NY. 16. Konopacki, S, Gartland, L, Akbari, H, and Rainer, I. 1998. Demonstration of energy savings of cool roofs. Paper

LNBL-40673. Lawrence Berkeley National Laboratory, Berkeley, CA. 17. http://www.bestroofing.net/roofingblog/wp-content/uploads/2011/10/White-Roof-Alliance-single-10.png 18. http://web.ornl.gov/sci/roofs+walls/facts/CoolCalcEnergy.htm 19. http://www.eia.gov/consumption/residential/maps.cfm

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[Abstract:0100][Energy Efficient Buildings] EVALUATING THE LCA OF A HOUSE WITH THREE DIFFERENT WALLS

Annual energy consumption and annual global warming potential (GWP) decreases with improving the energy performance of the facade, whereas the embodied energy and embodied GWP increases due to the extra material and products applied. This study analyses the relation between the embodied energy and the energy consumption of a house during the life span of the buildings, and the results represented separately in tables and figures. The study uses LCA framework as a tool to conduct a partial LCA, from cradle to site of the construction and energy consumption during usage

phase of the buildings with three different wall types through 50 years usage phase. According to this study, laminated timber and aerated concrete are better choices than cast concrete for both types of buildings because of lower density and lower U value.

Turkey’s importance in the energy markets is growing, both as a regional energy transit hub and as a growing consumer. Turkey’s energy has increased rapidly over the last few years and likely will continue to grow in the future. Turkey imports nearly all of its oil supplies. Turkey is increasingly dependent on natural gas imports as its domestic consumption rises each year. Natural gas is used domestically mainly in the electric power sector[1]. Hence, mandatory using LCA for all types of buildings can reduce country’s demand and affiliation to other states.

Studies assuming that operational energy contributes 52% to 82% of the total life cycle energy consumption during 40 to 50-yearlifespan. In addition, embodied energy consumes the rest of the LCA behind the little impact of demolition and transportation [2]. This is clear that with increasing the life span of a building the significance of the embodied energy becomes inconsiderable [3].

From life cycle point of view, there is a limit to the thickness of insulation can be applied on external walls [4]. Effort should be paid to the reduction of the energy consumption during the usage phase, as this stage still has the largest potential for improvement, both for new and old buildings [5]. So, maybe using an innovative material such as phase change materials in the building can be a real solution for this.

Values of embodied energy and amount of gas emission from each database are different from one country to another because of energy production, fuel type, transformation process, the technology of the system and how these factors change over time [6].Nevertheless, we use Inventory of Carbon and Energy (ICE) database for carbon footprint studies. As we are comparing different wall types, so it does not affect the results.

Figure 1 express briefly the life cycle steps from cradle to grave. Construction machinery consume lots of energy through extracting the raw materials. Transporting the raw material from mine to the factory, release lots of CO2 to the environment. Raw material should prepare to use as the construction building material and for this purpose facility consumes lots of energy to built the material. Through this process lots of CO2 and energy release. All these steps call cradle to gate, and the data about all of these steps gathered from databases which are mentioned before [7]. Materials must

Mustafa Erkan Karaguler1, Pooya Pakmehr2 1Faculty of Architecture, Istanbul Technical University, Istanbul, Turkey

2Construction Sciences Program, Engineering and Technology, Istanbul Technical University, Istanbul, Turkey karagulerm@itu.edu.tr

SUMMARY

INTRODUCTION

METHODOLOGY AND SIMULATION

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distribute from factory to the distributors or construction site. During transportation, up to the vehicle that does the transportation, lots of CO2 release to the environment.

Figure 1 cradle to grave process

Use phase starts with the occupants and their demand for electricity, heating and cooling, which is calculated with Design Builder simulation software. The last stage of the LCA according to ISO 14040 is disposal and consequently the emission parts. On-site construction and demolition contribute almost 1% of LCA at the end of its service life, both are ignored

through this study [8]. This study does not calculate the effect of other gas emissions and disposal.

Case studies are related to a house with three levels that located in Istanbul, Turkey. Three story house with 152 m2occupied floor area that occupants stay 12 hours a day at home except holidays. The object studied are an external wall of the building with the latitude of 40.97 and longitude of 28.82.

This study uses three different wall types with 3 different U values and thickness of almost 0.23 meters. According to TS 825 (2008), the standard wall for Istanbul should have 0.6 W/m2Kand 0.4 W/m2K for flat roof and 0.6 W/m2K for the ground floor. For this study, three different wall types were considered with various U values and different materials. The U value for all glazing types set to be 2.4 W/m2K. Tables below shows the detailed feature of each wall type, which stucco is the inner face of the wall and marble tile, laminated timber and external rendering are the outer layer of the wall.

Table 1 Wall type 1 Material Thickness (m) U-value W/m2K Stucco

Cast concrete Aerogel

Cast concrete Marble tile

0.019 0.08 0.02 0.08 0.03

0.559

Table 2 Wall type 2 Material Thickness (m) U-value W/m2K Stucco

Softwood Rock wool

laminated timber

0.019 0.04 0.08 0.09

0.346

Table 3 Wall type 3

Material Thickness (m) U-value W/m2K Stucco

Aerated concrete EPS

External rendering

0.019 0.17 0.013 0.019

0.6

The life cycle analysis (LCA) recommends the use of PCM in buildings [9]. There are different ways to use aerogels for insulation applications. Properties of an aerogel product strongly depend on the preparation. In thermal insulation, aerogel products show outstanding performance compared to traditional materials [10]. Using aerogel between two concrete should be valuable, so the first wall consists of two cast concrete with aerogel between them. Timber is the superior

Condition

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material for structural and cladding systems, so the second wall designed to include timber products [3]. To challenge the issue the third wall chose to design with aerated concrete and EPS (expanded polystyrene).

Table 4Technical details and embodied energy/carbon of materials which are used in walls [7, 11]. Material Density

Kg/m3 Material

Quantity m3 Embodied

energy MJ/Kg Embodied Carbon

Kg CO2/Kg

Wall type 1 Stucco

Cast concrete Aerogel

Marble tile

1800 2000 120 2800

5.82 46.6 6.11 8.73

1.8 0.75 53

3.33

0.12 0.107 4.2

0.187

Wall type 2 Stucco

Softwood Rock wool

Laminated timber

1800 510 92

470

5.82 11.64 23.28 26.19

1.8 7.4 16.6 12

0.12 0.39 1.2

0.45

Wall type 3

Stucco Aerated concrete

EPS External rendering

1800 500 25

1300

5.82 49.47 3.78 5.52

1.8 3.5

109.2 10

0.12 0.3

4.39 0.65

As shown in figure 2, the contribution of each material embodied energy in the wall presented. The quantity of stucco for all wall types is the constant amount so the EBE will be the same. Wall type 3 with aerated concrete and EPS have the minimum EBE among others while the second wall with soft wood, rock wool and laminated timber with total 246049 MJ have a maximum of that. Dark blue, light green and orange colour show the amount of insulation on each wall. Aerogel insulation represents almost 19% of wholly embodied energy in the wall type one while the rock wool depicts 15% effects on wall type 2, and EPS in wall 3 just include 9% of the whole wall since a little amount of that used. Besides, in wall three aerated concrete was used as insulation. The external surface of these walls has the theconsiderable difference between each other. As figure 1 shows, marble tile has 81398 MJ embodied energy, which is more than 39% of whole EBE of wall type 1. Laminated timber with 147711 MJ embodied energy includes almost 60% of wall type 2.

Figure 2 embodied energy of each wall

Considering each material in a wall may be a better choice to compare embodied energy because most of the materials have limitation about their thicknesses and we do not allow using thinner than that amount. Therefore, if the wall will design to

be thinner, the embodied energy of the whole wall will be reducing.

In terms of embodied CO2, all three types of walls have impacts between 11000 and 19000 KgCO2/Kg. Cast concrete in the first wall contains 53% of total CO2 emissions. The laminated timber in the second wall with 48% effect and aerated concrete in the third wall with 54% footprint are materials in each wall that have maximum impacts. Among these

EVALUATING EMBODIED ENERGY

Greenhouse gas emission

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insulations, aerogel with 3078 and rock wool with 2569 KgCO2/Kg have close CO2 emissions, despite being different in the volume amount (24 m3 for wool and 6 m3 for aerogel). As we can distinguish from figure 3, wall type 1 with total 18869 Kg CO2 has maximum impact, while the wall type 2 has 38% less embodied carbon than the first wall, and wall type 3 has 27% less embodied carbon.Several observations can be achieved from Fig. 3, since the thickness, density and the quantity of material have a significant role in evaluating the footprint of each building.

Figure 3 embodied carbon of each wall

This study attempts to calculate the CO2 emission during transportation with different vehicles such as, truck and ship. Most of the information about the embodied energy and embodied carbon of materials derived from Sustainable Energy research team of BATH University. Gabi simulation used to calculate the CO2 emission during transportation. Inputs for this software include distance and material weight while the total volume of the walls is almost 67 m3. From table 5, wall type 1 with total 128773 Kg has the maximum amount of CO2 emission, while wall type 2 with 30862 Kg is the lowest. Therefore, the wall with the utmost weight has higher CO2 emission during transportation.

Table 5CO2 emission during transportation for home material Wall type Wall quantity

(Kg) Kg CO2 emission

1 128773 11500 2 30862 2760 3 42492 3800

Whole CO2 emission through exploiting the material to the site is gathered in figure 4. Wall type 2 have minimum footprint while wall type 1 has maximum effect on the environment.

Figure 4 total CO2 emission from cradle to site

Transportation

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The operation phase includes energy for space heating, cooling, ventilation, appliances, miscellaneous, catering and lighting. The hot water demand and electricity consumption largely depend on the users. Since the building construction has minimum impact on the energy demand for hot water and household electricity. These requirements were regulated for all buildings using Turkish Standard Institution [12, 13]. Home electricity and lighting for all walls are the same, since changing the wall type do not affect them. Home heating for both of the wall one and wall three are the same because they have close U value, whereas building with laminated timber wall has minimum heating consumption.

Figure 5 Operating energy of home Figure 5 shows the distribution of different operating energies in two building. Wall type 2 consumes less energy because its U-value is less than the other two walls. Wall type 3 consume a little bit more energy than wall type one, and this is

because of the small difference between U values of them. Despite climate and other differences, the study of some residential and non-residential cases from nine countries revealed a linear relation between operating and total energy[6]. Figure 6 shows the life cycle energy of two buildings with 3 wall type in a home over 5, 10, 15 …, and 50 years. It includes embodied energy of the whole external façade at the end of construction, embodied carbon from cradle to gate, and the operating energy. This amount will increase gradually during usage phase of the building. Since each year, both buildings consume a high quantity of energy to prepare comfort condition for occupants. Figure 6 consists of triple lines which relate to the house with three different wall types. LCA amounts increase annually for each year the building consumes some energy. It is obvious that due to aging, defects and material obsolescence the performance of all walls will be reduced but as this fact occurs for all walls, it does not change the results much more. In construction part, home with wall type 1 has the highest life cycle among others. The third wall type has minimum life cycle assessment. During usage phase building with wall type 1 and 3 consumes more operating energy. Yearly energy consumption adds up to the previous year and through each 5-year, the total amount rolled in (figure 6). Wall type one with higher embodied energy, higher embodied CO2 and mean U value is a pioneer in LCA for all 50 years. Comparing wall type two and three at each triple shows that the wall type 2 has embodied energy and embodied carbon more than wall type 3 while it has less U value. After 10 years, LCA for all wall types becomes almost the same and after 15 years buildings with wall type 1 and 3 passes wall type 2. During 50 years usage phase, LCA of building with wall type 3 become closer to the building with wall type 1. From figure 4 and 6 we can conclude that second wall in every building is the better choice albeit the building should be used more than 10 years even if it has high embodied energy. Moreover, this type of wall has minimum impact on the environment because of its lower weight. Also, it releases a minimum amount of CO2 through transportation.

From the figure, it is distinguishable that for wall type 1 CO2 emission during transportation increase the footprint almost 61% while this percentage for light materials is approximately 22%. Therefore, transporting heavy materials may increase the CO2 emission more than 60% of wholly embodied carbon of the building.

Operating energy

LCA in building scale

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Figure 6 LCA of two building during 50 years life span The construction industry uses vast quantities of raw materials that also has a high amount of energy consumption. Choosing materials with high embodied energy cause not only a high level of energy consumption in the building production stage but also determines future energy consumption in order to fulfil heating, cooling, ventilation and air conditioning demand [14]. Therefore, outcomes of this paper summarized as:

Thickness, quantity and density of each material play a significant role in embodied energy of a wall. Laminated timber with embodied energy of 12 MJ/Kg and density of 470 is tolerable among the other materials of table 4. In the time, using a high amount of laminated timber in wall type 2 caused a significant amount of embodied energy among other walls.

Insulation materials at each wall have the highest embodied energy and embodied carbon, so finding the optimum amount of insulation in each wall helps to reduce the EBE and CO2 emission.

The U value for wall type 2 is less than the other two walls, accordingly operating energy consumption by the building with wall type 1 and 3 will be more than the second wall, and this variation will gradually increase during the usage phase.

Building with laminated timber and aerated concrete is superior to the cast concrete. Moreover, if we do not intend to substitute the second wall material before 10 years laminated timber with soft wood and rock wool should be the right choice.

Even using aerogel between layers of cast concrete does not a good solution for building industry, because CO2 emission of such wall is high.

Although laminated timber is one of the privileged materials, but using a thick layer of that in façade flourish the embodied energy and using a thin layer of that may disturb the performance of the façade.

For most of the buildings around the world with a similar climate to Istanbul, that is made with aerated concrete and EPS are preferable to cast concrete and aerogel.

Different wall types have different embodied energies which are between 25% and 40% of its operating energy over 25 year period while this amount reduces over 50 years lifespan of the building. For this home, the consumption from cradle to the site of wall type one is 20% of whole LCA through 50 years. This amount for wall type 2 is 17% and for wall type 3 is 14%.

With reducing the operational energy, the impacts of embodied energy increases [15]. Rarely the building demolishes before the end of its life, unless, natural disasters such as earthquake does not happen. Therefore, evaluating the life span of each wall will help us to decide more confidently about the performance of that.

CONCLUSIONS

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1. EIA. 2013. Turkey country analysis brief, U.S. energy information administration, available from http://www.eia.gov/countries, accessed on 20st November 2013.

2. Adalberth, K. 1997. Energy use during the life cycle of buildings: a method. Lund University, Sweden. 3. Thiel, c l, Campion, N, et al. 2013. A materials life cycle assessment of a net-zero energy building, energies, University

of Pittsburgh, USA. 4. Ramesh, T, Prakash, R, Shukla, K K. 2010. Life Cycle Analysis of Building: an overview. Motilal Nehru National

Institute of Technology, Allahabad, India. 5. Verbeeck, G, Hens, H. 2010. Life cycle inventory of buildings: A contribution analysis. Building and environment,

PHL University College, Diepenbeek, Belgium. 6. Sartori, I, Hestnes, AG, 2007. Energy use in the life cycle of conventional and low-energy buildings: A review article,

Norwegian University of Science and Technology, Trondheim, Norway. 7. Menzies, G F. 2011. Embodied energy considerations for existing buildings, technical paper 13, Scotland. 8. Ramesh, T, Prakash, R, Shukla, KK. 2012. Life-cycle energy analysis of a residential building with different envelopes

and climates in theIndiancontext, Motilal Nehru National Institute of Technology, Allahabad, India. 9. Cabeza, L F, Barreneche, et al., 2013. Low carbon and low embodied energy materials in buildings: A review,

University of Lleida, Spain. 10. Rovers, R, Kimman, J, Rovesloot, C, 2010. Towards 0-Impact building and built environments, Amsterdam. 11. Hammond, G, Jones, C. 2008. Inventory of carbon and energy (ICE), sustainable energy research team, University of

Bath, UK. 12. Turkish standard regulation on the energy performance of buildings, 2008. No: 27075 13. Turkish Standard, 2008. Thermal insulation requirements for buildings No: 825. 14. Bribian, I Z, Capilla, A V Uson, AA, 2011. Life cycle assessment of building materials: a comparative analysis of

energy and environmental impacts and evaluation of the eco-efficiency improvement potential, University of Zaragoza, Zaragoza, Spain.

15. Moncaster, A.M., Symons, K.E. 2013. A Method and Tool for ‘Cradle to Grave’ Embodied Carbon and Energy Impacts of Uk Buildings in Compliance With The New Tc350 Standards, University of Cambridge, UK.

REFERENCES:

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[Abstract:0101][Exergy] THERMODYNAMIC ANALYSIS OF CENTRAL HEATING SYSTEMS AND

INVESTIGATION EFFECTS OF EXERGY EFFICIENCY OF TEMPERATURE CHANGING IN AIR BURNING

Zafer UTLU 1 , Mesut YENİGÜN2

1 Istanbul Aydın University, Engineering Faculty, Mechanical Engineering Department, Istanbul 2 Istanbul Aydın University, Science Institute, Mechanical Engineering Department, Istanbul

SUMMARY

The main aim of this study, called; “Thermodynamic analysis of central heating systems and examination of various parameters effect on exergy and fuel consumption”. Detail information about energy and exergy calculations of central heating components, like heating boiler, air conditioning plant, fan coil and heat exchanger are presented. The components of central heating systems and energy, exergy efficiencies calculated. The effects of exterior temperature, amount of combustion air, combustion air temperature, hot water flow rate and return water temperature difference on exergy efficiency and fuel consumption analysed and these effects expressed by digital data. System enhanced theoretically. Keywords; energy efficiency, energy analyses, exergy analyses, central heating systems, reduction of fuel consumption INTRODUCTION Approximately 35% of the energy consumption in our country and the world are used in the building. In about 80% of the energy used in buildings is spent for heating, cooling, air conditioning and hot water supply. The average residential energy consumption in Europe is 100 kWh / m^2 , in Turkey is found normally 200 kWh / m2 .The average energy consumption in passive house building design is criteria consumed under of 15 kWh / m2 [1]. Values are investigated when energy consumption in homes nationwide, and it reached twice the amount of energy consumed is compared with households in Europe. This energy consumtion gives that short and outstanding information about our housing. In this studies performed suggests that we have a great potential for energy savings. Even in this case, established in the country in our building HVAC (Heating Ventilation and Air Conditioning) systems is not adequate. When installing HVAC systems, starting from the project stage, Construction and operation process takes into consideration the cost of life, energy and exergy analysis should be opened the system by the establishment of minimum energy consumption. In addition, building energy use in our country is predicted to decrease by 30-40%.With these savings potential, our country is dependent on foreign energy imports 72%, this issue demonstrate the necessity of more and more number of studies and analyzes [2]. Central heating systems and thermodynamic analysis and investigation of the effect of various parameters on exergy fuel consumption; Heating boilers, air handling units, fan coils and heat exchangers, such as detailed information about central heating component of the energy and exergy calculations are presented. Central heating system components by examining the energy and exergy analysis were conducted and each energy and exergy efficiency is calculated. DESCRIPTION OF SYSTEM

Model building, was selected an existing building in İstanbul Bakırköy which has a covered area of 17,000 m^2. Total heat load of the building is 1750 kW. The central heating system with natural gas-powered is made 2 heating boiler. As the heat exchanger has 8 air handling and 160 fan-coil unit. Summer cooling is done with air-cooled chiller unit. The air handling unit is designed as a Line 2-pipe fan coils and 4-pipe system lines. The model building, can be shown in Figure 1.

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Fig.1. Plan of heat center

System Components A hot water system is generally hot water boiler, water carrier pipes, heating elements, though the circulation pump, expansion tank, consisting of automatic control devices and a variety of equipment and spacers. Hot water boiler produced in volumes to be heated by hot water radiators mounted pipe fan coils, air conditioning plant and transported to the heating elements such as hot air apparatus. Here volume by cooling to room temperature, leaving the hot water returns to the boiler. Natural water circulations in the old system (by gravity) are provided with the circulation pump for the new system is more economical and comfortable. Increasing volume during the heating of the water available in the system is collected in a tank conservator given name. Employees used electronic panel system according to the outside temperature and the modern system. Water temperature 90/70 ° C instead of 70/55 ° C is selected and the low temperature heating comfort achieved. Model building's heating, cooling and ventilation systems, devices, and equipment vendors are listed in Table 1 with the quantities and specifications. Table 1. Model building equipment list

Equipment Quantity (pcs)

Specifications

Heating Boiler 1 1.000.000 kcal/h, Construction pressure 3 bar Heating Boiler 1 800.000 kcal/h, Construction pressure 3 bar

800 000 k l / h 3 b C iHeating Expansion Tank 2 1500 LT Boiler Chimney 2 Ø450 Stainless Steel Burner 1 1163 kW

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Burner 1 1512 kW Heating Boiler 1 500 LT

Heat Exchanger 1 25,000 kcal / h, Plate Heat Exchanger, Test Pressure 8 bar

AC line heat pump 2 12,7 m3/h - Motor 2,2 kW - 1450 d/d - 11 mss Fan coil line heat pump 2 10 m3/h - Motor 0,75 kW - 1450 d/d - 9 mss DHW pump 2 5 m3/h - Motor 0,37 kW - 1450 d/d - 7 mss Radiator heating pump 1 3,9 m3/h-Motor 2,2 Kw-1450 d/d-7mss Boiler Circulation Pumps 1 33,4 m3/h - Mot1,1 kW - 1450 d/d - 5 mss Heat Exchanger Circulation Pump 1 2,3 m3/h-Motor 0,4 kW-1450 d/d-5mss Air-condition Unit 1 %20 - 26.500m3/h - IK123kW -SK65kW - Motor 11kW Air-condition Unit 1 %100 - 17.700m3/h - IK184kW -SK122kW - Motor 11kW Air-condition Unit 1 %100 - 6.100m3/h - IK66kW -SK40kW - Motor 3kW Air-condition Unit 1 %100 - 13.300m3/h - IK138kW -SK95kW - Motor 7,5kW Air-condition Unit 2 %100 - 7.500m3/h - IK74kW -SK50kW - Motor 4kW Air-condition Unit 1 %100 - 3.500m3/h - IK30kW -SK30kW - Motor 2,2kW Air-condition Unit 1 %100 - 4.500m3/h - IK39kW -SK30kW - Motor 2,2kW Fan 2 7500m3/h - Motor 2,2kW Aspirator 1 14.310m3/h - Motor 5,5kW Aspirator 1 4.140m3/h - Motor 1,5kW Aspirator 1 11.970m3/h - Motor 5,5kW Aspirator 1 3.150m3/h - Mot1,5kW Aspirator 1 4.050m3/h - Mot1,5kW Aspirator 1 14.000m3/h - Motor 3kW Fan coil 40 Heating Capacity 1,9 kW Fan coil 50 Heating Capacity 2,6 kW Fan coil 70 Heating Capacity 3,2 kW

THERMODYNAMIC ANALYSIS OF SYSTEM Fan coil Energy and Exergy Analysis Fan coil, heat gain obtained from the water circuit, the scene with the airline uses for heating and cooling. Calculation of yield fancoil energy and exergy analysis is illustrated by the following formula. Fan coil mass balance equations;

∑ ∑ (1)

Fan coil energy balance equations;

, , (2)

, (3)

Energy efficiency of Fan coil;

, (4)

Fan coil exergy balance equations;

, , (5)

∅ 1 (6)

, (7)

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Fan coil exergy calculation of earnings obtained from the water circuit equations;

ΔĖx , , . (8)

Exergy efficiency of fancoil equations;

, (9)

The analysis of fan coils in different capacities in the existing building has been made and is shown in the following table 2. Table 2. Fan coil exergy and energy efficiency

Heat Exchanger Energy and Exergy Analysis Heat exchangers, heat gain obtained from the hot water circuit, Use cold water to meet the hot water demand by transferring. Heating boiler water temperature entering the heat exchanger from 90 ° C, the temperature of the return water at 70 ° C. Cold water enters the heat exchanger at 10 ° C, 60 ° C using the hot water is sent into the building. Heat exchanger efficiency calculation of energy and exergy analysis is illustrated by the following formula. Heat exchanger energy balance equations;

(10)

Energy efficiency of Heat exchanger;

(11)

Heat exchanger exergy balance equations;

(12)

(13)

Hot water heat exchanger calculation of earnings obtained from the circuit exergy equations;

ΔĖx . (14)

Heat exchanger cold water circuit exergy amount Transferred equations;

ΔĖx . (15)

Exergy efficiency of heat exchanger;

(16)

The analysis of heat exchangerin the existing building has been made and is shown in the following Table 3.

Table 3. Heat exchanger exergy and energy efficiency

Fancoil ΔĖfc.w= ØH

(W) Qfcu (W)

ṁfc.w (kg/s) ΔĖxfc.w (W) Ėxfc (W)

Ėxfc,dest (W) Ψexergy ηenergy

Type1 1570 1430 0,0375 339,77 131,71 208,06 0,388 0,91

Type2 2225 1992 0,05314 481,52 186,66 294,86 0,388 0,90

Type3 3000 2711 0,07165 649,24 251,68 397,56 0,388 0,90

Heat Exchanger

ΔĖhot.water

(W) ΔĖcold.water

(W) ṁ.water (kg/s)

ΔĖxhot.wat

er

(W)

ΔĖxcold

water

(W) Ėloss (W) Ψexergy ηenergy

Type1 29307,6 26167,5 0,35 6676,3 2928,1 3140,1 0,438 0,893

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Air-Condition Energy and Exergy Analysis Air condition units, fan coil that such water heat gain obtained from the circuit, the scene with the airline uses for heating and cooling. Air-conditioning plant efficiency by calculating the energy and exergy analysis is illustrated by the following formula: Air Condition mass balance equations;

∑ ∑ (17)

Air Condition energy balance equations;

, , (18)

, (19)

Air-Condition water obtained from the circuit power gain;

ΔĖx . . (20)

The calculation of the amount of heat transfer which locus of air condition plant;

Q . (21)

Energy efficiency of Air-Condition;

, (22)

Air-Condition exergy balance equations;

(23)

(24)

Air Condition water obtained from the exergy circuit gain;

ΔĖx . (25)

Calculating the Amount of Mahale given Exergy Air Condition;

∅ 1 (26)

Air Condition exergy destruction;

(27)

Exergy efficiency of air-condition;

, (28)

Analysis of air conditioning plants in different capacities in existing buildings made and shown in the following Table 3. Table 3. Exergy and energy efficiency of air condition

Air Condition

ΔĖAc.water= ØH (W)

QAc (W)

ṁAc.water (kg/s)

ΔĖxAc.water (W)

ĖxAc (W)

ĖxAc.dest. (W) ηexergy ηenergy

AC01 275000 245711 3,28 62645,46 23070,47 39574,99 0,368 0,89 AC02 250000 223030 2,99 56950,42 20973,15 35977,26 0,368 0,89 AC03 87500 76863 1,04 19932,65 7340,60 12592,04 0,368 0,88 AC04 187500 167588 2,24 42712,81 15729,87 26982,95 0,368 0,89 AC05 105000 94504 1,25 23919,17 8808,72 15110,45 0,368 0,90 AC06 105000 94504 1,25 23919,17 8808,72 15110,45 0,368 0,90 AC07 49500 44102 0,59 11276,18 4152,68 7123,50 0,368 0,89 AC08 63500 56703 0,76 14465,41 5327,18 9138,22 0,368 0,89

Heating Boiler Energy and Exergy Analysis

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Figure 2 is a schematic diagram illustrating the boiler. The burner portion of the heat energy obtained by the combustion of fuel-air mixture is sent as reactive hot water into the building. Energy for heat transfer results in reduced water returns to the boiler to reheat places. Resulting combustion products are released to the atmosphere as flue gas.

Fig. 2. Heating Boiler

Natural gas is used as fuel to heat boilers examined. Gas largely methane (CH4), in a lower rate of ethane (C2H6), propane (C3H8), butane (C4H10), nitrogen (N2), carbon dioxide (CO2), hydrogen sulfide (H2S) and helium (He) As contains various hydrocarbons. This component of the rate varies according to the gas supply [5]. Heating boiler primarily for energy analysis and system building heat loss calculation should be made. In this study, the outside air temperature (T0 = 0 ⁰C) for building heat loss was calculated as 1750 kW. Mass balance and energy flow necessary to perform an energy analysis above shows the volume control. According to thermodynamic analysis, the mass and energy balance for the control volume is expressed as follows [4]. Heating Boiler Energy Analysis Assuming that all of the gas consists of methane, natural gas combustion equation can be written as follows: CH4 + 2 (O2 + 3.76N2) → CO2 + 2H2O + 7.52N2 Table 4. Entering the combustion reaction and enthalpy values of the compound [4].

Compounds

0

(k J/ km o l)fh

(273 )(kJ/ kmol)h K

(450 )

(kJ/ kmol)h K

0

(kJ/ kmol)h

Reacting compounds CH4 -74850 0 - 0 O2 0 7946 - 8682 N2 0 7937 - 8669

The (reaction of the compounds)

CO2 -393520 - 15483 9364 H2O -241820 - 15080 9904 N2 0 - 13105 8669

∑ . ∑ . (29)

nr: number of moles of the reacting chemical compounds nü : The number of moles of reaction from the chemical compounds

(30)

(31)

(32)

:Heat energy released during combustion value (kW); : 1750 kW (T0 = heat loss of the building to 0 ° C)

η = 0.90 (according to the boiler manufacturer's catalog information, boiler combustion efficiency)

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Heating Boiler exergy balance equations;

. . (33)

(34)

Formed in the boiler heating the total irreversibility;

. . (35)

(36)

Calculation of earnings obtained from the exergy heating boiler Water circuit;

ΔĖx . (37)

ΔĖx : Exergy obtained from the water loop gain (kW)

Exergy efficiency of Air-Condition;

(38)

For heating boilers, calculating TO = 0C and 90-70˚C ranges is shown above in detail. The calculations TO = -6, -3,3,6,9,12,15˚C and 90-70˚C range is performed in a similar manner is shown in Table 5. Table 5. TO = -6, -3, 0, 3, 6, 9, 12,15˚C and 90-70 ˚C thermodynamic analysis on working conditions

Qısıtma (kW)

Qyanma (kW)

qyanma (kj/kmol)

T0 (°C)

Ėxyakıt (kW) ∆Exsu (kW)

ṁyakıt (kg/h)

Ėxters. (kW)

ηekserji %

2095 2327,78 46364,65 -6 2977,55 427,405 180,74 2550,14 0,144 1923 2136,11 46416,87 -3 2729,40 412,545 165,67 2316,86 0,151 1750 1944,44 46469,09 0 2481,80 397,684 150,64 2084,11 0,160 1578 1752,78 46520,72 3 2234,76 382,823 135,64 1851,94 0,171 1405 1561,11 46572,94 6 1988,24 367,963 120,67 1620,27 0,185 1233 1369,44 46624,69 9 1742,26 353,102 105,74 1389,16 0,203 1060 1177,78 46676,91 12 1496,80 338,242 90,84 1158,56 0,226 888 986,11 46728,53 15 1251,89 323,381 75,97 928,51 0,258

Energy Consumption

For model building in Istanbul province in 2013 belonging to the December-January-February 3 months total, amount of energy used in this period was 827.921kWh and the total energy consumption cost paid 117.129 TL. Model building 3 monthly energy consumption per m2, 48.7 kWh / m2. Taking into account 3 monthly measurements the annual energy consumption 194.8 kWh / m^2. Total amount of energy used for heating is 699.448 kWh and the consideration paid 70.961 TL(unit price of Natural Gas 0.10145 TL / kWh. IGDAS in 01/03/2013) and total of electricity consumption for the same period is 128.473 kWh and consideration paid 46.168 TL (Unit Price of Electricity 0.360 TL / kWh. Istanbul BEDAS in 01/03/2013 date, is the unit price of electricity retail tariff, including KDV.) [3]. In this period, 84% of the energy consumption by using natural gas for heating purposes, while the rest was realized in the form of 16% of electricity consumption. The energy consumption model building can be shown in Figure 2. a) kWh b) TL

Fig. 2. Energy Consumption. a) kWh, b) TL

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CONCLUSION "Investigation of the effect of Center Thermodynamic analysis of Exergy efficiency of heating and combustion air temperature change" is the content of the work; heating boilers, air handling units, fan coils and heat exchangers, such as detailed information about central heating component of the energy and Exergy calculations are presented. Central heating system components by examining the energy and Exergy analysis were conducted and each energy and exergy efficiencies are calculated. Outside temperature, combustion air temperature and roundtrip examined the effects of the water temperature difference of the Exergy efficiency and fuel consumption, and these effects were expressed as numerical data and made theoretical improvements on the system. Energy efficiency for fan coil units of type 1 is 0.91 and Exergy efficiency is 0.388; for AHU 01 energy efficiency is 0.89 and Exergy efficiency is 0.368; energy efficiency Boiler is 0.90 Exergy efficiency is 0.16; Exchanger unit energy efficiency was calculated to be 0.893 and for Exergy efficiency to be 0.438 which Examined with several variables; when outside air temperature increases, irreversibility the amounts reduced, Exergy has been shown to decrease the amount of consumed fuel and the yield increased. Exergy efficiency in -6˚C of the external weather conditions is 0.144, fuel consumption is 180.74 kg / h; Exergy efficiency in -3˚C weather conditions is 0,151, fuel consumption is 165.67 kg / h and Exergy efficiency is 0.160 and fuel consumption 150.64 kg / h in the 0˚C weather conditions and where the outside air temperature was observed with the increase of Exergy efficiency. The fuel consumption of model building; boiler runs at 0˚C and range 90-70 ° assuming that calculated for a month heating season total amount of fuel consumed was found 1,710,818.48 kWh. The total cost of fuel consumed for a season was calculated as £ 171,595.09. TO = 0C and 90-70 ° C operating range of 1 m3 of natural gas combustion is carried out with 100% fresh air Exergy efficiency is realized in 0160. When combustion occurs with 90% fresh air and combustion occurs when the Exergy efficiency is 0.144 with 80% fresh air exergy efficiency is 0.128. Therefore when the amount of combustion air decreases Exergy efficiency also decreases. REFERENCES 1] S. Tekin., Central heating systems, thermodynamic analysis and investigation of the effects of Exergy efficiency and fuel consumption of various parameters change Supervisor: Assoc. Dr. Zafer Utlu., Istanbul Aydin University, Institute of Science and Technology, Department of Mechanical Engineering., Discipline Code: 625.04.00.Sunuş Date: 07/07/2013 [2]Ministry of Energy and Natural Resources, the www.enerji.gov.tr area [3] ", the area www.igdas.com.tr [4] Utlu, Z. and A. Hepbaşlı, "Parametrical Investigation of the Effect of the Dead (Reference) State on Energy and Exergy Utilization Efficiencies of Residential-Commercial Sectors: A Review and the Application", Renewable and Sustainable Energy Reviews, 11 (4), 603-634 (2007). [5] IGDAS, "Gas & Technology / Gas / Chemical and Physical Properties

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[Abstract:0102][Comfort Cooling] EXPERIMENTAL INVESTIGATION OF THE EFFECTS OF USING EVAPORATIVE

CONDENSER IN COOLING SYSTEMS

Abdulaziz Yıldız1, Ali Etem Gürel2, Emrah Deniz3

1 Karabuk University, Graduate School of Natural & Applied Sciences, Department of Mechanical Engineering, 78050, Karabuk

2 Duzce University, Vocational School, Department of Electrical and Energy, 81010, Duzce 3 Karabuk University, Faculty of Engineering, Department of Mechanical Engineering

78050, Karabuk Corresponding email: etemgurel@gmail.com

SUMMARY Energy efficiency of cooling systems can be improved by replacement the current system with higher performance systems, improving the operating conditions and using alternative cooling systems. These methods help to decrease costs, improve the service life and provide a more effective cooling. Adiabatic cooling method (adiabatic humidification) improves the system performance without increasing the energy consumption.

In this study, an adiabatic humidification system integrated to a cooling system was investigated theoretically and experimentally in terms of improving energy efficiency and cooling capacity. For this propose, an adiabatic humidification system was integrated to a cooling system and tested under actual conditions. Results show a decrease in the compressor work by 8.48% and increase in cooling system COP by 32.92% and in COP value of the whole system by 18.43%. It can be concluded that by integrating an adiabatic humidifier the energy consumption of cooling system can be decreased. Keywords: adiabatic humidification, COP, energy efficiency. INTRODUCTION Efficient use of energy is an important concern today and the civilization level of a country is defined as the energy intensity (ratio of added value to energy consumption). 30% of the global energy consumption today goes to the air conditioning applications [1]. Since one half of the energy consumed in air conditioning applications is used for mechanic vapor compression systems, improving the energy efficiency in cooling is very important [2]. Refrigeration is used both in domestic areas and industrial applications. Refrigeration systems are used to transfer heat from a low temperature area to a high temperature one in order to cool that area. Refrigeration systems are more important for the industry than they were before. Depleting energy resources and little use of renewable energy resources force the humanity to use existing energy resources carefully. Energy efficiency in refrigeration systems can be improved by decreasing the energy consumption without deteriorating the refrigeration quality. Preventing loses during the energy consumption, reusing the waste heat by heat recovery methods and some technological developments improve the efficiency. Saving energy is the main concern for improving energy efficiency by minimizing the energy loses during the energy production and consumption. Energy efficiency in refrigeration system can be improved by combined use of cooling towers, direct and indirect evaporative cooling and conventional vapor compression units [3]. Adiabatic cooling (humidification) method can improve the performance of a cooling system without increasing the energy consumption. There are many researches on the literature on adiabatic refrigeration systems. Korun (2011) investigated the advantages of evaporative cooling system types of water spraying and cellulosic pad on the air cooling groups [4]. Bilge (1999) investigated a combination direct and indirect evaporative refrigeration systems and compared it with classical refrigeration systems used in air conditioning and concluded that the proposed system is much more economic in terms of energy consumption at proper ambient air conditions [5]. Atikol and Hacışevki (2001) conducted a feasibility study of evaporative cooling for Lefkoşa (Cyprus) region [6]. Şen et al. (2011) made performance evaluation of evaporative refrigeration systems for İzmir (Turkey) and presented the benefits of these systems by an example [7]. Idrissi et al. (2007) developed a condenser model which is cooled by air-water spray. Based on the mass transfer laws for the air-water mixture sprayed onto the condenser they concluded that water spraying method, droplet diameter, location of the sprayer, spray flow rate and geometry of heat exchanger are the main factors that should be considered [8]. Hajidavallo and Eghtedari (2010) perform experiments to improve the COP of an air conditioner by spraying water onto the air cooled condenser surface. They found that the condenser with evaporative cooling decreased the energy consumption by up to 20% and increased the

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COP by up to 50% [9]. Hu and Huang (2005) used cellulosic pad on an air cooled cooling tower and investigated the performance change experimentally. Their results showed a decrease in condenser temperature and power consumption of the compressor along with an increase in COP value up to 3.45. They also emphasized that the system has a payback time of as short as ten months [10]. Yanga et al. (2012) investigated an air cooled chiller with an adiabatic cooling system and twin cooling system. Experimental results show that condenser inlet air temperature has decreased 8.8 °C, the refrigeration system COP value has increased by 25% and the annual electrical energy consumption has decreased 14.1% [11]. The aim of this study is to determine the ratio at which the adiabatic humidification method contributes to the cooling system performance and energy efficiency. EXPERIMENTAL SETUP Figure 1 shows the indoor and outdoor units of the cooling system used in the experimental study along with measurement setup. Table 1 gives the specifications of the cooling system used in the study. Main components of this duct-type 60.000 BTU cooling system are; scroll compressor, finned tube type evaporator, U type finned condenser and thermostatic expansion valve. Adiabatic humidification was provided by a high pressure pump (55-70 bar) through 0.4 mm diameter nozzles installed on the condenser.

a. Scroll compressor d. Liquid tank g. Expansion valve j. Nozzle group b. Accumulator e. Filter-dryer h. Evaporator k. Water filter c. Oil separator f. Sight glass i. Condenser l. Pump

Figure 1. Schematic view of the experimental system.

Table 1. Technical specifications of the cooling system used in the experiments. Parameters Values Cooling capacity (kW) 17.6

Air flow rate (L/s) 825

Refrigerant R-22

Energy efficiency ratio-EER (BTU/Wh) 9

Compressor type Scroll Current (A) 11.9

System mass (kg) 200

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Adiabatic humidification system used in the experiments saturates the condenser inlet air and brings it to the wet thermometer temperature by spraying water particles under 35 micron through the high pressure nozzles. Figure 2 shows the cooling system outdoor unit integrated with the adiabatic humidification system. Experimental studies were conducted as two stages with/without using the adiabatic humidifier with 0.4 mm diameter nozzles.

Figure 2. Cooling system outdoor unit integrated with adiabatic humidifier.

During the experiments, pressure measurements were done using Gems 2600 series (0-400 bar) pressure transmitter with 0.25% accuracy. Relative humidity of the air was measured by Ram DT 615 device. Energy consumption of the system was measured by Makel T300.2216 type electric meter. All temperature measurements were done by nickel chromium-nickel (K type) thermocouples with ±0.1°C accuracy. Pressure and temperature data were transferred to the computer by Advantech Adam 4018 data acquisition card. THERMODYNAMIC ANALYSIS Performance of refrigeration systems is defined as the coefficient of performance (COP) [12].

comp.

evap

.

W

QCOP (1)

Reverse Carnot cycle is the most effective cooling cycle within a certain temperature range. COP of a reversible cooling machine can be defined as below;

LH

Lcr TT

TCOP

(2)

This equation shows the maximum COP value a cooling machine operating between “TH” and “TL” can reach.

The COP value of all real cooling machines operating within this range would be less than tis value. Therefore, the

reversible cycles define the upper limit of the performance of real cycles.

Condenser, evaporator and compressor capacities for a cooling cycle can be calculated as below [13];

43r.

cond

.hhmQ (3)

67r.

evap

.hhmQ (4)

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21r.

comp.

hhmW (5) There are different energy consuming system components with different structure and properties such as compressor, air fans and water pump.

Considering the energy consumed by these components, the performance the whole system would be;

pump.

fans.

comp.

evap

.

wsWWW

QCOP

(6)

EVALUATION OF EXPERIMENTAL RESULTS This study aims to determine the effects of integrating an adiabatic humidification method to the cooling system on the operating parameters. Therefore, a split air conditioner and adiabatic humidifier were integrated to form an experimental setup and tested under actual conditions. At the first stage of the study, the condenser was used without humidification. At the second stage, a humidifier with 0.4 nozzle diameter was integrated to the condenser and the data was recorded. Table 2 gives the temperature and pressure variations for both condenser cases. Table 2 gives the evaporation and condensing temperatures as -9.6 and 41.75 C respectively for the condenser without humidification. At the same application conditions with adiabatic humidification the same values become -11 and 30.25 C respectively. Condensation temperature of the refrigerant should be as low as possible in operation conditions in order to maximize the COP of a cooling system. Decreasing the condensation temperature by adiabatic humidification method have increased the COP of the system. Refrigerant temperature and enthalpy at the condenser outlet with adiabatic humidification are less than those of conventional condenser without humidification. Therefore, the refrigerant enters the expansion valve as compressed liquid with lower temperature and enthalpy and directly increased the cooling capacity of the system.

Table 2. Experimental Parameters.

Parameters Unit Without water injection

With water injection (0.4mm)

Ambient temperature °C 38 38 Evaporator air inlet temperature °C 25 25 Evaporator air outlet temperature °C 10 6.8 Evaporator pressure bar 3.6 3.42 Condenser Pressure bar 16 12 Evaporator temperature °C -9.6 -11 Condensation temperature °C 41.75 30.25 Compressor inlet temperature °C 1.1 -6.4

Table 3 gives the experimental results of the cooling systems with and without adiabatic humidification (0.4 mm nozzle diameter). It can be seen that humidified cooling system consumes less energy in the compressor by 8.48%. Humidified system also has a higher COP (by 32.92%) and COPws (by 18.43%) than the system without humidification.

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Table 3. Comparison of the systems with and without water injection trough 0.4 mm nozzle.

Parameter Unit Without water injection

With water injection (0.4mm) Difference (%)

compW.

kW 4.6 4.21 -8.48

rm.

kg/s 0.08937 0.10159 13.67

evapQ.

kW 14.8 18 21.62

COP - 3.22 4.28 32.92 COPcr COPws

- -

5.13 2.55

6.35 3.02

23.78 18.43

Water flow rate L/h - 48 -

Figure 3 shows the Mollier diagram of the system developed by using data from the tests on the system with/without water injection through nozzle. Figure 3 clearly shows that adiabatic humidification method improves the energy economy and cooling capacity of the cooling system.

Figure 3. Comparison of the systems with and without water injection.

RESULTS AND DISCUSSION All experimental and theoretical investigations conclude that effective use of the energy sources requires high efficiency systems and low energy losses. Therefore, applications that increase the efficiency and decrease the losses should be promoted. Air conditioning and refrigeration is an energy intensive technology and it is highly required to develop new methods to use the energy more efficiently. Adiabatic humidification is one of the methods that can be used air conditioning and refrigeration applications. Experimental results from a duct type split air conditioner revealed that humidification decreased compressor work by 8.48%, and increased the cooling capacity and COP value by 21.62% and 32.92%, respectively. 18.43% increase in COPws value, which is calculated by including the effects of all components of the air conditioner, shows that a significant amount of energy can be saved. Adiabatic humidification method provides the benefits below along with its positive contributions to the air conditioning and refrigeration applications; 1. It is a reliable and low cost method for improving system performance and efficiency. 2. It can be easily applied to existing air cooled condensers with low cost. 3. Humidifying improves system performance and energy economy more than air cooled condensers. 4. It decreases the pressure ratio of the system and provides a safer compressor operation. 5. It facilitates the use of cooling system under extremely high outside temperatures. Further studies may include using adiabatic humidifying method with cooling systems using different refrigerants under various conditions. Another area of study may be the prevention of condenser fins from corrosion which can contribute to the widespread use of these systems.

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REFERENCES 1. Jahangeer, K.A., Tay A.A.O., Islam M.R., “Numerical investigation of transfer coefficients of an evaporatively-

cooled condenser”, Applied Thermal Engineering 31, 1655-1663 (2011). 2. Coulomb, D., “ IIR listing of refrigeration research priorities”, International Journal of Refrigeration, 28, 973-976

(2005). 3. Ettouney, H.M., El-Dessouky, H.T., Bouhamra, W., Al-Azmi, B., “Performance of evaporative condensers”, Heat

Transfer Engineering, 22:4, 41-55 (2010). 4. Korun, G., “Hava soğutmalı grupların evaporatif ön soğutma ile verim ve kapasite artışı uygulaması”, X. Ulusal

Tesisat Mühendisliği Kongresi ve Sergisi, İzmir, 413-420 (2011). 5. Bilge, D., Bilge, M., (1999), “İndirek/Direk evaporatif soğutma sistemleri kombinasyonu”, IV. Ulusal Tesisat

Mühendisliği Kongresi ve Sergisi, İzmir, 197-204 (1999). 6. Atikol, U., Hacışevki, H., (2001), “Lefkoşa bölgesi için evaporatif soğutma fizibilite çalışması”, V. Ulusal Tesisat

Mühendisliği Kongresi ve Sergisi, İzmir, 421-425 (2001). 7. Şen, E., Akdemir, Ö., Ülgen, K., “İzmir ili evaporatif soğutma sistemlerinin performans değerlendirmesi”, X.

Ulusal Tesisat Mühendisliği Kongresi, İzmir, 1359-1368 (2011). 8. Idrissi M.Y., Tejeda, H.M., Fournaison, L., Guilpart J., “Numerical model of sprayed air cooled condenser coupled

to refrigerating system”, Energy Conversion and Management 48, 1943–1951 (2007). 9. Hajidavalloo, E., Eghtedari, H., “Performance improvement of air-cooled refrigeration system by using

evaporatively cooled air condenser”, International journal of refrigeration 33, 982–988 (2010). 10. Hu, S.S., Huang, B.J., “Study of a high efficiency residential split water-cooled air conditioner”, Applied Thermal

Engineering 25, 1599–1613 (2005). 11. Yanga,J., Chana, K.T., Wub, X., Yuc, F.W., Yangb, X., “An analysis on the energy efficiency of air-cooled

chillers with water mist system”, Energy and Buildings, 12 (2012). 12. Çengel, Y. A. and Boles, M.A., “Thermodynamics: An Engineering Approach”, 5th edition, Mc and

GrawHill,New York, 1-55 (2006). 13. Ertunc H. M. and Hosoz M. “Artificial neural network analysis of a refrigeration system with an evaporative

condenser”, Applied Thermal Engineering, 26 (5-6): 627–635 (2006).

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[Abstract:0103][Heating, Climatization and Air-conditioning Applications in Buildings] SELECTION OF GRILLERS AND DIFFUSERS FOR COMFORT APPLICATIONS

Hüseyin BULGURCU1 and Bekir CANSEVDİ2

1Balikesir University Architectural and Engineering Faculty Mechanical Engineering Dep. BALIKESİR 2Üntes Heating Air Conditioning Refrigeration and Ventilation Company Kazan-ANKARA

Corresponding e-mail: hbulgurcu1962@gmail.com SUMMARY Thermal comfort is depends on some parameters such as indoor air temperature, humidity, air speed and turbulence, the temperature of the surrounding surfaces. HVAC system can be designed with proper capacities and air changes but still result in discomfort due to air greater than 0.25 m/s entering the occupied zone. There are a number of influences that could affect the choice of a supply register of diffuser. Some of the input criteria for making a selection are: Flow rate, face velocity, throw, and pressure loss. In this study, are described the terms of indoor air supply and ambient air distribution models, the selection criteria of the output device. In this last chapter are given some grilles and diffusers selection examples. Keywords: Grille selection, diffuser selection, comforts applications. 1. INTROUDUCTION Air conditioning systems consist an air handler unit, air distribution systems, a heating boiler and a chiller plant. Even if the other equipments are designed very well but air distribution systems play an important role for the success of the system. In conversations with design engineers, they estimate spending weeks and months laying out chillers, air handlers, controls and ductwork, but estimate they spend only minutes selecting and laying out air-distribution devices. Most of the HVAC consulting engineer’s time is spent designing a system to meet the required ventilation rate and space temperature, but occupant discomfort can still occur if the space air velocity is too high. Especially when by examining the HVAC systems in our country, it is noted that ait outlets aren’t well designed in many projects. These problems are thought to be due to the following reasons: lack of experience and knowledge, lack of adequate literature in this area. Grilles and diffusers is the latest element of a ventilation system and are located in the space. It is generally expected from outlets [1]:

to provide the required air flow to ensure the diffusing of the air in the room to create a disturbing air drafts the supply air don’t vent directly to return grilles should not make noise It is suitable for architectural design

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Terms and Definitions for Air Distribution [2] Throw or Radius of Diffusion: Forward travel of a jet to the point where the maximum velocity has decayed to a

nominated terminal velocity. (Vt) Terminal Velocity: Decaying velocity at centre of the jet, which in used to define the throw (Typically Vt =

0.15 to 0.5 m/s). Drop or Rise: Vertical distance between the jet centreline and supply outlet centreline at a nominated throw. Envelope: Jet area within the boundary of a nominated air velocity. Expansion or Spread: Normal divergence of a jet as it leaves an outlet and entrains surrounding air. Free Jet: A supply jet which is able to entrain surrounding air on all sides. Confined Jet or Ceiling Effect Jet: A supply jet which is located so close to one or more surfaces that

entrainment is reduced or eliminated. Ceiling Effect: The tendency of an air stream that is discharged close to and parallel to a surface, to cling to the

surface. This is also called the Coanda effect. Occupied Zone: Defined as the area up to 1.8m from the floor and as close as 150mm from any room surface. (Vr) Room Air Velocity: Average air velocity recorded within the occupied zone. (t) Temperature Differential: Difference between supply temperature and room air temperature. Isovel: A contour of equal velocity.

Figure 1. Air distribution terms and definitions [2]

2. INDOOR AIR DISTRIBUTION MODELS A well-designed room air diffusion scheme ensures that when conditioned air is supplied into a room, it causes no discomfort to the occupants. With a conventional diffusion arrangement, primary air is supplied over the occupied zone where it entrains and mixes with room or secondary air. This process results in a decay of the initial temperature and velocity difference between the supply and room air so that when the supply jet reaches the occupied zone, the velocity and temperature are close to room conditions.

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The location, type and size of the air terminal device will determine the manner in which the supply jet and resultant room air motion behave. With full air conditioning schemes, the change in supply air temperature from a cooling to a heating cycle will also modify the jet trajectory and room air movement pattern. 2.1 Mixed-Air System Concepts (Fig. 2) [3]

Supply air 13°-18°C Cold air supplied outside the occupied zone, thoroughly mixes with room air Creates an air pattern on the ceiling and/or walls Picks up heat and pollutants at the ceiling level Creates low velocity room air motion Ideally creates uniform temperature throughout the space and minimizes stratification

2.2 Fully-Stratified Concepts (Fig. 3) [3]

Supply air 17 - 20°C Cool air supply displaces warm room air at low velocities Uses the natural buoyancy of warm air to provide improved ventilation and comfort Cold air moves slowly across the floor until it reaches a heat source, then rises Improved IAQ

Figure 2 Mixed-Air System Concepts [3] Figure 3 Fully-Stratified Concepts [3]

2.3 Mixed Air-Distribution System One factor that is commonly overlooked in an air-distribution system design is the resulting room air velocities generated by the outlets and the spacing of the outlets. It is fairly common to see outlets spaced too close together and the downward draft that is generated. Also, it is common to see outlets placed too close to an architectural feature that protrudes below the ceiling, resulting in the horizontal jet being forced down into the space and causing drafts for the occupants. Several methods described below can be used to avoid high velocity jets and the resulting drafts when selecting outlets for an air distribution system. These methods can be implemented without adding a significant amount of design time for a project. Designing to control room air motion will lead to higher levels of comfort and a more productive environment [4]. 2.4 Defining a Fully Mixed System A fully mixed system ideally maintains a constant temperature gradient from the floor to the top of the occupied zone. Unlike displacement systems or under floor systems, the fully mixed system maintains a uniform temperature through di-lution of the space air with the air supplied into the space (Figure 4).

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Figure 4. Classification of air-distribution strategies (from 2009 ASHRAE Handbook Fundamentals) [5] Typically, a forced air system is designed to meet the following performance parameters: 23°C to 25°C with humidity at 25% to 60% and air velocity in the occupied zone of less than 0.25 m/s. The latest version of ASHRAE Standard 55-2010 allows for elevated air speeds “to be used to increase the maximum oper-ative temperature for acceptability under certain conditions.” These temperatures and velocities are shown in the standard’s Figure 5.2.3.1. This article covers air outlet selection where the desired temperature range is 23°C to 25°C. 2.5 Effect of Velocity on Comfort To maintain ideal comfort conditions in a space, ambient temperature, air movement, and humidity must meet design criteria. Heat losses in winter and heat gains in summer must be controlled. Sufficient “conditioned” air must be introduced unobtrusively into the space to mix with room air so that the resulting diluted conditions meet comfort requirement. The usual method is to supply air through circular or linear ceiling diffusers. Wall or floor grilles are often used in residences. Heating or cooling devices located in the space, supply air through specially designed outlet grilles. The art is to supply the air so that there will be no objectionable drafts in the occupied zone. High velocity streams of conditioned air should be supplied outside this zone, in the space one 0.15 m from the walls and above 1.8 m from the floor [6]. Some interesting relationships exist between room air motion and the feeling of occupant comfort. Figure 5 shows the effect of air motion on comfort. The charts show that the feeling of comfort is a function of the local room air velocity, local temperature and ambient temperature. The basic criteria for room air distribution can be obtained from the curves shown in (Figure 5). The chart shows the equivalent feeling of comfort for varying room temperatures and velocities at the neck. The % curves indicate the number of people who would object to the temperature and velocity conditions. The same comfort perceptions are shown in (Figure 6) for the ankle region.

Figure 5. Comfort Chart - Neck Region [6] Figure 6. Comfort Chart - Ankle Region [6]

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If 20% objections or 80% acceptance at the same velocities are allowed between (Figure 1 and Figure 2), the temperature deviation allowed between the ankle and neck levels would be about 2.2°C (less than ASHRAE values of 3°C). Table 1 shows the relationship between local velocities and temperatures on occupant comfort. As an example, at a local velocity of 0.4 m/s, the local temperature can be maintained at 23.88°C to reach an 80% comfort level in the space. The same 80% comfort level can be maintained with local air velocity of 0.075 m/s and a local temperature reduced to 21.66°C.

Table 1. Effect of velocity and temperature on comfort [7]

Local Velocity [m/s] Temperature [°C] (tX˗tC) Percentage Objection 0.4 0.3 0.2

0.075

23.9 23.3 22.7 21.7

0 -0.55 -1.11 -2.22

20% 20% 20% 20%

0.3 0.2 0.15 0.075

23.9 23.9 23.9 23.9

0 0 0

1.66

10% 5%

Neutral Warm

Generally, the acceptable level of comfort for a space is considered to be at the point where 20% or less of the room occupants may object to the room conditions. This would indicate that the given condition is acceptable to 80% of the occupants. 3. AIR OUTLET SELECTION METHODS When designing a fully mixed system for comfort, it is important to define the occupied and unoccupied zone for a space. The occupied zone is the space we live and work in. It is typically the volume from the floor up to a height of 1.8 m and 152 mm to 305 mm from the walls. The occupied zone depends on the specific space and use of that space. We can use the unoccupied zone to deliver all the high velocity supply air and do all the mixing. The tool that can be used by a designer to predict the location of a discharge jet is listed as “throw” in a manufacturer’s catalogue. The current standard for outlet testing is ASHRAE Standard 70-2006 (RA 2011), Method of Testing the Performance of Air Outlets and Air Inlets, which defines how throw data is obtained and allows the testing of both isothermal and non-isothermal air [8]. Most of the manufacturer’s throw data is, in fact, isothermal air. The reason for this is the data is easier to obtain and is repeatable. Non-isothermal air requires a balanced room, which does require a significant amount of time to conduct testing. 3.1 Discharge Jets Outlets generate a discharge jet. The designer should select the type of outlet and space the outlets so that the maxi-mum velocity in the occupied space does not exceed 0.25 m/s. There are several principles of jet behaviour that if used along with the manufacturer’s catalogue throw data, the designer can better predict the resulting velocities in the occupied zone. These principles include:

A free jet expands at approximately 22°. The expansion of the jet is due to the induction of air around the jet. If a jet is directed toward a surface, the jet will want to stay on that surface regardless if the surface is a

ceiling, wall or floor. At some point, a cooled jet projected along a horizontal ceiling service will separate from the surface and

drop downwards. The point where the jet leaves the ceiling is defined as the separation point.

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Figure 7. Zone of expansion of isothermal jet [6]

In a mixed air system, the outlet is an “engine” that produces a discharge jet into the unoccupied zone causing air in the occupied zone to be induced into the supply jet. A typical diffuser induces 20 to 30 times the amount of supply air discharged. This rate of induction is how we maintain comfort in a mixed air system. By discharging jets of air into the unoccupied zone, we create air motion and circulation in the occupied zone. Because of the high rate of induction generated by outlets, the temperature of the supply jet reaches near space temperature within a very short distance from the outlet. Most of the exchange of load from the supply air to the room air occurs within several decimetres from the outlet. The outlets should be spaced to ensure that the entire occupied zone has some induction and air motion to avoid stagnant areas. Placing returns in stagnant areas will not create air motion to pull room air to the return.

Figure 8. Extract grille and two way diffuser [9]

3.2 Selection by Noise Criteria The most common method of outlet selection is by using noise criteria (NC) for a device at the required flow rate. The designer determines the outlet type and size by using the required flow rate for an outlet and the maximum NC levels desired for a space. The NC targeted values most commonly used are from Table 1 of the 2011 ASHRAE Handbook HVAC Applications, Chapter 48, Noise and Vibration Control (examples of recommended NC values from the Handbook are shown in Table 2). Table 2. Examples of recommended NC values (2011 ASHRAE Handbook HVAC Applications) [5]

Room Types NC/RC dBA dBC Rooms with Intrusion from Outdoor Noise Sources

Traffic Noise N/A 45 70 Aircraft Flyovers N/A 45 70

Residences, Apartments, Condominiums

Living Areas 30 35 60 Bathrooms, Kitchens, Utility Rooms 35 40 60

Hotels/Motels

Individual Rooms or Suites 30 35 60 Meeting/Banquet Rooms 30 35 60 Corridors and Lobbies 40 45 65 Service/Support Areas 40 45 65

Office Buildings

Executive and Private Offices 30 35 60 Conference Rooms 30 35 60 Teleconference Rooms 25 30 55

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Open-Plan Offices 40 45 65 Corridors and Lobbies 40 45 65

Although this method will select devices that meet required sound levels, it does not address the primary requirements of maximizing thermal comfort in the occupied zone. Using this method, the diffusers can be spaced too close together creating a draft with a velocity of more than 0.25 m/s in the occupied zone. This method can also lead to stagnant zones where there is little or no induction from the space into the supply outlet. 3.3 Selection Using Throw Data Using the published throw data obtained from a manufacturer’s catalogue is one method of outlet selection that can be used to predict room air velocity. This information will indicate the performance of the discharge jet once it leaves the outlet and can be used to fairly accurately predict resulting space velocities (Figures 9).

Figure 9. Plan view of ceiling-mounted outlet indicating location of 0.75, 0.50 and 0.25 m/s isovels (from ASHRAE Standard 70-2006) [5] Throw data obtained per ASHRAE Standard 70-2006 is typically given in a manufacturer’s catalogue at 0.25, 0.50 and 0.75 m/s (Figure 9). Throw is defined as the distance, in m, from the centre of the outlet perpendicular to a point in the mixed airstream where the velocity has been reduced to a specified terminal velocity. In most cases, throw data is based on isothermal air. If either cooled or heated air is shown, the throw data does not indicate the drop or rise of the jet or the spread of the jet. Using the throw information, a designer can map out the location of outlets so as to predict and maintain velocity in the occupied zone below 0.25 m/s. 3.4 Selection by T50/L and ADPI Another available method to predict air velocities and comfort in the occupied zone is by using the T50/L ratio to predict the resulting ADPI value for a space where T50 is the catalogued throw data to 50 fpm (0.25 m/s) and L is the characteristic length of the space being evaluated. Characteristic Length L is the horizontal distant from the outlet to the outside of the zone the outlet serves. If the outlet is placed so that there are differing distances from the outlet out to the zones served (i.e. a four way pattern), the outlet may have up to four different L values that are used to calculate the T50/L ratio in each direction. Characteristic Length L is defined in Table 3 in Chapter 57 of the 2011 ASHRAE Handbook—HVAC Applications. 3.5 What is ADPI

Statistically relates local temperatures and velocities to occupant comfort Ratio of diffuser T50 to characteristic length of the room being served ADPI > 80 is acceptable Currently only applies to cooling applications Soon may be expanded to include more diffuser types and add heating applications

Air diffusion performance index (ADPI) is a single number that quantifies the overall comfort of a space when in cooling. ADPI is the percentage of points in a space where the effective draft temperature is between –19°C and –17°C and the air velocity is less than 0.36 m/s. A high percentage of people have been found to be comfortable in cooling applications for

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office type occupations where these conditions are met. High ADPI values generally correlate to high space thermal comfort levels with the maximum obtainable value of 100. The effective draft temperature provides a quantifiable indication of comfort at a discrete point in a space by combining the physiological effects of air temperature and air motion on a human body. A point is considered comfortable if the results of Equation 1 are between –19°C and –17°C, and the measured velocity at the point is less than 0.36 m/s.

Ted=(Tx – Tc) – 0.07(Vx – 30) (1)

where Ted = effective draft temperature, °F Tx = local airstream dry-bulb temperature, °F Tc = average (control) room dry-bulb temperature, °F Vx = local airstream centreline velocity, fpm By determining the value of T50 from a manufacturer’s catalogue, and measuring the characteristic length L from the proj-ect’s plans, the ratio can be determined and the predicted ADPI value can be estimated from Table 2.

Table 3. Air diffusion performance index (ADPI) selection guide [4]

3.6 Location of Return/Exhaust Grilles ASHRAE Handbook—HVAC Applications, Chapter 57, Room Air Distribution, recommends that “A return inlet affects room air motion only in its immediate vicinity. The intake should be located in the stagnant zone to return the warmest room air during cooling or the coolest room air during heating. The importance of the location depends on the relative size of the stagnant zone, which depends on the type of outlet.” Also, the return or exhaust grille does not short circuit the supply jet. In fact, if a supply jet is at or above 0.76 m/s and directed over a return, the jet will either pull air out to the return or exhaust or create a “blanking effect” over the return or exhaust where no air is taken out of the space. If the return or exhaust is placed at the end of the supply jet where the velocity is 0.76 m/s or less, the air being returned is at or near room temperature and the amount of air being returned is close to 5% of the total air motion in the space. 4. AIR TERMINAL DEVICE SELECTION Frequently the type and location of grille or diffuser will be determined by architectural or other requirements. If this is the case, performance data can be applied directly to determine whether the resulting performance is acceptable. If an entirely free choice is available, refer to Sections 2 and 5 where comparative selection data can be used to determine the most suitable air terminal device. Usually, the sizing of a particular terminal device is based on the throw, but at each stage, it is necessary to check that any acoustic or pressure loss specification is satisfied. Having decided which type of grille or diffuser is required, apply the following techniques for selection in conjunction with necessary information, such as total airflow rate and room size, it is helpful to have scale drawings of air terminal device layouts.

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4.1 Linear slot diffusers These diffusers can be selected or set to provide horizontal diffusion in one or two directions across a flat ceiling surface drawing up on the ceiling effect. As the supply jet entrains room air, it expands in the vertical plane and must be prevented from prematurely entering the occupied zone. Using the following table, determine the maximum throw according to the ceiling height:

For continuous slot diffuser arrangements, divide the ceiling area into convenient strips, based on the maximum throw.

Determine the available active lengths of diffuser sections. Calculate the diffuser duty by dividing the active length into the total airflow rate to be supplied. With the available information of maximum throw and diffuser duty, draw two lines on the selection nomogram;

one passing through the minimum radius of diffusion and the other passing through the maximum radius of diffusion. This produces a band of possible selections.

It is now necessary to find the optimum selection, which is usually a compromise between economy (minimum number of slots) and comfort (maximum number of slots to produce the ideal room air movement for the particular application).

If the optimum selection falls below a one-slot diffuser, then the active length can be reduced as necessary. If the optimum selection is greater than eight slots, it is possible that a slot diffuser arrangement is not practical

Figure 10. Maximum and minimum throw distances

4.2 Circular, Square and Rectangular Diffusers Circular diffusers produce a radial air diffusion pattern while square and rectangular devices can be selected or adjusted to produce 4, 3, 2 or 1 way directional air patterns and drawing up on the ceiling effect. Wherever possible, select a 4 way or radial pattern as this results in the most efficient air diffusion. Using the following table, determine the maximum throw or radius of diffusion based on the zone ceiling height. This will prevent the supply jet from entering the occupied zone prematurely, as it expands in the vertical plane.

Using a scaled ceiling plan, divide the area into convenient squares twice the sizes of the derived maximum throw. A circular or square diffuser at the centre of each area can now be selected to handle its proportion of the total airflow rate.

Using selection tables or nomograms, determine the diffuser sizes which satisfy the throw parameter. The most economical selection will produce a minimum radius of diffusion very close to the required throw.

However, the optimum selection will probably be a compromise between the most economical selection and that which will produce the most comfortable room air movement.

If the maximum radius of diffusion produced by the smallest available diffuser is less than the required throw, then insufficient room air movement and high level stagnation will result. An alternative air terminal device should be considered.

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Similarly, if the minimum radius of diffusion produced by the largest available diffuser is greater than the required throw, it is probable that the particular air terminal device is unsuitable.

Wherever possible, diffuser selections should be within the limits given in the tables; extrapolating data down to very low-neck velocities will usually result in poor air diffusion, for example; high-level stagnation with heating cycles, and draughts due to dumping with cooling cycles

Figure 11. Maximum and minimum throw distances on side view

4.3 Examples Using Throw Data The following examples show how the throw data can be used to predict the resulting air motion in a space. In this example, the space is 6 m × 6 m with 3 m ceilings and the required airflow is 448 L/s.

Figure 7 (left): Diffuser Performance. Two louver-faced 381 mm × 381 mm inlets. Each diffuser is 221 L/s, totalling 442 L/s. Throw data is 18 (0.25 m/s) and NC<20. Use the T50/L ratio to obtain ADPI range for 90%: 1.4 to 2.7 and T50/L calculation 18/5 = 3.6. Figure 8 (right): Diffuser Performance. One louver-faced 533 × 533 mm inlets with four-way discharge. Each diffuser is 434 L/s. Throw data is 24 (0.25 m/s) and NC<20. Use the T50/L ratio to obtain ADPI range for 90%: 1.4 to 2.7 and T50/L calculation 24/10 = 2.4. Example 1 In this example (Figure 7), colliding airstreams will generate a high velocity in the occupied zone. The T50/L calculation shows a value of 3.6, higher than the desired 1.4 to 2.7, indicating a draft in the occupied zone. We also can predict too high a velocity in the center of the space by mapping out the throws of the outlets. Example 2 To reduce the room air velocity, one diffuser is selected and placed in the center of the room (Figure 8). The value of T50 /L is now 2.4 and within the recommend range to maximize the ADPI value for the space. Although the supply jet does reach the floor and enters the occupied zone, from Figure 2 we know that the ankle region is much more tolerant to higher velocities compared to the neck region. This selection has eliminated one diffuser compared to Example 1 and the associated costs while providing a higher level of comfort.

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5. CONCLUSION Most HVAC designers spend a significant amount of time selecting chillers, air handlers, pumps and cooling towers with the purpose of maintaining occupant thermal comfort in a space. Most conventional air distribution systems in North America are mixed air systems where space comfort is maintained through the induction of air from the occupied zone into the unoccupied zone. The definition of comfort includes maintaining room air velocities below 40 fpm to 50 fpm (0.20 m/s to 0.25 m/s). This is sometimes overlooked by designers, resulting in discomfort to the space occupants caused by ex-cessive air motion. By using some simple selection techniques such as mapping cataloged throw data and the T50/L method to predict ADPI, designers can select and space outlets to maximize the comfort level in an occupied space. Using these methods also offers the opportunity to reduce the required number of outlets and the associated costs. REFERENCES

1. Bulgurcu, H., Ventilation Systems, Chamber of Mechanical Engineering Publication no: 650, İstanbul-December 2015.

2. Waterloo The Green Book, http://www.waterloo.co.uk/wp-content/uploads/2012/07/The-green-book-technical.jpg (Retrieved on 20.12.2015).

3. Zimmerman, R., Room Air Distribution, www.cenpenn.ashraechapters.org/.../ashrae%2 (Retrieved on 15.01.2016) 4. Int-Hout, D., Miler, K., Cenci, P., Basic of Room Air Distribution and ADPI,

http://www.ashraebistate.org/sites/all/files/events/Basics_of_Room_Air_Distribution_%26_ADPI_0.pdf (Retrieved on 20.01.2016)

5. David A. John, Selecting Air Distribution Outlets Designing for Comfort, ASHRAE Journal, September 2011. 6. Engineering Guide, Nailor Indutries Inc. http://www.nailor.com/onlineCatalog09/CAT-06/CATGENIN.pdf

(Retrieved on 20.01.2016) 7. Titus Grilles and Diffusers-Engineering Guidelines, https://www.titus-

hvac.com/file/7561/diffusers%20eng_guidelines2013.pdf (Retrieved on 18.01.2016) 8. ASHRAE Standard 70-2006 (RA 2011) -Method of Testing the Performance of Air Outlets and Air Inlets (ANSI

approved) 9. Int-Hout, Daniel, Best Practices for Selection Diffusers, ASHRAE Journal, June 2004. 10. Internet http://www.arca53.dsl.pipex.com/index_files/grille3.htm (Retrieved on 19.01.2016)

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[Abstract:0106][Economy of Energy and Environment] THE OBSERVATION OF TARGETS ACHIEVED DURING DISTRICT HEATING

DEVELOPMENT IN RIGA CITY

E. Dzelzītis1 and N. Talcis2. 1Riga Technical University 2Riga Technical University

Corresponding email: egils@lafipa.lv

SUMMARY Te vajag summary ar 150 vārdiem!

INTRODUCTION A modern district heating company is a power utility whose operations are based on the use of cogeneration or even trigeneration technologies and which provides district heating services. Improvement of efficiency of district heating systems plays the decisive role in securing the company competitiveness which, in turn, is a precondition for utilising the advantages of the district heating system in power generation, in minimisation of the environmental impact that finally improves the quality of life and comfort of heat consumers and presents the hidden reserve of the economic development related with re-distribution of free financial resources and introduction of modern production technologies. JSC „RĪGAS SILTUMS” produces heat at 5 heat plants with the capacity ranging from 50 to 405 MW, two of which operate in a cogeneration mode, and at 38 automated gas fired boiler houses with the capacity ranging from 0.05 to 20 MW, two of which operate in a cogeneration mode. The total installed heat capacity of JSC „RĪGAS SILTUMS” equals 1082 МW. The total length of heat networks in Riga is more than 800 km, of which JSC “RĪGAS SILTUMS” owns 684 km and more than 116 km belong to individual companies, cooperative societies of apartment houses, state institutions, etc.

METHODS Joint Stock Company “RĪGAS SILTUMS” which is the biggest district heating company in Latvia and also in the Baltic countries is the operator of the district heating system in Riga. JSC “RĪGAS SILTUMS” was founded in 1996 by merging the Heat Supply unit of the state-owned JSC “Latvenergo” and municipal companies operating heat supply systems. At the beginning stage of its operation JSC “RĪGAS SILTUMS” had 6 heat plants and 111 non-automated boiler houses, i.e. 18 gas-fired and 93 coal-fired boiler houses. Approximately 50% of the heat consumers connected to the district heating system in Riga received heat via central heat sub-stations operating on the basis of a four-pipe system. There were totally 185 central heat sub-stations in the district heating system of Riga. The length of heat networks equalled approximately 1000 km, including 135 km of networks for supply of hot water. In order to secure the required amount of supply of heat at the appropriate quality level to heat consumers, improve safety and energy efficiency of heat supply, minimise heat losses in pipelines and thus reduce harmful emissions to the environment, the implementation of the project for renovation of the district heating system in Riga was launched in 1997. The project for renovation of the district heating system in Riga was implemented in compliance with the requirements of the Energy Law and on the basis of the concept of heat supply development in Riga. In the course of the implementation of the project for renovation of the district heating system in Riga priorities were set and it was decided to implement the modernisation of the district heating system in two stages. Taking into account that the biggest heat losses were found at central heat substations and in the four-pipe system, the elimination of central heat substations and reconstruction of individual heat substations was performed at the first stage. During the second stage heat networks and heat sources were upgraded. This program included the following measures: a) replacement and reconstruction of the existing heat networks in bad condition; b) closing of small and medium inefficient boiler houses and connection of consumers to the existing heat networks or construction of automated boiler houses; reconstruction of large heat sources and installation of cogeneration units at heat sources.

Elimination of central heat substations The program of elimination of central heat substations included the following: a) installation of individual heat substations in houses which used to receive heat via central heat substations; b) transition from a four-pipe system to a double pipe system and elimination of hot water networks; c) reconstruction of individual heat substations in buildings connected to the heat networks. In the result, all 185 central heat substations were eliminated until 2001, 3008 new modern automated individual heat

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substations were installed, as well as 135 km of the networks of hot water supply were eliminated. 8138 out of totally 8147 individual heat substations were upgraded during the period until 2015 (see Figure 1).

Figure1. The number of constructed automated individual heat substations Construction of individual heat substations will allow adjusting the temperature in premises and setting the required hot water temperature. The equipment installed at individual heat substations provides the following: a) the possibility of maintaining comfortable heat supply to a building in compliance with customers’ demand; b) the necessary temperature in the building irrespective of the incoming temperature to the individual heat substation; c) automation of the heat supply process to the building; d) the possibility of rational use of heat preventing inefficient consumption; e) the possibility of accurate adjustment of hydraulic regimes of the internal heat supply system in the building; f) stable hydraulic regime of the system of the building irrespective of the hydraulic regime of heat networks; g) minimum water leakage in cases of damage to the heat system.

Upgrading of heat networks The security and continuousness of heat supply in Riga depends to a large extent of the technical condition of heat networks. In this regard, repairs and reconstruction of the sections of heat networks in a bad condition and where heat losses exceed the standard indices is performed every year. First of all, the sections of heat networks where a bad condition of insulation, internal or external corrosion of pipes was identified and where there were a high number of accidents which may cause interruptions of heat supply to consumers, are replaced. Reconstruction of the sections of heat networks is performed by using non-channel pre-insulated pipelines the operation of which is not impacted by a high level of ground water. Replacement (re-installation) of the sections of main and distribution heat pipelines has considerably reduced the company expenditure related with heat losses, elimination of leakages in pipelines, as well as considerably improved the security of the district heating system in Riga. Development of replacement of heat networks during last ten years (see Figure 2).

Figure 2. Replacement of heat networks during last ten years In order to secure continuous and secure supply of heat to consumers, a four year program (2002 – 2006) was developed for replacement of sleeve joints with expansion joints in all main heat networks and installation of new shut-off fittings in all main chambers. In the course of implementation of this program, 1111 expansion joints with diameter up to Dn1200 were installed and 2164 units of shut-off fittings with diameters up to Dn800 were replaced. Completion of this program provided the possibility: a)to disconnect sections of heat networks in a fast and secure manner; b) to change heat supply regimes, thus improving the efficiency of district heating; c) not to disconnect consumers and to provide heat during summer repair periods and in case of accidents of heat networks. Minimisation of the amount of losses in heat transmission is among the most important directions for the improvement of efficiency of district heating systems. Complete elimination of these losses is not viewed as a feasible possibility and they have to be included in the final costs of the product. Expenditure related with the heat carrier transportation may also reduce the effect of the use of cogeneration. Major measures which allowed reducing heat losses: a) Reconstruction of heat networks by using pre-insulated pipelines according to the non-channel technology; b) Replacement of insulation of above ground heat networks and the heat networks installed in building basements; c) Reconstruction of insulation in heat chambers; d) Installation of modern equipment and shut-off fittings (ball valves, expansion joints); e) Systematic monitoring of leakages and their elimination. f) Systematic analysis of the increase of return temperature in heat networks and at customers and elimination thereof; g) Periodic flushing of drainage systems in heat networks; e) Application of modern technologies for connection of consumers to heat networks - pressure cutting; d) Hydraulic testing of heat networks; c) The system of collection of heat meter readings which was implemented in JSC “RĪGAS SILTUMS” allowed introducing continuous control over the alarm system of heat networks. Fulfilment of the above listed measures allowed JSC “RĪGAS SILTUMS” to reduce the proportional share of heat losses from 19.97% in fiscal year 1996/1997 to 12.43% in fiscal year 2014/2015 (see Figure 3).

Figure 3. Reduction of heat losses in fiscal year 1996/1997 to fiscal year 2014/2015 Minimisation of the level of feeding heat networks is among the main goals in the course of operation of heat networks. In order to identify the places of leakage, district teams prepare plans for surveying main and distribution heat networks and the Dispatch Service analyses the use of feeding into heat networks and causes of any change therein. At present the identified leakages of the heat carrier in heat networks are low and it is difficult to identify them even with the help of the existing devices. Air photography is used as an additional tool for searching for leakages in heat networks. This has presented the possibility of identifying leakages and heat losses in a fast manner. In order to speed up the search for leakages in heat networks, heat carrier is coloured green by using fluorescent for this

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purpose. This allows fast identification and elimination of a defect. The set of the above described measures for searching and eliminating leakages has provided results which are demonstrated by the annual decrease of the amount of feeding. Thus, in 2014/ 2015, in comparison to 1996/ 1997, the total feeding decreased by 189.4 t/h in JSC “RĪGAS SILTUMS”. Change in the annual consumption of feed water per fiscal years (see Figure 4).

Figure 4. Change in the annual consumption of feed water per fiscal years Upgrading of heat sources In the course of implementation of the project of renovation the district heating system in Riga, JSC “RĪGAS SILTUMS” eliminated inefficient small- and medium-scale boiler houses, including the ones fired with fossil fuel. Heat consumers of these boiler houses were connected to boiler houses with high operational efficiency or new automated gas-fired boiler houses were built, in the result of which the specific fuel consumption for production of 1 MWh of heat decreased and the environment pollution was reduced. By utilising the advantages of the district heating system which allow securing the lowest possible harmful emissions to the environment, JSC “RĪGAS SILTUMS” has been gradually reducing the negative environmental impact of heat production sources. Among the most successful projects of the recent years there is the complete reconstruction of a boiler house, during which old steam units were replaced by modern, highly efficient water heating boilers equipped with condensation economisers providing for the rate of useful operation of the boiler house above 100% according to the calculations based on the lower heating value. In 2010, in the course of continuing implementation of modern technologies improving energy efficiency, in heat plant “Imanta” a condensation economiser with the heat capacity 10 MW, which is the biggest one in the Baltics, was installed for the boiler KVGM-100. In the result of its operation the efficiency was considerably increased (up to 102%), and the use of produced condensate allows minimising expenditure (which had been quite high) for purchase of water. In 2010 a unique project was implemented at the heat plant “Imanta” - installation of an industrial absorption heat pump/ cooler which allows efficient utilisation of approximately 2 MW of low potential heat emitted from the cooling system to the atmosphere during the operation of the cogeneration power unit with heat capacity 47 MW. During the heating season when the heating load is sufficiently high, the absorption heat pump allows increasing the operational efficiency rate of the cogeneration power unit by 2%, reducing the amount of harmful emission to the environment, as well as reducing the amount of water used for the needs of the cogeneration cycle. Improvement of energy efficiency, saving of emission allowances, reduction of power consumption, considerable reduction of the volume of harmful emissions, minimising of the rick of icing, reduction of the consumption of cooling water by 48,000 tons per year – this is the list of the major advantages gained by the installation of the industrial heat pump at the cogeneration unit of the heat plant “Imanta”. The heat plant “Vecmilgravis” was commissioned at the end of the 1970-ies and supplies heat to residential buildings and industrial companies of Vecmilgravis district. As no reconstruction of equipment had been performed since the commissioning of the heat plant, it was decided to radically upgrade this heat plant. In 2011 the modernisation of the heat plant, which was implemented in several stages, was completed. During the first stage the obsolete boilers KVGM-100 and PTVM-30 in the existing premises were replaced with modern highly efficient water heating boilers equipped with condensation economisers and with the total heat capacity of about 50 MW; During the second stage, in the course of active introduction of biofuel at its heat production sites, thus contributing to the fuel diversification with the goal of intensifying the use of local renewable energy resources and contributing to the reduction of the emission of carbon dioxide, a completely new boiler house was constructed with wood-chips fired water heating boilers with the total capacity 14 MW; Following the start of operation of the new highly efficient equipment at the heat plant “Vecmilgravis” the total operational efficiency increased from 92-94% to 107-110%. In 2013 other two wood-chips fired plants were commissioned – a cogeneration plant with heat capacity of up to 22 MW and electrical capacity 4 MW and a water heating boiler house with capacity 20 MW, where the fluidised bed combustion technology is implemented allowing combustion of a broad range of biofuel, which allowed increasing the proportional share of heat produced on the basis of renewable energy sources from 4.4% to 19.4%. For the purpose of improving the efficiency of production of heat on the basis of biofuel, flue gas condensers are being installed. Flue gas condensers are used for improving the operational efficiency of boilers by utilising the residual heat of flue gases by reducing their temperature which allows improving the environment and using fuel more efficiently. This allows increasing the efficiency of use of biofuel from 85% to 105% (based on the lower heating value).

Figure 5. Total operational efficiency of heat sources of JSC “RĪGAS SILTUMS” Note: Operation of several heat sources within a single zone when, in compliance with hydraulic regimes or due to technical reasons, relevant units may be put into operation accordingly, can impact the total operational efficiency of heat sources.

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System of collection of heat meter readings Because of continuous fluctuations of the gas price, it was necessary to ensure accurate metering of sold heat to consumers on monthly basis. Therefore, JSC “RĪGAS SILTUMS”, after having analysed technical possibilities, in 2011 introduced a two-way system of collection of consumers’ heat meter readings. The system of collection of consumers’ heat meter readings implemented in JSC “RĪGAS SILTUMS” ensures two-way communication with metering sets of all the consumers connected to heat networks in Riga. Implementation of this large-scale project ensured collection of sufficiently complete information regarding the condition of and the problems in the heat supply system of the city. The whole city (300 sq.km) is serviced by a minimum system infrastructure. The whole system infrastructure consists of 41 base stations of transceivers and 45 retranslators (i.e. just one base station and one multiplicator per 5.5 sq. km of the system coverage area). The system of collection of heat meter readings is multifunctional and also provides functions like reading of an individual heat meter upon request, remote software updating of metering modules, parallel (asynchronous) collection of data from all the meters in the system, immediate transmission of an alarm signal regarding interruptions in the power network, immediate transmission of an alarm signal regarding attacks to the site, identification of accidents in the heat network and transmission of an alarm signal thereof. Among the most important achievements there is the fact that the system for communication with 7500 end units located in the area of more than 300 sq. km requires little resources. The system of meter reading was integrated with the existing software of JSC “RĪGAS SILTUMS” allowing implementation of technological control and recording of various measurements, as well as performing their detailed analysis. Taking into account the technical possibilities of the project, JSC “RĪGAS SILTUMS” collects data on daily basis at 9:00 a.m. This allows all the services to implement control over the consumption of heat by consumers. Control over consumers or consumer groups according to any time mode, as well as the possibility of individual inquiries of data from meters from any computer in the company network or from a mobile telephone has improved the efficiency of work of the heat inspection. The project which was implemented in Riga serves as the confirmation of technical achievements in the area of two wireless communications. The project success is based on the fact that the expanded system has demonstrated security, operationality and a high level of availability and communication. The system has turned out to be very secure also under extreme operational conditions. Successful communication is maintained in more than 99% of all 7500 installed measuring sets. With the help of developed software the employees of JSC “RĪGAS SILTUMS” ensure continuous control over the operation of individual heat substations and, in case of necessity, notify problems to the maintenance organisation.

Installation of allocators

Implementation of the program for renovation of the district heating system in Riga regarding installation of modernised heat substations provided the savings of approximately 8-10% of the total amount of heat consumed by a house. However, there were also buildings where heat bills increased. For example, after installation of heat substations circulation in the system of hot water supply and heating was improved. In the quarters where it used to be cold before, temperature increased to 18-20 degrees, thus the total heat consumption by a house also increased. As heat supply is a very expensive service, JSC “RĪGAS SILTUMS” wants to ensure that customers continue saving heat also in future. We have reached a new level now and offer consumers to save not only on the level of a house, but also individually in their apartments. For this purpose allocators, which should be correctly referred to as distributors of payment for heating, may be installed in apartments and regulators on radiators allowing residents to change heating temperature in a room. The allocator is not a heat meter, it is the distributor of the payment for heat. For example, a water meter records how many cubic meters of water were received per month. The allocator cannot meter how many kilowatts were used for heating the apartment, however these devices record what proportional share of the total heat consumption of the house was used in the particular apartment. Thus, also heat bills in different apartments will differ. This principle was invented 99 years ago in Switzerland, and in 1981 the mandatory installation of allocators in all the apartment houses in Germany was commenced, as the energy crisis started there at that time. According to the estimations by German experts, thanks to adjustment and individual metering of consumed heat it is possible to achieve savings of up to 20% and in some houses even more. In Latvia the use of heat payment distributors started in the beginning of the 2000-ies. The efficiency of allocators was confirmed by our pilot project implemented in a municipal house. In autumn 2014 we launched a project – research regarding the impact of allocators upon heat consumption in individual apartments of a municipal house. In the beginning we proposed this idea to residents in regular privatised houses, offered to test this system in their apartments, however in such houses it is difficult to come together for taking a decision. Due to this reason a new municipal house was selected for this pilot project (see Drawing 1). We installed allocators on all the radiators in 168 apartments of this house. It is important to understand that in this house there was a modern two-pipe heating system, there were regulators on radiators allowing each resident to adjust heating temperature in rooms. However, earlier the residents of the building were not interested in changing the adjustments of

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radiators because their heating bills were the same as those of their neighbours who were not saving heat. After the installation of allocators everything changed. People understood the main principle: one pays for what was actually consumed. By implementing the pilot project we offered a full set of services, in particular, we installed allocators and covered the related costs, we were recording their readings and issued heat bills accordingly. For us it was important to understand for ourselves how the process of individual metering of heat should be correctly organised, in order to be able to offer the same service also in other houses in Riga. We found a manufacturer of allocators whose devices not only allow to record readings of the proportional share of heat consumed in the particular room once a month, but also record the mean temperature in this room. It seemed important for us in order to see how the quality of life changed after people started adjusting their radiators and saving money paid for heat. Now, when somebody is complaining that heat bills are too high, we can check the data about this apartment and explain why the person pays for heating more than the neighbour. Now once a month an employee of JSC “RĪGAS SILTUMS” remotely reads the data of all the allocators by using a mobile device. Then residents receive accurate heat bills for the share of heat consumed by each apartment. Prior to launching the pilot project, the experts of JSC “RĪGAS SILTUMS” very carefully studied methodologies of assessing the payment for heat by using allocators applied both in Latvia and abroad. We also heard complaints from people that after the installation of allocators some of them have to pay for heat more than before. It turned out that the problem is not caused by the allocators, but by the assessment methodology used in each equipped house, as everything depends on the proportional share! Payment for heat in such houses consists of two items, in particular: the payment for the proportional share of the total heat consumption of the house consumed by the particular apartment, and the payment for heating of the premises of common use, which amount is distributed among apartments, as well as heat provided by the heating system pipelines. Based on a decision of the general meeting of apartment owners, any proportions of the two above components may be applied. For example, residents may decide that 70% of heat will be accounted for by using allocators and 30% will be the fixed payment depending on the area of an apartment. Under the system of individual metering, by adjusting radiator settings residents can only impact the first component of the payment. The proportion of 70/30 was borrowed in Latvia from the Western Europe and it was incorrectly accepted. First, because in Europe settlement is done once a year based on the mean annual consumption, but in Latvia this is done once a month. Second, in Latvia there is a high number of not reconstructed houses where the heating systems are in a bad condition. It is clear that the heat distributed among apartments, heat supplied from the heating system pipelines and heat consumption in premises of common use account for a much higher proportional share than 30% of the total payment for heat. If the proportion of 70/30 is applied in such a house without further consideration, the residents who switch off their radiators completely for the sake of saving are winners. They will only pay for the “common heat” which is quite cheap according to this proportion. However, the residents who switch on radiators at least time from time will pay the biggest part of the total heat bill of the house. This situation can be easily corrected by introducing correct proportions for payment for heat in the house. We at JSC “RĪGAS SILTUMS” chose the mean outdoor temperature as the starting point for the methodology. For example, if the mean outdoor temperature during this month was-10, the proportion 70/30 is applied in the house for calculation the payment for heat by each apartment. If the mean temperature is 0, the applied proportion will be 50/50 and if the temperature is +10 we consider that the correct proportion would be 30/70. On our scale there are several steps, which means that any outdoor temperature fluctuations impact the calculation of heat bills in the relevant month. Many service organisations have independently arrived to the same conclusion and are also changing the payment proportions depending on, for example, the season of the year. However, in houses where the administrator and apartment owners apply the same proportion all the time there are huge problems! Moreover, in the course of calculation of the payment for heating we also apply the coefficient of the location of each apartment. We understand that if an apartment is located on the top floor or in the corner of the house, its owner will have to use more heat for heating the rooms than his/ her neighbours located in the central part of the house. Historically, people were privatising apartments and did not have much choice. Due to this reason the application of such coefficient of location is just fair. These coefficients for each house are defined by experts on the basis of the basic estimation of heat losses. Recently the Cabinet Regulation has entered into force imposing an obligation upon energy auditors to estimate these coefficients and to provide recommendations regarding their application.

Drawing 1. Implementation of the pilot project of installation of allocators in a new municipal building.

Figure 6. Comparison of heat consumption prior to and after implementation of the pilot project of installation of allocators in a new municipal building. The comparison of consumption of thermal energy before and after the pilot project implementation on installation of allocators to new municipal building shows that in comparison to heating season year 2013 /2014 the residents of building, where the installation of allocators took place, saved approximately 20% of thermal energy in heating season year 2014 /2015. In the same time the residents of similar new buildings, where allocators were not installed, consumed on the average 2% more heat in heating season year 2014 /2015 if compare with a year earlier (see Figure 6).

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DISCUSSION

The assessment of the policy of the development of district heating system in Riga during the operation of the company leads to a justified conclusion that it has been successful. This is attested by the fact that the heat price in Riga is lowest among the capitals of the Baltic countries and even among the biggest cities in Latvia. JSC “RĪGAS SILTUMS” has managed to achieve this thanks to the invested effort for minimising heat losses. At present heat losses amount to less than 13% of the heat supplied to consumers, and in 1996/1997 heat losses amounted to approximately 20%. In 2013, thanks to the launch of operation of biofuel-fired heat sources where wood chips which are cheaper than natural gas are used, the heat price was reduced by 3%. Thinking about future plans, JSC “RĪGAS SILTUMS” is actively working towards maximum reduction of operational costs, which will allow offering even lower prices, at the same time providing high quality and security of heat supply, efficient use of natural resources, maximum reduction of the environmental impact. This means that long-term solutions for improving energy efficiency are required, including the use of the state-of-art technologies, minimisation of dependence on imported energy resources and broader use of local bio resources which is not only a more profitable solution for the company, but also contributes to the economic development. REFERENCES

REF1 REF2 REF3

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[Abstract:0107][Heating, Climatization and Air-conditioning Applications in Buildings] NEXT GENERATION BUILDING AUTOMATION TECHNOLOGIES

Jagdish Naik

Honeywell GmbH, Cooling Solutions SUMMARY Technology is influencing and reshaping the life cycles of Buildings Industry. The design strategies ( usage of Solar energy from light as well as power perspective ) , the design tools such as use of Building information Modelling during the design and maintainance Cycle , The enterprise wide integrations with the “scope of enterprise” itself widening to integrate larger business operations, the technology tools to maintain and operate the buildings with a goal to enhance the lives of equipment apart from enhancing the productivity of operations are some of the examples we can think of . Technology is impacting the design cycle times, the productivity efficiencies, safety in maintenance cycles and processes. They bring benefits to Owners, Facility mangers, Occupants and Visitors. Automation systems are no more expected to limit centralizing the intelligence in one location (control room) as the past-generation systems did. The control room intelligence in fact is in fact being redistributed back to the various stakeholders wherever they are located . The client Server architectures evolved out of the reality of computing hardware being excessively expensive to buy, install and run. The computing hardware being a commodity in today’s world the computing architectures too are influenced by it and have changed substantially. The Automation landscapes have changed. Honeywell has always been in frontline to adopt to the changes and stay ahead of the Global technology changes. This discussion is to introduce you to how Honeywell is leading the way through their ongoing improved product lines and what they intend to bring in near future . This will cover Building Information Modelling, Mobility , Total Work flow Management on Mobile, discussions on Analytics.