Thesis Evaporative Cooling Towers

THESIS FOR THE DEGREE OF DOCTOR OF PHILOSOPHY

Evaporative Cooling Tower and Chilled Beams

Design Aspects for Cooling in Office Buildings in Northern Europe

BENGT BERGSTEN

Building Services Engineering

Department of Energy and Environment

CHALMERS UNIVERSITY OF TECHNOLOGY

Göteborg, Sweden 2009

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Evaporative Cooling Tower and Chilled Beams Design Aspects for Cooling in Office Buildings in Northern Europe

Bengt Bergsten

ISBN : 978-91-7385-348-4 © Bengt Bergsten Doktorsavhandlingar vid Chalmers tekniska högskola Ny serie nr 3029

ISSN 0346-718X Document D2009:05 Building Services Engineering Department of Energy and Environment Chalmers University of Technology SE-412 96 Göteborg, Sweden 2009 Telephone: + 46 (0)31-772 10 00 http://www.chalmers.se Printed by Chalmers Reproservice, Göteborg, Sweden 2009

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Evaporative Cooling Tower and Chilled Beams Design Aspects for Cooling in Office Buildings in Northern Europe

Bengt Bergsten

Building Services Engineering Department of Energy and Environment Chalmers University of Technology

Abstract

The purpose of this thesis is to examine a comfort cooling system consisting of a hydronic cooling system with an evaporative cooling tower and chilled beams. The system has no conventional chiller. The analysis of the comfort cooling system is made through simulations in a building simulation tool, IDA Indoor Climate and Energy (IDA ICE) and monitoring of a pilot plant equipped with this system. A mathematical model of a cooling tower is developed and validated in this project and added to IDA ICE. The base case condition comprises a normal office building with a normally sized cooling system. The total internal heat gain, including solar radiation, is between 50 – 70 W/m2 and the climate conditions is equal to those in northern Europe, i.e. north of latitude 48 – 49°N. The pilot plant consists of an office space of 450 m2 which is cooled by a free cooling system. The free cooling system has an outdoor evaporative cooler which is connected to a chilled beam system in the office building. The monitoring was made during May – August 2007. The results form the simulations indicate that the comfort cooling system can maintain a thermal climate where the annual maximum indoor air temperature is between 24 – 26°C in a Nordic climate, and between 25 – 27°C in the rest of the northern Europe. The annual duration of indoor air temperature during working hours exceeding 24°C is between 1 – 5% in Nordic climates and between 3 – 8% in the rest of the northern Europe. The annual COP of the cooling tower and the cooling system is about 7 at base case conditions. Thus, the use of electric energy is about a third of the energy used in a conventional cooling system. The outcome from the pilot plant basically confirms the results regarding the indoor climate. However, the COP of the evaporative cooler during the measured period is lower compared to the results from simulations and findings from literature. A hydronic cooling system with chilled ceilings or beams and an evaporative cooling tower can be applied to both new and refurbished buildings. This comfort cooling system represents well-established techniques and no parts of the system are new or unproven on the market. Preliminary data from other sources indicate approximately equal total investment costs for a hydronic cooling system with a cooling tower compared to a conventional hydronic cooling system with a mechanical chiller. Keywords: free cooling, low energy cooling, evaporative cooling, commercial buildings, cooling tower, cooling tower model, simulation, hydronic cooling system, chilled beams, pilot plant.

IV

This doctoral thesis has been financed by the Swedish National Energy Agency, BELOK, Swedish Building Research Programme Competitive Building, IMI Indoor Climate and CIT Energy Management

V

Preface

The idea behind this research project originates from the early 90’s when working as an HVAC engineer, designing HVAC systems in buildings. The question came up when learning about evaporative cooling of air; is it possible to chill a hydronic cooling system (with e.g. chilled beams which is very common in Sweden) by evaporative cooling only and still get an acceptable indoor temperature? At that time, no written sources or nobody I knew could give an answer. Many years later, I got the opportunity to work at CIT Energy Management. Here the doors to the research world were opened for me and I was encouraged to conduct research. I am truly grateful for that. Research is seldom performed in solitary and for me that was certainly not the case. Obtaining the results in this thesis, and write it, has been possible only by support from several people and organisations. I would first like to express my sincere gratitude to all my colleagues at CIT Energy Management (CIT EM) for contributing to a work environment impregnated with sincerity, helpfulness and joy. Professor Emeritus Enno Abel deserves my gratitude for initiating the project of building a pilot plant. My supervisors have been Per Erik Nilsson (CIT EM) and Jan-Olof Dalenbäck (CIT EM and Department of Energy and Environment, Building Services Engineering, Chalmers). Some people outside CIT Energy Management deserve special appreciation. Leif Nilsson at Monitoring Centre, Chalmers, who both helped me a lot in the deep waters of multi variable regression analysis, developing a curve fit equation for a numerically tricky equation in the cooling tower model. Leif also supplied and rigged the monitoring system at Kvarnberget, Gothenburg, with the highest level of professionalism. Lars Eriksson and Mika Voulle at Equa, who did a great job weeding my first primitive versions of the cooling tower model (in NMF language), and patiently answering all my NMF-related questions. Per Sahlin at Equa, for lending me the tool IDA SE before CIT EM purchased it. The real estate company Vasakronan who granted access to their building at Kvarnberget in Gothenburg for the pilot plant and financed it together with BELOK. Thank you all! The financial recourses have been supplied from the Swedish Energy Agency, Swedish Building Research Programme Competitive Building, IMI Indoor Climate and CIT Energy Management. Last but not least, I would like to express my very deepest gratitude to my beloved family for their patience, understanding and support. Without that, dear reader, this thesis would never have seen daylight. Göteborg, November 2009 Bengt Bergsten

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A better burden can no man bear

on the way than his mother wit (Ej bättre börda man på vägen bär än kunskap mycken)

Havamal

VII

Table of contents

Abstract III

Preface V

Nomenclature XI

Glossary XV

1 Introduction ………………………………………………… 1

1.1 Motivation of study 1

1.2 Scope and limitations 2

1.3 Previous work 3

1.4 Outline of this work 6

2 Cooling towers – introduction and modelling …………...... 9

2.1 Introduction 9

2.2 Modelling heat and mass transfer in evaporative cooling towers 15

2.2.1 The effectiveness – NTU method 18

2.3 Visualizing heat and mass transfer in an evaporative cooling tower 23

2.4 Legionella in evaporative cooling towers 25

3 Cooling system with evaporative cooling tower ………...... 27

3.1 Introduction 27 3.2 System configurations 30

3.1.1 Parallel coupling 30

3.1.2 Series coupling 33

4 Simulation model development …………………………...... 37

4.1 Prerequisites 37

4.2 The cooling tower model (CTM) 38

4.2.1 General description 39

4.2.2 Modes of operation 40

4.2.3 Control of supply temperature of cooling tower liquid 40

4.2.4 User input data 41

4.2.5 Limitations 42

4.2.6 Equations 42

4.2.7 Determining the UA-value 57

4.3 Validation of the model 61

4.3.1 Validation by published data 61 4.3.2 Validation by data from pilot plant 66

VIII

5 Simulation of a free cooling system with

evaporative cooling tower ………………………………….. 69

5.1 Simulation tool 69

5.2 Prerequisites at base case 71

5.2.1 Model building 72

5.2.2 Analysed room 74

5.2.3 Internal heat and moisture generation 74

5.2.4 Cooling tower 76

5.2.5 HVAC system 78

5.3 Simulation methodology 81

5.3.1 Heat wave simulation 81

5.3.2 Full year simulation 81

5.4 Parameter analysis methodology 82

5.4.1 List of parameter alterations 82

5.4.2 Discussion and motivation of parameter alterations 84

6 Monitoring of free cooling system with

evaporative cooling tower ………………………………….... 95

6.1 Introduction 95 6.2 Pilot plant lay-out 95 6.3 Monitoring system 98

7 Results and analysis of simulations……………………….... 99

7.1 Cooling tower performance 100

7.2 Indoor thermal climate 105

7.2.1 Base case 106

7.2.2 Single parameter variations 111

7.2.3 Multi parameter variations 115

7.3 Energy use 123

7.3.1 Base case 123

7.3.2 Single parameter variations 125

7.4 Findings from other sources 128

8 Results and analysis of measurements…………………….. 131

8.1 Ambient conditions 131 8.2 Indoor thermal climate 133 8.3 Performance and energy use 136

IX

9 Conclusions and discussion......……………………………... 139

9.1 Conclusions 139

9.2 Discussion 142

9.2.1 Design 142

9.2.2 Indoor climate 147

9.2.3 Investments and annual costs 148

9.2.4 Applicability 150

9.2.5 Miscellaneous 150

9.3 Further research and development 152

References 153

Appendix A: NMF model 161

Appendix B: Specifications pilot plant 175

Appendix C: Specifications monitoring system 177

X

XI

Nomenclature

Latin letters

=A Area of control volume, Area of wet surface in cooling tower [m2] =a Fill interface area per fill volume [m2/m3]

=aC Heat capacity flow of air [W/°C]

=daC , Design heat capacity flow of air [W/°C]

=1liqC Heat capacity flow of Liquid 1 [W/°C]


=apc , Specific heat of dry air [J/kg, °C]

=eapc ,, Effective specific heat of air [J/kg, °C]

=ficpc ,

_

Fictitious mean specific heat [J/kg, °C]

=1,liqpc Specific heat of Liquid 1 [J/kg, °C]


=wpc , Specific heat of cooling liquid [J/kg, °C]

=Ctrl Signal from controller, 0< Ctrl <1 [-]

=∆h Enthalpy difference of moist air in the cooling tower [J/kg]

=ah Enthalpy of the surrounding air [J/kg]

=inah , Enthalpy of entering ambient air [J/kg]

=outah , Enthalpy of leaving air [J/kg]

=adh Enthalpy change of air in a control volume [J/kg]

=sh Enthalpy of saturated air at the same temperature as the water film [J/kg]

=mK Mass transfer coefficient [kg/s, m2]

=K Mass transfer coefficient [kg/s, m2]

=aM& Mass flow of air [kg/s]

=aM& Air mass flow through cooling tower [kg/s]

=daM ,& Design air mass flow through cooling tower [kg/s]

=daMMin ,_ & Minimum fraction of air mass flow through cooling tower for fan with

variable speed drive [-]

=1liqM& Mass flow of Liquid 1 [kg/s]

=dliqM ,1& Design mass flow of Liquid 1 [kg/s]


=

da

liq

M

M

&

&

Design mass flow relation of Liquid 1 and air [-]

=wm& Mass flow of cooling liquid [kg/s]

XII

=wM& Mass flow of cooling liquid [kg/s]

=speedn Speed of fan [s-1]

=dspeedn , Design speed of fan [s-1]

=∆ Fp Pressure difference in cooling tower fan [Pa]

=∆ dFp , Design pressure difference in cooling tower fan [Pa]

=FP Fan electrical power [W]

=1liqP Electrical power for the pump in Circuit 1 [W]

=∆ 1liqp Pressure difference in pump in Circuit 1 [Pa]



=sprayP Electrical power for the spray water pump [W]

=∆ sprayp Pressure difference over spray water pump [Pa]

=totalP Total electric power ( sprayliqliqF PPPP +++ 21 ) [W]

=Q Heat flux through a control volume [W]

=dQ Heat flux through a control volume [W]

=Q& Cooling capacity [W]

=dQ& Design cooling capacity [W]

=dbinaT ,, Dry bulb temperature of ambient air [°C]

=ddbinaT ,,, Design dry bulb temperature of ambient air [°C]

=dboutaT ,, Dry bulb temperature of ambient air [°C]

=wbinaT ,, Wet bulb temperature of entering ambient air [°C]

=dwbinaT ,,, Design wet bulb temperature of ambient air [°C]

=wboutaT ,, Wet bulb temperature of leaving air [°C

=∆ lnT The log-mean temperature difference

=wliqT ,1 Temperature of the warmer side of Liquid 1 [°C]

=cliqT ,1 Temperature of cold side Liquid 1 [°C]

=dcliqT ,,1 Design temperature of cold side in Liquid 1 [°C]

=dwliqT ,,1 Design temperature of warm side in Liquid 1 [°C]

=inliqT ,2 Temperature of incoming (return) Liquid 2 [°C]

=outliqT ,2 Temperature of outgoing (supply) Liquid 2 [°C]

=dAT , Design temperature approach [°C]

=dRT , Design temperature range [°C]

∆Tm = Design temperature difference between chilled water and room air for a room cooling device, e.g. a cooling beam [°C] (average water temp. and air temp. at 1,1m above floor level according to NordTest method NT VVS 078 ed. 2)

=wT Water temperature in control volume [°C]

XIII

=wdT Temperature of the water film in a control volume [°C]

=wdT Temperature change of cooling liquid in the control volume [°C]

=∆ wbT Wet bulb temperature difference of moist air in the cooling tower [°C]

=wbT Wet bulb temperature of moist air in control volume [°C]

=U Overall heat transfer coefficient (U-value) [W/ m2, °C]

=ficU Fictitious U-value [W/ m2, °C]

=eUA Effective heat transfer coefficient-area product [W/°C]

=UA Heat transfer coefficient-area product [W/°C]

=V The volume of the cooling tower fill [m3]

=sprayV& Volume flow through spray water pump [m3/s]

=w Humidity ratio of ambient air [kg water/kg dry air] Greek letters

=1liqη Total pump efficiency in Circuit 1 [-]


=sprayη Total efficiency in spray water pump [-]

=Fη Total fan efficiency [-]

=dF ,η Design total fan efficiency [-]

=hexη Efficiency of heat exchanger [-]

=1liqρ Density of Liquid 1 [kg/m3]


=aρ Air density [kg/m3]

XIV

XV

Glossary

Approach The different approach temperatures are illustrated below in a heat

exchanger diagram. Sometimes approach is called grädigkeit in heat-exchanger literature.

ASHRAE American Society of Heating, Refrigeration and Air-Conditioning

Engineers

CAV Constant Air Volume, Ventilation system with constant air flow

CIBSE The Chartered Institution of Building Services Engineers, UK

COP Coefficient of Performance

CTM The Cooling Tower Model developed in this thesis

DOE 2 Building simulation tool issued by US Department of Energy, USA

Energy Plus Building simulation tool issued by US Department of Energy, USA (successor of DOE 2)

NMF Neutral Model Format, A model language which the CTM are written in

Range See Approach and diagram above

SPF Seasonal Performance Factor (sometimes denoted SEER, Seasonal Energy Efficient Ratio)

TMY-year Test Meteorological Year

TRY-year Test Reference Year

WBT Wet Bulb Temperature

dWBT Design Wet Bulb Temperature

Liquid in primary circuit

Wet bulb (Ambient air)

temp. in

Area

Temperature

Liquid in secondary

circuit (water)

Moist air

Secondary

approach

Primary

approach

Primary

range

Secondary

range

Total

approach Cold side

Return

side

Supply side

Warm side

Wet bulb

temp. out

XVI

1

1 Introduction

This chapter gives an introduction to this thesis by discussing the motivation, scope

and limitation of this project. Previous work made by researchers is furthermore

discussed as an introduction to the field of study. Finally, the outline of the thesis is

presented with recommendations of reading for different categories of readers.

1.1 Motivation of study

The use of electricity in commercial buildings has increased over the last decades

(CIBSE, 2004). The main causes are the increased amount of office equipment in the

commercial sector together with a rapid escalation of air conditioning equipment for

handling the high internal heat gains. Adnot (2002), reports an increase in building floor

area supplied with central air-conditioning in the EU, by 200% from 1990 to 2002. The

growth in air-conditioned building floor area from present to 2020 is predicted to be an

additional 60%. Increased use of air-conditioning leads to higher demands of electric

energy, which will inevitably lead to higher emissions of carbon dioxide in the EU. This

is quite contrary to the intentions laid down in the Kyoto protocol.

To subdue further increase and reduce the energy demand in the building sector in the

EU, a directive concerning the energy performance of buildings was launched in

December 2002 (EU, 2002) and legislated in 2003. The intention with the directive is to

reduce the energy use in buildings in the EU, thus reducing the import of fossil fuels

and increasing the prospect of reaching the goals set in the Kyoto protocol.

The use of electricity in commercial buildings for air-conditioning and comfort cooling

plays a non-negligible role. According to CIBSE (2004), the share of total electric

energy use for air-conditioning for offices, industrial buildings and other buildings with

mixed use in the UK is between 2 – 20%. Figures in Sweden for office buildings show

similar values, 10 – 30% (Energimyndigheten, 2007).

Together with other major categories of electricity end-users, such as lighting, fans and

office equipment, air-conditioning is a category which definitely plays an important role

in the demand for electricity in most commercial buildings.

Conventional air-conditioning and cooling systems not only use electric energy, which

has an environmental impact, they are also a source of emission of environmentally

hazardous gases from refrigerants. Parallel with the discussion about the increase in the

use of electric energy in buildings, an ongoing process of phasing out the use of

environmentally hazardous refrigerants is carried out on an international basis, based on

the Vienna convention and the Montreal protocol.

To explore new and environmental friendly ways to control excess temperatures in

buildings therefore serves dual purposes, i.e. reduction of the use of electricity and

environmentally hazardous refrigerants.

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In this thesis, a new application of already established cooling techniques is

investigated. The application comprises the use of a cooling tower, as the sole provider

of chilled water, in conjunction with a hydronic cooling system with chilled ceilings or

chilled beams. Cooling towers is a technique which originates from the late 19th

century

and is mainly used as heat rejection in industrial applications or in chiller systems.

Hydronic cooling systems with chilled ceilings or chilled beams was introduced as a

technique a few decades ago. These systems became increasingly popular during the

1990’s and are today the predominant system in Swedish commercial buildings. In other

northern European countries this type of system has a growing market share.

Hydronic cooling systems with chilled ceilings or chilled beams are normally operated

with a chiller based on vapour compression technique. Although the common use of a

conventional chiller, these systems are sometimes classified as a low energy cooling

system per se, compared to a CAV-system (Liddament, 2000; Behne, 1997).

A cooling system comprising a cooling tower, as the sole provider of chilled water, and

a hydronic cooling system with chilled ceilings or chilled beams, is a low energy

alternative and completely free of refrigerants, thus interesting to further explore.

Apart from the environmental advantages, one key feature with this system is its high

level of applicability in both existing and new buildings. Both parts in the system, the

cooling tower and the system with chilled ceilings or beams, are well-established

techniques. When refurbishing commercial buildings in Sweden and adding comfort

cooling, systems with chilled ceiling or chilled beams are the predominant choice.

As discussed in section 1.3 Previous work, there is a limited knowledge of what indoor

climate such cooling system can provide in commercial buildings together with the

annual use of energy. There is however indications of a satisfying indoor climate and a

low energy use in the published literature. Further study of the possibilities and

limitations of this system is therefore motivated.

1.2 Scope and limitations

This thesis explore the possibilities and limitations of a comfort cooling system

consisting of a cooling tower, as the sole provider of chilled water, in conjunction with a

hydronic cooling system with chilled beams.

The basic hypothesis is that such a system in many cases can be a serious alternative to

conventional comfort cooling systems with a mechanical chiller, especially in climates

similar to conditions in the northern Europe, i.e. north of latitude 48 – 49° N in Europe.

The objective is to investigate the resulting indoor thermal climate and annual energy

use in a commercial building with such system applied. The investigation is carried out

by using an advanced building simulation program. Monitoring of temperatures and

other variables on a free cooling system as described in an existing office building has

also been made.

1 Introduction

3

The building type, which is used in the simulations, is a typical office building. The use

of the cooling system is however not restricted to offices only but can also be applied in

other types of commercial buildings.

The investigation is limited to climates similar to conditions in the northern Europe, i.e.

north of latitude 48 – 49° N in Europe, and is only discussing comfort cooling systems.

Cooling of products or processes as well as the use of water for evaporation is excluded.

1.3 Previous work

Many contributions have been made in the research for environmental-friendly and

CFC-free alternative cooling techniques and systems. Among many others the IEA

ECBCS annex 28 (Liddament, 2000) and the Pascool project (Santamouris, 1995) have

pointed out the variety of alternative cooling principles and cooling systems that can be

used in buildings.

The use of a cooling tower as a free cooling source, however in conjunction with a

vapour compression chiller, in all-air systems has been discussed by many, e.g.

Goswami and Reveliotty (1987), Mumma et al. (1990), Murphy (1991), Hipskind et al.

(1991) and Hensley (1994).

Research dealing with the application of cooling towers as the sole free cooling source

in air-water systems, i.e. systems comprising chilled ceilings or chilled beams for

cooling, is however limited. The earliest found is by Niu and Kooi (1993) and Niu et al.

(1995). Their work consisted of measurements of chilled ceilings in a laboratory and

development of a simulation program, ACCURACY, with the aim of studying the

performance of chilled ceilings. In their work, they also studied the performance of

chilled ceilings in connection with an evaporative cooling tower. The modelling of the

cooling tower was however very simplified; the supply cooling water temperature was

defined as the sum of the ambient wet bulb temperature and a constant approach

temperature. They found that with an internal heat gain of 50 W/m2 and a constant

approach temperature of 4°C or 60 W/m 2 and 2,5°C, the indoor thermal climate on a

yearly basis was acceptable for a Dutch climate, i.e. the indoor operative temperature is

not exceeding 25°C with more than 50 – 100 hours a year.

Almadari et al (1998) report about a test performed by BRE (Building Research

Establishment, London UK) where an office room was equipped with a chilled ceiling

connected to a cooling tower. At a sensible heat gain of 60 W/m2, a maximum indoor

air temperature of 25,2°C was registered.

Between 1997 - 2000 a partly EU financed research project under the Joule IV program,

called EcoCool, was conducted with Portugal, Switzerland, Germany, United Kingdom

and Finland as participating countries. The objectives of the project were:

1. To improve and optimize an existing closed wet cooling tower technology to suit a

system with chilled ceilings.

2. To develop a simple and cheap control strategy for the system considering the

building mass.

4

3. To improve the system with regard to minimizing investment and running costs as

well as energy consumption.

4. To gain knowledge of the applicability of the system, depending on climate and use

of the building.

5. To give a simple guideline, which show how to install and run the system, to building

services, consultants and HVAC installers.

The EcoCool project generated a number of published articles; Gan and Riffat (1999),

Facão and Oliveira (2000), Sprecher et al (2000), Oliveira (2000), Gan et al. (2001),

Hasan and Sirén (2002) and Hasan and Gan (2002).

Gan and Riffat (1999) use computational fluid dynamic (CFD) calculations to predict

the thermal performance of a closed wet cooling tower for chilled ceiling system. Facão

and Oliveira (2000) carry out measurements of a closed wet cooling tower for chilled

ceiling systems and obtain correlation factors which were used in existing thermal

models for such cooling towers. The article of Sprecher et al (2000) is more of a

description of the cooling system used in the EcoCool project. However, they report of

measurements of indoor temperature in an office building equipped with a chilled

ceiling system connected to a cooling tower. During the hottest days of the

measurement period from August to October 1999 with ambient air temperatures up to

32°C, the maximum indoor temperature was 27°C. Indoor temperatures over 26°C

occurred, but only for a limited time. The report by Oliveira (2000) is the final and

official report from the Ecocool project. Gan et al. (2001) use CFD calculations to

predict the thermal performance of two closed wet cooling towers for a chilled ceiling

system, and compare with experimental measurements. This work also includes

optimization of the performance of one of the cooling towers. Hasan and Sirén (2002)

use measurements from Facão and Oliveira (2000) to validate a thermal model for a

closed wet cooling tower. The model is then used to optimize the cooling tower for the

required cooling load to achieve a high coefficient of performance (COP) for the

cooling tower. The performance of the cooling system to cool an office building is also

investigated using the TRNSYS simulation program. The outcome of the simulations is

however described very briefly, as well as the prerequisite conditions. Hasan and Gan

(2002) compare analytical models with the use of CFD for a closed wet cooling tower

for chilled ceiling system, and find good agreement with the two calculation

approaches.

All articles with connection to the EcoCool project, with the exception of Hasan and

Sirén (2002) and Sprecher et al (2000), mainly discuss the potential for a cooling tower

to deliver chilled water in the range of 18-20°C, with a focus on detailed analysis of

heat transfer characteristics of a cooling tower.

Another research project with the objective to examine the ability of a cooling tower to

deliver chilled water in the range of 18-20°C and other properties of a tower has been

carried out and published by Costelloe and Finn (2000), (2001), (2003), (2007) and

(2009). Costelloe and Finn (2000) describe findings from measurements on a laboratory

test rig. The rig consists of an open counter flow cooling tower, a primary and

secondary circuit with an intermediate heat exchanger. The aim of the test facility is to

deliver cooling water at low approach temperatures for free cooling opportunities with

1 Introduction

5

chilled ceiling panels and chilled beams. Costelloe and Finn (2001) are focusing more

on the energy performance of the cooling tower test rig and publish outcome from

measurements concerning Primary energy rate, which is the inverse of the COP of the

cooling tower. Costelloe and Finn (2003) use the measurements from the test rig and

climate data of Dublin, Ireland and Milan, Italy, to calculate the availability of

maintaining a certain cooling temperature in a coolant at these climate locations and at a

total approach temperature of 3 K. The paper Costelloe and Finn (2007) outlines how

the thermal effectiveness of a cooling tower are affected by five key operating

parameters; load, ambient wet bulb temperature, primary and secondary water-flow rate

and air-flow rate. In Costelloe and Finn (2009) they discuss heat transfer correlations

from experimental results on their test rig. They also publish correlation coefficients for

calculating the total heat transfer coefficient in a cooling tower.

The most comprehensive published material concerning the resulting indoor climate is

published by Bohler et al. (2002). They investigated the performance of a cooling

system comprising an open evaporative cooling tower in conjunction with chilled

ceilings, however with an intermediate heat exchanger to avoid fouling in the chilled

water circuit. The investigation was made using the simulation tool ConsoClim. The

research work included analysis of the operative temperature in an office at a number of

different prerequisites, e.g. three design cooling capacities of the chilled ceilings, two

building inertia, two solar gains, two internal heat gains, two orientations (east and

west) and three different climates (locations of Trappes, Nice and Carpentras, all in

France). The results from all the simulation runs are quite extensive. The resulting

maximum indoor operative temperature in Trappes, located close to Paris, is in the

range of 25 – 28°C when maximum cooling loads were 40 – 70 W/m2. For Nice, with a

seaside Mediterranean climate, and Carpentras, with an inland Mediterranean climate,

the maximum cooling loads were in many cases above 70 W/m2, hence the operative

temperatures were in the range 27.5 – 30°C. However, when the maximum cooling

loads for these locations were between 50 – 70 W/m2, the operative temperatures were

in the range 25,5 – 27,5°C with only a few exceptions. Bohler et al. (2002) also

presented the operative temperature for each case when there was no cooling of the

room.

Sprecher and Tillenkamp (2003) published an investigation on a system with a closed

evaporative cooling tower and a water circuit embedded in the concrete slab. They used

the simulation program TRNSYS for investigating such a system. Sprecher and

Tillenkamp processed dynamic simulations of a large shopping centre, 10 000 m2, in

Lucerne, Switzerland, with internal heat gain from people and lighting at 42 W/m2.

During the warmest weak of the year, the indoor air temperature was kept below 26°C.

Hasan et al. (2007) uses a simulation tool (IDA-ICE) to find out the best performance of

a cooling tower combined with chilled ceiling, serving a four storey residential building.

The highest yearly mean COP achieved was 8,3.

It can be concluded that there are few published data concerning the resulting indoor

climate in buildings equipped with chilled ceilings or chilled beams connected to an

evaporative cooling tower, especially where the cooling tower is the sole provider of

chilled water. The available data indicate the possibility to achieve an indoor thermal

6

climate during hot periods that in most cases can be acceptable. The sparse data and the

limited experience of this type of cooling system is one, of many, reasons why

conventional vapour compression cooling systems still is dominating in the building

stock.

In light of the sparse but promising results from published works on the above

mentioned cooling system, it is therefore important to further investigate the resulting

indoor thermal climate and annual use of energy under different conditions.

1.4 Outline of this thesis

This thesis comprises nine main chapters and one appendix. In this section, each chapter

is briefly described to give an overview of the thesis and its content.

Chapter 1: Introduction

This chapter gives an introduction to the thesis by discussing the motivation, scope and

limitation of this project. Previous work made by researchers is furthermore discussed

as an introduction to the field of study. Finally, the outline of the thesis is presented.

Chapter 2: Cooling towers – introduction and modelling

The chapter introduces cooling towers and air-cooled heat exchangers. It also includes a

brief introduction to the physical and mathematical modelling of cooling towers. The

focus is on the effectiveness – NTU method as a theoretical base for the development of

the cooling tower model in chapter 4. Furthermore, the evaporative cooling tower

process is visualized to facilitate a better understanding of the complex nature of this

process. The chapter is finished with a short discussion about the risk of growth of

legionella bacteria in a cooling tower applied as a sole free cooling source in a hydronic

comfort cooling system.

Chapter 3: Cooling systems with evaporative cooling tower

The chapter discuss different system configurations of cooling systems with evaporative

cooling towers for free cooling in combination with a conventional chiller.

Chapter 4: Simulation model development

In this chapter a cooling tower model is developed. The development is introduced with

a discussion about the prerequisites followed by a discussion about the features of the

model and a description of its equations. The chapter is finished with a presentation of a

validation of the cooling tower model. The model is validated against both published

data and measured data by the author.

Chapter 5: Simulation of a free cooling system with evaporative cooling tower

Here the basis and the prerequisites for the simulation of a cooling system with an

evaporative cooling tower are laid out. The simulation methodology and the parameter

analysis methodology are described including a discussion and motivation of the

considered parameter alterations.

1 Introduction

7

Chapter 6: Monitoring of a free cooling system with evaporative cooling tower

This chapter describes the monitoring of a free cooling system with an evaporative

cooler in a pilot plant. The system lay-out and the monitoring system are introduced.

Chapter 7: Results and analysis of simulations

The results from the simulations described in chapter 5 “Simulation of a free cooling

system with evaporative cooling tower” are presented and analyzed in this chapter. The

results are separated into sections including Cooling tower performance, Indoor thermal

climate and Energy use. The chapter also contains relevant findings from other

researchers.

Chapter 8: Results and analysis of measurements

The measurement results from the pilot plant described in chapter 6 “Monitoring of a

free cooling system with evaporative cooling tower” are presented and analyzed in this

chapter. The results are separated into sections including Ambient conditions, Indoor

thermal climate and Free cooling system.

Chapter 9: Conclusions and discussion

In this chapter, conclusions from previous research and the results in this thesis are

presented. Conclusions are followed by a discussion concerning various aspects of

comfort cooling in general and the comfort cooling system examined in thesis in

particular.

8

9

2 Cooling towers – introduction and modelling

Although the focus is on cooling towers the title also cover air-cooled heat

exchangers. The chapter gives a short introduction to cooling towers and air-

cooled heat exchangers. It also includes a brief introduction to the physical and

mathematical modelling of cooling towers with focus on the effectiveness – NTU

method. This method forms a theoretical base for the development of the cooling

tower model in chapter 4. Furthermore, the evaporative cooling tower process is

visualized to facilitate a better understanding of the complex nature of this process.

The chapter is finished with a short discussion about the negligible risk of growth

of legionella bacteria in a cooling tower applied as a sole source in a comfort

cooling system. For more in depth studies of cooling towers, Kröger (1999) and

Hill et al. (1990) are recommended. A more detailed description of the

effectiveness – NTU method can be found in most engineering handbooks, e.g.

ASHRAE (2001a)

2.1 Introduction

Air-cooled heat exchangers and cooling towers have many applications in a broad range of industrial and commercial areas. In industrial processes like power plants, chemical and process plants and computer centrals as well as in buildings equipped with a refrigeration process, excess heat has to be discharged. The excess heat can be discharged to different type of heat sinks, e.g. to a water mass like a river or lake, to the ground water or an aquifer or to the ambient air. When excess heat is discharged to the ambient air an air-cooled heat exchanger or a cooling tower is used. The difference between an air-cooled heat exchanger and a cooling tower is not exactly defined in the literature. A simple but not exact characterisation between the two is that an air cooled heat exchanger is based on exchange of sensible heat while a cooling tower exchange heat via both sensible and latent heat. In practical terms an air-cooled heat exchanger is based on a coolant in a closed loop typically connected to a finned tube heat exchanger, see figure 2.1. The coolant, usually water, is never in direct contact with the ambient air. The coolant is chilled only through heat transfer driven by the difference of the temperature of the coolant and the dry bulb temperature of ambient air. The air movement through the heat exchanger is often accomplished with one or several fans. An air cooled heat exchanger is sometimes called a fluid cooler. This type of cooler is occasionally used in Swedish commercial buildings as a device for free cooling. Normally it is used in connection with a conventional chiller which is activated when the outdoor temperature is above 10 - 12°C. At that temperature the air cooled heat exchanger cannot produce chilled water at the normal set-point temperature, i.e. 13 - 14°C.

10

Figure 2.1 Schematic illustration of an air-cooled heat exchanger A cooling tower is a device, which typically uses a combination of heat and mass transfer to cool water, see figure 2.2. The water to be cooled is distributed in the tower by spray nozzles, splash bars or film fills in a manner that exposes a very large water surface to the ambient air. The cooling tower fill is sometimes also called packing. The movement of the air is achieved by fans (mechanical draft), natural draft or the induction effect from water sprays. A portion of the water is evaporated because the moisture content of the air is not saturated at the temperature of the water. Since this process of evaporation requires energy to change the water from liquid to vapour, heat is taken from the water, i.e. the water is cooled.

Figure 2.2 Schematic illustration of a cooling tower Cooling towers can be categorized according to different criteria:

By the air transport system

• naturally ventilated towers (natural draught cooling towers)

• mechanically ventilated towers (fan draught towers). Ventilation can be of either induced type or forced draught.

Coolant in a closed

loop system

Ambient air

Fan

Coolant in an open

loop system Ambient air

Fan

Cold water basin

Spray nozzles

Finned tube air-coil


11

By the air channelling system

• counterflow cooling towers

• cross flow cooling towers

• a combination of these two designs.

By the design of the heat exchanging surface

• open circuit systems, where water is cooled by direct contact with the surrounding

air (wet type cooling tower).

- water film fill

- splash fill

- cooling towers without fill for special applications

• closed circuit systems, where the medium to be cooled is not in direct contact with the surrounding air

- steel tube bundles with smooth external wall tubes suitable for external water spraying

- finned tube bundles, generally used for dry towers only

• hybrid cooling towers, a combination of wet and dry tower, with different fills and possibilities to switch between different modes of operation, i.e. wet only, dry only and a combination of wet and dry mode.

In this thesis the application of cooling towers is focused on cooling systems which are normally used together with building HVAC (Heating, Ventilation and Air Conditioning) and refrigerating systems. The following text is therefore narrowed to deal only with the types of cooling towers normally occurring in this context. Cooling towers in conjunction with building cooling systems can be divided in to open or closed towers. A schematic illustration of an open type of cooling tower is shown in figure 2.3. The water to be cooled is directly exposed to ambient air by either a falling water film in a fill arrangement or by droplets caused by an arrangement of splash bars. The air stream passing by catches small droplets, which are captured in a drift eliminator. In the drift eliminator, the small droplets accumulate to form larger drops that, by gravity, return to the fill or splash bars. The cooled water is collected in a basin in the base of the tower from where it is returned to the heat source.

12

Figure 2.3 Schematic illustration of an open type cooling tower In the closed type cooling tower, figure 2.4, the water or coolant is kept in a closed loop and is not directly exposed to ambient air. The heat and mass transfer is taking place in a heat exchanger, usually a smooth tube heat exchanger, which is kept wet through spraying of water over the exchanger. This type of closed loop cooling tower is sometimes also referred to as an evaporative cooler.

Figure 2.4 Schematic illustration of a closed type cooling tower The air movement is always facilitated by one or several fans regardless if it is an open or closed type cooling tower. The air can be either pushed or pulled through the tower. When pushed (by overpressure) the type is called forced draft and the fan(s) is placed where the ambient air stream enters the tower, see figure 2.5. The fan(s) forces the air

Coolant in open

loop system Ambient air

Fan

Cold water basin

Fill arrangement

Spray nozzles

Drift eliminator

Coolant in a closed loop system

Ambient air

Fan

Spray water basin

Drift eliminator

Spray nozzles


13

through the fill or the tube heat exchanger. With a forced type cooling tower, the heat exchange configuration is typically of counter flow type. Forced draft tower are characterized by relatively high air inlet velocities and low exit velocities and are thus susceptible to recirculation of the hot and moist plume air. A radial fan is the most common type of fan use in forced draft towers.

Figure 2.5 Schematic illustration of a forced draft cooling tower (illustrated with an

open type cooling tower) When the air is pulled through the cooling tower (by underpressure) the type is called induced draft, see figure 2.6. Induced draft towers may be of the counterflow or crossflow type. An axial fan is the most common type of fan used with induced draft towers and the fan is placed on top of the tower. Plume recirculation is a smaller problem in induced draft towers than it is in forced draft towers, since exit air velocity is higher. However, noise reduction is a more difficult task with this type.

Coolant in an open

loop system

Ambient air

Fan

Cold water basin

Fill arrangement

Spray nozzles

Drift eliminator

14

Figure 2.6 Schematic illustration of an induced draft cooling tower (illustrated with

an open type cooling tower). As mentioned above one or several fans cause the air movement through the cooling tower. The fan(s) can be either single speed, dual speed or equipped with variable speed control (with a variable frequency drive, VFD). Adjusting the airflow through the tower is the normal and primary way to control the temperature of the leaving fluid. If a cooling tower has only one fan it is normally cycling on/off or cycling between two or three steps of speed to control the heat rejected. If a tower has several fans the heat rejected from the unit can be controlled either by cycling each fan on/off or between two or three steps of speed and operating the fans in several steps. The lowest operating cost is achieved with fans equipped with variable speed drives (ASHRAE/IESNA, 1999 and Stout & Leach, 2002). VFD-control of fan(s) has become more common in cooling tower applications in later years. In cold climates operation of cooling towers, a fog plume can occur which can be of significant size, especially at low temperatures and high humidity. For this reason and for water saving causes, cooling towers with basically two operation modes, dry or wet operation, has been developed. Cooling towers with two or more operational modes is often called hybrid cooling towers. An example of a principal hybrid cooling tower can be seen in figure 2.7. Wet mode is activated, i.e. the spray water pump is started, when the dry bulb temperature of the ambient air is too high for sufficient heat rejection. During wet mode the coolant can pass through the dry finned coil only, or continue through the wet surface coil using a modulating three-way valve. When the dry bulb temperature is low enough the operation can be switched to dry mode and the spray water pump is stopped.

Coolant in an open

loop system

Ambient air

Fan

Cold water basin

Fill arrangement

Spray nozzles

Drift eliminator


15

Figure 2.7 Schematic illustration of a hybrid type cooling tower. Note that hybrid

cooling tower can have many different designs. This figure shows only the principle.

As mentioned earlier a cooling tower is a device which typically uses a combination of heat and mass transfer to cool water. This fact makes the physical and mathematical description more complex when engineers want to make calculations on a cooling tower. The following chapter gives a short overview of modelling heat and mass transfer in cooling towers, with focus on modelling based on the effectiveness – NTU method.

2.2 Modelling heat and mass transfer in evaporative cooling towers

In the literature, many different mathematical models have been presented to predict the heat and mass transfer in an evaporative cooling tower. These models can be separated into two main categories; analytically based and empirically based. The analytically based can further be separated into two sub categories; those requiring numerical integration of one or more differential equations and those requiring evaluation of more simple equations. The empirically based models use vendor-supplied cooling tower performance data as the base for performing calculations. When calculating cooling tower performance, different kinds of regression-techniques are used. This approach makes this type of modelling and calculating very fast compared to the analytically based models. One

Coolant in closed

loop system

Ambient air

Fan

Spray water basin

Drift eliminator

Spray nozzles

Dry finned coil

Wet surface coil

Wet deck surface

Spray water pump

Modulating three-way valve

16

example of using an empirically based model is found in DOE 2 (Benton et. al., 2002) which is a well known and widely spread building simulation tool. Analytically based models including differential equations, on the other hand, require substantially more computer time to be solved (Benton et. al., 2002). Many of these models are very complex and require specific and comprehensive data to calculate the heat and mass transfer. The analytically based models with less complexity regarding model equations lies somewhere between the two other groups mentioned when it comes to computational speed. Two of the most known in this category is the Merkel model and the effectiveness – NTU method. The Merkel model developed by Fredrich Merkel in 1925 (Merkel, 1925) has been widely used when developing cooling tower models. Cooling tower performance curves and the CTI Toolkit, released by the Cooling Technology Institute (www.cti.org), is one example where applications are based on the Merkel model. The effectiveness – NTU method is a well-established method for calculating heat exchangers in general. This method is also applicable in combined heat and mass transfer calculations of evaporative coolers and cooling towers (Jaber and Webb 1989). The method enables the cooling tower/evaporative cooler to be treated as a “black box”, i.e. the physical properties such as size and kind of fill arrangement in an open type of cooling tower, or size and dimensions of the tube heat exchanger in a closed type cooling tower, may not be known when calculating the heat transferred. A cooling tower model to be used at practical HVAC design and engineering must have a limited complexity with a manageable amount of data, yet accurate enough for applications in building simulation. This is especially important in the early stages of the building design phase when often very little information is available concerning details of the building and its technical installations, including a cooling tower. A model based on the effectiveness – NTU method has several advantages compared to models that are more complex:

- It requires relatively few input data prior to calculation - The same model can be used for calculating both wet cooling towers, with open or

closed configuration, and dry cooling towers, i.e. air cooled heat exchangers. - The accuracy of a model based on the effectiveness – NTU method is sufficient

for applications in a building simulation program. The accuracy of the model has been examined by Benton et al. (2002), see table 2.1. The error presented in table 2.1 is the difference between the supply cold water temperature predicted by the effectiveness – NTU method and that predicted by a cooling tower vendor. The figures in table 2.1 are based on 669 data points for counterflow towers and 347 data points for crossflow towers. For counterflow arrangement 95% of the errors is in the interval of –0,8°C to 1,2°C. For crossflow ditto 95% of the errors are in the interval of –1,0°C to 0,6°C. The accuracy of the model developed in this thesis, named CTM (Cooling Tower Model), which is also based on the effectiveness – NTU method, is examined in chapter 4.3 Validation of the model.


17

Table 2.1 Accuracy of the effectiveness – NTU method compared to vendor supplied data, (Benton et al., 2002)

Error [°C] Counterflow

cooling tower

Crossflow

cooling tower

Average error 0,2 -0,4

Maximum error 3,9 2,3

Standard deviation 1,0 0,6

95% Confidence Interval ±2,0 ±1,2

Bourdouxhe et al., (1994), who published the version of the effectiveness – NTU method which is used in this thesis, validated the model with catalogue data. The model was validated against a counter flow cooling tower manufactured by Baltimore Air Coil with nominal cooling capacity of 254 kW. The error presented in table 2.2 is the difference between the supply cold water temperature predicted by the effectiveness – NTU method and that predicted by Baltimore Air Coil in the same way as in table 2.1. The error figures are calculated by the author from 13 values of cold water supply temperature presented by Bourdouxhe (1994). Table 2.2 Accuracy of the effectiveness – NTU method compared to catalogue

data, (Bourdouxhe et al., 1994)

Error [°C] Counterflow

cooling tower

Average error 0,09

Maximum error 0,44

Standard deviation 0,15

95% Confidence Interval ±0,16

Hernandez et al., (1994) validated the model published by Bourdouxhe et al., (1994), with experimental data. The maximum relative difference between calculated and measured exiting cold water temperature was 0,3 % using the same cooling tower as Bourdouxhe et al. (1994). The relative difference is defined as the difference between calculated and measured supply cold water temperature divided with the difference between entering (warm) water temperature and entering (ambient) air wet bulb temperature. The effectiveness – NTU method is used in cooling tower models in well known and wide spread building simulation programs such as TRNSYS (www.trnsys.com) and EnergyPlus (Energy Plus, 2004). The cooling tower model in ASHRAE Primary HVAC Toolkit package (ASHRAE, 1999) is also based on the effectiveness – NTU method.

18

2.2.1 The effectiveness – NTU method

The effectiveness – NTU method applied on cooling towers and evaporative coolers is based on the theories of Merkel (1925). His equation states that the total heat transfer taking place at any position, e.g. a control volume, in the tower is proportional to the difference between the enthalpy of air saturated at the same temperature as the water film at one point and the enthalpy of the surrounding air at the same point.

( )dAhhKdQ asm −= (2.1)

Where

=Q Heat flux through a control volume [W]

=mK Mass transfer coefficient [kg/s, m2]

=sh Enthalpy of saturated air at the same temperature as the water film [J/kg]

=ah Enthalpy of the surrounding air [J/kg]

=A Area of control volume [m2] The difference (hs – ha) is the enthalpy driving potential proportional to the heat flux. The Merkel theory is based on the following assumptions:

1. The total heat flux is written in terms of an enthalpy difference of the moist air. 2. The effects of loss of water through evaporation are neglected in the energy

balance equation. Also the effects of drift and blow down water loss are typically neglected.

3. The leaving moist air enthalpy is assumed a function of the wet bulb temperature.

4. The thermal resistance of the water film is neglected. 5. The Lewis number is equal to unity.

The Merkel equation integrated for the whole cooling tower yields

∫ −=

2

1

w

w

T

T as

w

w hh

dT

m

KaV

& (2.2)

where =K Mass transfer coefficient [kg/s, m2]

=a Fill interface area per fill volume [m2/m3]

=V The volume of the cooling tower fill [m3]

=wm& Mass flow of cooling liquid [kg/s]

=wdT Temperature of the water film in a control volume [°C]

=− as hh The enthalpy difference and driving potential [J/kg]


19

The expression wm

KaV

& is called the cooling tower coefficient or sometimes the Merkel

number. The derivation of equation 2.1 and 2.2 can be found in Merkel (1925) and also in Kröger (1999). The first one to publish a correct derivation of a general effectiveness – NTU method for an evaporative cooling tower from the works of Merkel was Jaber and Webb, (1989). Later Braun et al., (1989) and Bourdouxhe et al., (1994) also presented a cooling tower model based on the effectiveness – NTU method. The main part of the following derivation of a cooling tower model, based on the effectiveness – NTU method, is from Bourdouxhe et al., (1994). In a control volume in an evaporative cooling tower, the following relation can be stated:

aawwpw dhMdTcMdQ && == , [W] (2.3)

where

=dQ Heat flux through a control volume [W]

=wM& Mass flow of cooling liquid [kg/s]

=wpc , Specific heat of cooling liquid [J/kg, °C]

=wdT Temperature change of cooling liquid in the control volume [°C]


=adh Enthalpy change of air in a control volume [J/kg]

Based on Merkel’s theory and the assumption that Lewis number is equal to unity, the total heat transfer (sensible plus latent) between air and water in the control volume may be defined as follows:

)(,

as

ap

hhc

dAUdQ −

⋅= [W] (2.4)

where

=U Overall heat transfer coefficient (U-value) [W/ m2, °C]

=dA Wet surface area in the control volume [m2]


It is assumed that the moist air enthalpy can be defined by the wet-bulb temperature. To enable calculations of moist air enthalpy using the wet-bulb temperature, a different definition of the specific heat of the moist air must be used. The moist air is therefore

20

treated as a fictitious ideal gas characterized by the following fictitious mean specific heat:

wb

ficp

T

hc

∆

∆=,

_

[J/kg °C] (2.5)

where

=ficpc ,

_

Fictitious mean specific heat [J/kg, °C]

=∆h Enthalpy difference of moist air over the cooling tower [J/kg]

=∆ wbT Wet bulb temperature difference of moist air over the cooling tower [°C]

The liquid side conductance is much greater than the air side conductance. Therefore, the wetted surface temperature is also assumed equal to the water temperature. Based on these assumptions, the expression in (2.4) and (2.5) yields:

)( wbwfic TTdAUdQ −⋅= [W] (2.6)

where

=ficU Fictitious U-value, defined in equation 2.7 [W/ m2, °C]

=wT Water temperature in control volume [°C]

=wbT Wet bulb temperature of moist air in control volume [°C]

and

ap

ficp

ficc

cUU

,

,⋅= [W/ m2, °C] (2.7)

The energy balance equation (2.3) can now be written as follows:

wbficpawwpw dTcMdTcMdQ ,

_

,&& == [W] (2.8)

Equation (2.8) can be rearranged

wpw

wcM

dQdT

,&

= [°C] (2.9)

and

ficpa

wbcM

dQdT

,&

= [°C] (2.10)


21

Now subtract (2.9) with (2.10), which gives

( )

−=−=−

ficpawpw

wbwwbwcMcM

QdTTddTdT,,

11&&

& [°C] (2.11)

Combine (2.6) with (2.11) gives

( )

dAcMcM

UTT

TTd

ficpawpw

fic

wbw

wbw

−=

−

−

,,

11&&

[-] (2.12)

This equation is applicable in an evaporative cooling tower. It also corresponds with the equation that occurs in the effectiveness-NTU development for an ordinary heat exchanger transferring sensible heat only:

( )

dAcMcM

UTT

TTd

CpCHpHCH

CH

−=

−

−

,,

11&&

[-] (2.13)

where index H stands for the hot side and index C stands for the cold side of a simple heat exchanger. Consequently, the cooling tower can be modelled, in steady state regime, as a fictitious indirect contact heat exchanger, see figure 2.8.

Figure 2.8 Fictitious indirect contact heat exchanger The two fluids can now be treated in the same way as in ordinary heat exchangers models. The first fluid is water (or any liquid) and the second fluid is a fictitious fluid entering the heat exchanger at the temperature Twb,in and characterized by the fictitious specific heat cp,fic . The heat exchanger is characterized by one parameter, its global heat transfer – area product, AUfic. The actual cooling tower heat transfer–area product, AU, is related to AUfic by the following expression.

cpw

Tw,w

AUfict

Twb,in

cp,fic

Twb,out

Tw,c

Air side

Liquid side

22

ficp

ap

ficc

cAUAU

,

,= [W/°C] (2.14)

Consider a case with a counterflow cooling tower, see figure 2.9. Figure 2.9 Temperature diagram of fictitious indirect contact counterflow heat

exchanger

If wpwcm ,& is the smaller heat capacity rate, the effectiveness, ε, of a cooling tower can be

defined by analogy with the effectiveness of a simple heat exchanger.

inwbww

cwww

TT

TT

,,

,,

−

−=ε [-] (2.15)

or as a function of CR and NTU

( )[ ]

( )[ ]RR

R

CNTUC

CNTU

−−⋅−

−−−=

1exp1

1exp1ε (counter flow) [-] (2.16)

where max

min

C

CCR = [-] (2.17)

( )ficpawpw cMcMC ,,min ,min && ⋅= [W/°C] (2.18)

( )ficpawpw cMcMC ,,max ,max && ⋅= [W/°C] (2.19)

Area

Temperature

Liquid to be

cooled, e.g. water

Moist air

Tw,w

Twb,in

Twb,out

Tw,c

Area 0 100%


23

and

minC

AUNTU

fic ⋅= [-] (2.20)

In the case with a crossflow cooling tower, Jaber and Webb (1989) recommends use of

the unmixed/unmixed ε-NTU relation.

( )

⋅

−⋅⋅−−=

kC

kCNTU

R

R 1expexp1ε (cross flow) [-]

(2.21) where

22,0−= NTUk (2.22)

A big advantage with applying the effectiveness – NTU method on a cooling tower model is that it is applicable on both open circuit and closed circuit cooling towers as well as air cooled heat exchangers. This is utilized in the simulation model, called CTM (Cooling Tower Model), presented in chapter 4 Simulation model development.

2.3 Visualizing heat and mass transfer in an evaporative cooling tower

To facilitate the understanding of the complex nature of the coincident heat and mass transfer in an evaporative cooling tower, the process can be visualized in a diagram. In this case, the enthalpy – humidity diagram (Mollier chart) in figure 2.10 can be illustrative. In figure 2.10, an example of the evaporative cooling tower process is shown. In this

case, a counter flow configured tower is used as an example. The cooling of the liquid,

e.g. water, is illustrated as a line on the saturation curve from 19°C (Tliq,in) to 15°C

(Tliq,out) . The ambient air is entering the cooling tower at 15°C and 50% RH, which

equals a wet bulb temperature of 9,7°C (Twb,in). The wet bulb temperature of the

ambient air is the lower limit to which the coolant can be chilled in an evaporative

process. The exhaust air from the cooling tower has a dry bulb temperature of 17°C and

a relative humidity of about 90% RH, which equals a wet bulb temperature of about

16°C (Twb,out).

24

Figure 2.10 Enthalpy – humidity diagram (Mollier chart) with an illustrated example of the evaporative cooling tower process

By using the concept of the effectiveness – NTU method, described in section 2.2.1 The

effectiveness – NTU method, the example of an evaporative cooling tower process in figure 2.10 can be visualized in an ordinary heat exchanger diagram in figure 2.11. Figure 2.11 Temperature diagram of a counter flow cooling tower. Example from

figure 2.10 is visualized together with definitions of Approach and Range.

humidity (kg/kg)

100%RF

-20

-15

-10

-5

0

5

10

15

20

25

0,000 0,005 0,010 0,015 0,020

Liquid Moist air

Enthalpy difference of entering and leaving moist air, equal to total heat exchange rate

Saturation curve

(100% RH)

Area

Temperature

Moist air

Liquid

Approach

Range

20°C

15°C

10°C

Twb,out

Twb,in

Tliq,out

Tliq,in

50% RH

Approach

Range

Twb,in

Twb,out Tliq,out

Tliq,in


25

In figure 2.11 the concept definitions of Approach and Range are visualized. The terms Approach and Range are commonly used in the cooling tower literature. Approach is simply a temperature difference showing how close the temperature of the leaving liquid is to the theoretical limit of the wet bulb temperature of the entering moist air. Range is the difference between the temperature of the entering and leaving liquid.

2.4 Legionella in evaporative cooling towers

It is beyond the scope of this thesis to discuss the whole issue about the risk of spreading legionella bacteria through aerosols originating from cooling towers and related problems with cooling water treatment. However, the issue of legionella and cooling towers has been well known within the HVAC society since the first known epidemic outbreak in 1976 in Philadelphia, USA. Cooling towers in general have since then been considered as a potential source of legionella bacteria. There are many factors that influence the rate of growth, or killing, of the legionella bacterium. The presence of water, nutriment, oxygen, time for growth, still water, microbial biofilm and pH-value between 5,5 – 9,2 together with right temperature conditions are some of the most important parameters for a positive environment for the growth of the legionella bacterium (Stålbom and Kling, 2002). The factor, of the above mentioned, that is significantly different from conventional use of a cooling tower compared to a cooling tower applied to a hydronic comfort cooling system as a sole free cooling source, is the temperature conditions. The temperature range with conventional use of a cooling tower, e.g. as a condenser or in an industrial process, is typically between 30°C – 45°C. The temperature range of a cooling tower in the context in this thesis is typically considerably lower. The hydronic cooling system temperatures is normally above the indoor dew-point temperature, but low enough to provide sufficient cooling in the room cooling devices, hence the temperature range is typically 13°C – 20°C. As seen in figure 2.12 the growth rate of legionella bacteria in conventional use of cooling towers is high or near the maximum rate, whereas for the application of cooling towers in the context of this thesis, the bacteria is in resting mode. Thus, there is a considerable difference in risk for bacteria growth between these two cooling tower applications. It can therefore be concluded that the risk for growth of legionella bacteria in a cooling tower applied as a sole free cooling source in a hydronic comfort cooling system is reasonably small.

26

Figure 2.12 Growth or killing rate for legionella bacteria at different temperatures (Stålbom and Kling, 2002)

Further reading in this matter can for example be found in ASHRAE (2000b), and Stålbom & Kling (2002).

Resting Growth Killing

Killing

Growth

Temperature

Conventional use

of cooling towers

Cooling tower as a sole

free cooling source

Growth – killing rate curve for

legionella bacteria

27

3 Cooling system with evaporative cooling tower

In this chapter, system aspects of cooling systems with evaporative cooling

towers for free cooling combined with a conventional chiller are discussed.

3.1 Introduction

This thesis explores the ability of an evaporative cooling tower to be the sole provider

of chilled water to a hydronic cooling system with chilled beams. This application can

clearly work for buildings with sufficient prerequisites, see chapter 5 Simulation of a

free cooling system with evaporative cooling tower, 7 Results and analysis of simulation

and 8 Results and analysis of measurements. However, when prerequisites are not

favourable, e.g. high cooling loads, strict temperature control, warm and humid climate

or need for dehumidification, the cooling system can be supplemented with mechanical

cooling.

In cases when a cooling system consists of both a conventional chiller, based on vapour

compression, and an evaporative cooling tower for free cooling opportunities, the

system configuration can have different layouts. The most common configurations are

discussed in section 3.2 System configurations.

However, in section 3.2 System configurations only water-side applications are

discussed, since this thesis is focusing merely on hydronic cooling systems. A

discussion concerning evaporative air-side free cooling is outside the scope of this

thesis and can be found in De Saulles (1996) and Lindholm (2000).

The question how big portion the free cooling part, i.e. the cooling tower, of the cooling

system can provide of the total annual cooling needs is of vital importance. The answer

is dependent on a number of factors. The most important are:

� The annual variation and duration of the wet bulb temperature for a given

location; the ambient wet bulb temperature constitutes the lower natural limit for the

provision of chilled water from an evaporative cooling tower and varies with climate

and geographical location.

� System configuration; different system configurations can provide different amount

of free cooling on a yearly basis, see section 3.2 System configurations.

� Required chilled water temperatures; the higher the required chilled water supply

temperature the higher the portion of annual free cooling from the cooling tower. If

dehumidification is required in the air handling unit a lower supply and return

temperature for the air cooling coil will be necessary, e.g. 6 – 7°C in supply and

about 10 - 12°C in return. The room cooling devices have higher temperature,

normally 13 – 15°C, for chilled ceilings, chilled beams or fan coils. In this case the

cooling coil will have a shorter utilization time than the room cooling devices. The

set point for the chilled water temperature is normally constant throughout the year.

However, there is usually no need for a constant temperature in an ordinary comfort

cooling system. Hence, the temperature can be allowed to rise with for example

28

lower ambient temperature. Variable chilled water temperature prolongs the potential

free cooling time.

� Design conditions of the cooling tower, i.e. primary approach and range at

design wet bulb temperature; primary approach and range at design temperature

determines the size of the cooling tower and the ability to chill the cooling water. A

further discussion about this topic is given in section 5.4.2 Discussion and motivation

of parameter alterations; Design cooling capacity (cooling tower). The actual

approach varies with the ambient wet bulb temperature and the cooling load, see

figure 3.1.

� Cooling load characteristics; how the cooling load varies diurnal and annual

determines the required load on the cooling tower. The actual approach temperature,

and hence the possible cooling water temperature, is dependent on the cooling tower

load; the lower the load, the smaller the approach.

� Conditions in control system; consists of different control regimes and its present

conditions. When the chiller starts and the system switch over from free cooling to

conventional cooling operating can be determined by the ambient wet bulb

temperature, the cooling system supply temperature or a representative indoor

temperature. Depending on the temperature chosen as the switch over condition, it

will influence the annual free cooling potential.

An example of the relation between the actual approach and the ambient wet bulb

temperature at two different ranges is displayed in figure 3.1. It should be noted that the

curves varies with the design conditions of the cooling tower.

Figure 3.1 Relation between actual approach temperatures and ambient wet bulb

temperature for an evaporative cooling tower. NB: the curves and

conditions in the figure is an example of a possible case. (data from

Marley (1982))

0

2

4

6

8

10

12

0 5 10 15 20 25

Ambient wet bulb temperature (°C)

Ap

pro

ach

te

mp

era

ture

(°C

)

Full load and a temperature

range between supply and

return water at 4,4 °C (8°F)

Half load and a temperature

range between supply and

return water at 2,2 °C (4°F)


29

A complication with evaporative cooling towers is that the performance, i.e. the ability

to chill water to a certain degree, varies with the ambient wet bulb temperature, which is

shown in figure 3.1. The performance also varies with the actual cooling load as shown

in figure 3.1. Furthermore, the performance is also determined by the design conditions

for the cooling tower, i.e. the cooling tower design approach and design range at a

certain design ambient wet bulb temperature.

In figure 3.2, the duration of the ambient wet bulb temperature and associated cooling

water temperatures from an evaporative cooling tower are shown. The available supply

cooling water temperature from a cooling tower at full and half load is indicated in the

figure and is based on the relation shown in figure 3.1. The actual cooling load is

however not constant during a year in a building. In the figure, the available cooling

water temperature is displayed resulting from a fictional example of an annual cooling

load profile, ranging from full load at high wet bulb temperatures to a fractional load at

lower wet bulb temperatures.

Figure 3.2 Duration of ambient wet bulb temperature and associated cooling water

temperatures from an evaporative cooling tower. NB: the duration curves

and conditions in the figure are only a fictional example.

In the next section, 3.2 System configurations, different configurations of free cooling

systems with an evaporative cooling tower are presented and discussed.

-10

-5

0

5

10

15

20

25

0 1000 2000 3000 4000 5000 6000 7000 8000

Time (h)

Te

mp

era

ture

(°C

)

Ambient wet

bulb temperature

Available cooling

water temperature

from cooling tower at

half load

Available cooling

water temperature

from cooling tower

at full load

Available cooling water

temperature from cooling

tower at a fictional example

of a cooling load profile

Example of approach

temperature at a wet

bulb temperature of

10°C and at full load

30

3.2 System configurations

There are a number of different system configurations where an evaporative cooling

tower can be utilized for free cooling opportunities. They are published in e.g. Murphy,

(1991), Hensley, (1994), De Saulles, (1996), Bahnfleth & Rehfeldt, (1996) and

Lindholm, (2003). The different system lay-outs can be grouped in two categories;

systems with the cooling tower coupled in parallel or in series with the cooling load.

Most of the cooling towers in the figures below are open type cooling towers. In most

cases, the open tower can be replaced by a closed loop cooling tower. In cases when the

cooling system is operating in ambient temperatures below zero, freeze protection of the

system must be considered.

3.2.1 Parallel coupling

Direct connection

Figure 3.3 Free cooling with direct connection between cooling tower and cooling

load. Dashed lines indicate closed flow paths. NB: The scheme is

intentionally simplified to improve the visual impression.

In figure 3.3, a direct system for free cooling with an evaporative cooling tower is

presented. Direct systems physically interconnect the chilled water and the condenser

water circuits during free cooling operation, enabling heat to be rejected directly by the

cooling tower.

Filter

Condenser

Evaporator

Cooling load

Filter

Condenser

Evaporator

Cooling load

Mechanical cooling Free cooling

Chilled

water

Chilled

water


31

The main benefit of this technique is that the difference between the chilled water and

the ambient wet bulb temperature, i.e. the approach, is kept to a minimum. This in turn

maximises the free cooling availability for a direct system.

The drawback of a direct system is the need for a strict water treatment management to

reduce the risk of corrosion and fouling in the chilled water circuit. A filter should

therefore be considered an integral component in this type of system. Direct systems are

sometimes denoted as a “strainer cycle”.

Indirect connection

Figure 3.4 Free cooling with indirect connection between cooling tower and cooling

load. Dashed lines indicate closed flow paths. NB: The scheme is

intentionally simplified to improve the visual impression.

In figure 3.4, an indirect system for free cooling with an evaporative cooling tower is

presented. In an indirect system the chilled water circuit is maintained as a closed loop.

The heat is rejected by means of a heat exchanger through which chilled water and

cooling tower water flow. A plate-and-frame heat exchanger is usually used in this

system configuration to minimize the temperature difference between the chilled water

and the cooling tower water.

Condenser

Evaporator

Cooling load

Heat exchanger

Condenser

Evaporator

Cooling load

Heat exchanger


Chilled

water

Chilled

water

32

As an alternative, an indirect system may incorporate a closed circuit cooling tower,

thus excluding the heat exchanger. In this case, the system is laid out as in figure 3.3.

The indirect system has the advantage of avoiding the risk of increased corrosion and

fouling since the chilled water circuit is separated from the cooling tower water. The

penalty for using a heat exchanger is an increase in the chilled water temperature to wet

bulb temperature approach. This will cause a corresponding reduction of the free

cooling potential. In many cases, this is considered to be worthwhile, since the costs for

maintenance and risk of fouling associated with direct systems may be regarded as too

high (De Saulles, 1996).

Thermosyphon

Figure 3.5 Free cooling with thermosyphon, i.e. refrigerant migration between

condenser and evaporator. Dashed lines indicate closed compressor or

flow paths. NB: The scheme is intentionally simplified to improve the

visual impression.

A thermosyphon system, figure 3.5, is based on a conventional mechanical chiller,

added with a compressor bypass arrangement. With this arrangement, when the

compressor shuts down, valves in bypass channels open to permit free migration of

refrigerant vapour from the evaporator to the condenser and the flow of liquid

refrigerant from the condenser to the evaporator. Many chiller manufacturers offer an

accessory package that can enable this free cooling application to be used.

Condenser

Evaporator

Cooling load

Condenser

Evaporator

Cooling load


Chilled

water Chilled

water

Compressor

by-pass valve

Compressor

by-pass valve


33

The heat transfer is limited to refrigerant phase-change and the load capability of

thermosyphon systems rarely exceeds about 25%, although some chiller manufacturers

can provide systems with an efficiency of up to 35% of full load. In addition, the chilled

water from the cooling tower must usually be 7°C or colder to accomplish sufficient

heat transfer (Murphy, 1991). Together, these two circumstances limit the use of this

type of system to a relatively small portion of the year and also consequently require

that the cooling load is lower (about 25%) during winter operation.

3.2.2 Series coupling

The operating principle of the tower based systems described under section 3.2.1

Parallel coupling enables the chiller to be shut down and switch the entire cooling load

to be handled by the cooling tower. Under this section, series coupling, also labelled

load shaving or load sharing, the system provides an alternative by allowing

simultaneous free and mechanical cooling. This has the advantage of allowing partial

free cooling to be exploited at times when ambient temperatures are too high for full

free cooling.

Load shaving with a common cooling tower

Figure 3.6 Free cooling with load shaving through a common cooling tower. Dashed

lines indicate closed flow paths and grey lines or objects indicate optional

usage. NB: The scheme is intentionally simplified to improve the visual

impression.

Condenser

Evaporator

Cooling load

Condenser

Evaporator

Cooling load

Heat

exchanger

Heat

exchanger


Chilled

water Chilled

water

34

Free cooling systems with load shaving, as shown in figure 3.6, are indirect by its nature

because the free cooling takes place on the return side of the chilled water circuit

through a heat exchanger. The heat exchanger is usually a plate and frame heat

exchanger to enable a close temperature approach between the cooling tower water and

the chilled water circuit. De Saulles (1996) recommend a design approach of about 1,5 –

2 K for systems operating 12 hours a day or less and about 1 K for systems operating on

a 24-hour basis.

Load shaving with a dedicated cooling tower

Figure 3.7 Free cooling with load shaving through a dedicated cooling tower. Dashed

lines indicate closed flow paths and grey lines or objects indicate optional

usage. NB: 1) The scheme is intentionally simplified to improve the visual

impression. 2) The cooling tower on the condenser circuit can also be an

air heat exchanger if desired.

In figure 3.7 free cooling with load shaving through a dedicated cooling tower is shown.

This application is also indirect, i.e. cooling tower water circuit and chilled water circuit

is separated. System control is rather simple since the level of pre-cooling can be

Condenser

Evaporator

Cooling load

Condenser

Evaporator

Cooling load


Chilled

water Chilled

water


35

allowed to float with the ambient wet bulb temperature while the chiller controls the

chilled water supply temperature in the normal way.

Disadvantages of using a dedicated tower are the higher capital and maintenance costs

together with the required extra space. However, the load shaving tower could be used

to supplement the tower in the condenser loop at times when free cooling is not

possible, i.e. at times with high ambient temperatures. In cases when the cooling load to

a sufficient degree is dependent of the ambient weather conditions this would be a

practical arrangement. When ambient temperatures are high the cooling tower rejecting

heat from the chiller will be working at full load and would benefit from the additional

capacity afforded by the load shaving tower which would otherwise be idle. This

arrangement should enable smaller towers, e.g. air heat exchanger, to be specified for

chiller heat rejection, which would partly or wholly offset the cost of the load-shaving

tower (De Saulles, 1996)

When the ambient wet bulb temperature is at least 0,5 K lower than the return chilled

water temperature the free cooling system can start pre-cool the chilled water circuit.

During such times the level of pre-cooling will vary with the ambient wet bulb

temperature. Since it is the return and not the supply chilled water temperature which

determines when energy saving is possible, the system can begin operate at relatively

high wet bulb temperature, enabling free cooling for a significant part of the year.

36

37

4 Simulation model development

In this chapter a cooling tower model is developed. The development is introduced

with a discussion about the prerequisites followed by a discussion about the

features of the Cooling Tower Model, CTM, and a description of its equations. The

chapter is finished with a presentation of the validation of the CTM.

4.1 Prerequisites

The basic reason for developing a cooling tower model was the lack of such a model in

the chosen building simulation program, IDA Indoor Climate and Energy (IDA ICE).

IDA ICE is described in chapter 5.1 Simulation tool together with a discussion about the

motives behind the selection of building simulation program.

The prerequisites for the Cooling Tower Model, CTM, are that it should comply with

the following requirements:

� Be complex enough for detailed studies of cooling system temperatures, both in

the primary and secondary circuits, air and liquid flows and power levels of

pumps and fan in cooling tower together with energy use of the pumps and fan.

� Be simple enough to use in early stages of the building design phase, i.e. have a

limited amount of input data, since detailed information about a cooling tower is

normally not available.

� Have sufficient accuracy for an application in a building simulation program. The

prime interest is not detailed analysis of the conditions inside the cooling tower,

instead the focus is on analysis of the performance of the cooling tower and the

indoor thermal climate in a building.

� Allow for different control strategies of maintaining a constant supply coolant

temperature, i.e. single speed fan with on-off control as well as fan with variable

frequency drive (VFD).

� Allow for parameter variation of design parameters.

� The required qualifications of the person using the CTM should be moderate. For

example HVAC engineers with some, but not extensive, experience of building

simulation should be able to handle the model.

Of the categories of cooling tower models described in section 2.2 Modelling heat and

mass transfer in evaporative cooling towers, the one which comply best with the

requirements above is based on the effectiveness – NTU method (ε - NTU method). In

addition to the use of the ε - NTU method in the model in this thesis, the CTM, the ε -

NTU method is also used in cooling tower models in well known and wide spread

building simulation programs such as TRNSYS (www.trnsys.com) and EnergyPlus

(Energy Plus, 2004). The cooling tower model in ASHRAE Primary HVAC Toolkit

package (ASHRAE, 1999) is also based on the ε - NTU method.

By using the ε - NTU method the same basic model can be used for simulating both wet

cooling towers, open and closed, and dry fluid coolers, i.e. air cooled heat exchangers,

which gives the CTM extra fields of application.

38

4.2 The cooling tower model (CTM)

The main purpose of the model developed in this thesis is for simulating an evaporative

cooling tower, hence the model is called the Cooling Tower Model (CTM). However,

the CTM actually describes a general air/liquid heat exchanger with the purpose to cool

a liquid with outdoor air with the ability to be used with several combinations of

arrangements; open or closed and wet or dry, see table 4.1. The CTM has therefore a

wide field of application and can be used to simulate both cooling towers, open and

closed circuit type, and any air/liquid heat exchanger with the purpose to cool a liquid

with outdoor air. The matrix in table 4.1 shows the different arrangements the CTM has

including only sensible heat exchange as well as both sensible and latent heat exchange.

The CTM implies that cooling is supported by one or several fans inducing forced air

movement.

Table 4.1 Table showing different possible arrangements of the CTM in this thesis.

Open Closed

Wet

Cooling tower with spray nozzles

distributing cooling water over fill packing

in an open cooling circuit. Heat rejection

contains both sensible and latent heat. See

figure 2.3

Cooling tower with closed circuit system.

Spray nozzles distribute spray water over

tube bundles containing a closed circuit

coolant. Heat rejection contains both

sensible and latent heat. See figure 2.4

Dry

This combination is impossible in practical

terms.

Air cooled heat exchanger with coolant

contained in a closed circuit system. Heat

rejection contains only sensible heat. See

figure 2.1

Apart from the different combinations of arrangements described in table 4.1, the CTM

has two possible modes of fan speed control; single speed or variable speed drive

(VFD), and two possible heat exchange configurations; counter flow or cross flow.

The CTM is based on a cooling tower model published by Bourdouxhe et al. (1994),

which also is described in ASHRAE Primary HVAC Toolkit (ASHRAE 1999). The

ASHRAE model is based on the ε - NTU method. In this thesis, the ASHRAE model is

a part of the CTM which has been supplemented with different combinations of

arrangements, described in table 4.1, together with different modes for fan speed control

and heat exchange configurations as described above. Equations for handling part load

operation together with sequenced control strategies for maintaining a constant supply

cooling water temperature at different cooling loads and ambient conditions are also

added to the CTM. All these features make the CTM complex but also increase its

possible fields of application.

The CTM is written in the Neutral Model Format (NMF) language. NMF is a formal

language used in IDA ICE to describe the equations and conditions in a model. In

simple terms NMF describes what to be calculated but not how it is executed. The

numerical execution is handled by a general-purpose equation solver (IDA Solver). The

CTM consists of 47 equations comprising 60 variables and 59 parameters.


39

4.2.1 General description

In this section, an overview is given describing the main features of the CTM. The

complete model, written in NMF language, is presented in appendix A.

In the CTM, the user can choose between wet or dry operation mode. The wet mode

implies the option of cooling liquid spraying either through a cooling tower fill (open

mode) or a separate water circuit spraying over the closed circuit tube bundles

containing the cooling liquid (closed mode).

The CTM includes a freeze-protected circuit, Circuit 1 or the primary circuit, which is a

closed loop between the cooling tower and the heat exchanger, which separates Circuit

1 from Circuit 2. Circuit 2, or the secondary circuit, is equal to the closed cooling water

loop connected to the air cooling coil in the Air Handling Unit (AHU) and the cooling

device(s) in the zone(s), e.g. chilled ceilings, chilled beams or fan coils. A scheme of the

CTM and its connection with the rest of a possible comfort cooling system is shown in

figure 4.1. Note that the scheme is simplified concerning fittings.

Figure 4.1 The Cooling Tower Model (CTM) and its connection with the rest of a

possible cooling system.

The liquid in Circuit 1 is called Liquid 1 and subsequently the liquid in Circuit 2 is

called Liquid 2. If Circuit 1 has a freeze protective function, Liquid 1 is a freeze

protecting brine. Circuit 2 is usually containing water. If the user wants to omit the

freeze-protected circuit a few parameters in the CTM is simply altered. A detailed

description for omitting the freeze-protected circuit is presented in appendix A.

Cooling tower or air-

cooled heat exchanger

Heat exchanger

Hydronic room cooling

devices, e.g. chilled

beams or fan coils

Possible chiller

Circuit 2 (secondary):

Water based circuit

Circuit 1 (primary):

Freeze-protected

brine

To the air cooling coil in the

Air Handling Unit (AHU)

System boundaries

for the cooling

tower model

40

4.2.2 Modes of operation

The cooling tower model can be set to different modes by the user.

� Mode_WetDry; is wet or dry operation as described above in table 4.1.

� Mode_Fan; decides if the fan(s) is/are single speed or equipped with variable

speed drive. The single speed mode is applicable when cooling tower has only one

fan (or maybe two fans) with single speed. If the cooling tower has several fans

with single or two speed operation the user should choose the mode with variable

speed drive. This mode simulates the case with multiple fans cycling on-off in

sequence fairly well. If cooling tower is equipped with variable speed drive the

choice of mode is obvious.

� Mode_ClOp; decides if the cooling tower or the air cooled heat exchanger has an

open circuit or closed circuit. The mode controls if spray water is on or off. This

mode is added because the combination of Closed circuit mode and Wet mode use

spray water. All other combinations of Open/Closed and Wet/Dry modes use no

additional spray water.

� Mode_Tower; lets the user choose between a counterflow configuration or a

crossflow configuration. The crossflow configuration assumes both streams

unmixed.

4.2.3 Control of supply temperature of cooling tower liquid

The cooling tower model has two different ways to control the tower and the supply

temperature of the cooling tower liquid.

� First is a simple on/off control of the whole tower via an external signal from a

time schedule through a link (Tower_control).

� Secondly, the leaving Liquid 2 temperature is controlled in two sequences; first by

modulating the mass flow through the pump in Circuit 1 and of air in the fan(s).

This is done via an external signal from a PI-controller, see figure 4.2. In single

speed mode the real life cycling on-off is simulated as a linear relation of the mass

flow from a minimum flow (15% of max.) up to maximum flow (ASHRAE/

IESNA 1999). In variable speed mode, the fan laws are applied controlling the

mass flow.

Figure 4.2 Control of cooling tower model through sequenced control

PI-controller


41

4.2.4 User input data

The CTM gives the user the following optional design parameters as input data. In early

stages of a building design phase, only some of the first category parameters are

necessary to address. The rest of the parameters can normally remain default until later

stages in the design phase.

1) Input parameters affecting size and cooling capacity of the cooling tower and

the intermediate heat exchanger:

QTower_d Design cooling capacity in tower (in kW)

TApproachP Temperature difference between inlet air (TAir_in_d) and outlet

Liquid 1 (TLiq1Cold_d) at design stage. P = Primary circuit.

TRange_P Temperature difference between inlet Liquid 1 (TLiq1Warm_d) and

outlet Liquid 1 (TLiq1Cold_d) at design stage

TApproachS Design difference between TLiq1Cold_d and TLiq2Out in secondary

circuit. S = Secondary circuit.

TRange_S Design difference between TLiq2Out and TLiq2In in secondary circuit

QCooling_d Design cooling load in building (in kW)

TAir_indb_d Design inlet dry bulb air temperature

RelHum_d Design relative humidity of ambient air [%)

MLiq1MAir_d Massflow relation of Liquid 1 and Air (MLiq1_d/MAir_d). Normally

in the range 0,5 to 2.

2) Input parameters primarily affecting energy efficiency of the cooling tower:

EtaFan_d Fan efficiency (in tower) at design airflow rate

dpFan_d Pressure difference over fan at design rate mass flow of air in tower

dpLiq1 Pressure difference in Circuit 1 (Liquid 1), assumed constant

dpLiq2 Pressure difference in Circuit 2 (Liquid 2 passing heat exchanger),

assumed constant

dpSpray Pressure difference in spray water circuit (open) assumed constant

Mode_Fan Control mode; 0=Fan with variable speed control, 1=Single speed fan

EtaLiq1 Total pump efficiency in circuit with Liquid 1

EtaLiq Total pump efficiency in circuit with Liquid 2

EtaSpray Total pump efficiency in (open) circuit with spray water

3) Miscellaneous input parameters:

Min_MAir Min. airflow in fraction of design rate mass flow of air in tower. When

using variable speed drive Min_MAir is normally in the range 0,1 to

0,2

cpLiq1 Liquid 1 specific heat (default 3685= 40% propylene glycol at 15°C)

rhoLiq1 Liquid 1 density (default 1039= 40% propylene glycol at 15°C)

42

Mode_WetDry Control mode; 0 = Wet mode, 1 = Dry mode

Mode_Spray Control mode; 0 = Closed circuit tower, 1 = Open circuit tower

Mode_Tower Control Mode; 0 = Counter flow tower, 1 = Cross flow tower

TLiq2Out_set Liquid 2 outlet set temperature

4.2.5 Limitations

The CTM has like most other mathematical models some limitations. The main

limitations are listed below:

• The CTM does not calculate the consumption of makeup water for evaporation,

drift losses or blow down.

• Hybrid cooling towers with combination of, or alteration between, wet and dry

mode during operation cannot be simulated within the CTM.

• Natural draft cooling towers cannot be simulated in the CTM.

• The CTM works with good accuracy between -20°C to + 40°C wet bulb

temperature of ambient air. It should not be used outside this temperature range.

For further information, see equation 4.11 to 4.13 and the adjacent discussion.

4.2.6 Equations

The cooling tower part of the CTM complies with the cooling tower model in ASHRAE

(1999). The equations presented in this chapter are the most important ones governing

the CTM. For a complete presentation of all equations of the CTM, see appendix A.

Figure 4.3 shows the layout and basic terms in the cooling tower model.

Enthalpy of entering air stream

The enthalpy of ambient air entering the cooling tower is:

( )dbinaapdbinaina TwcTh ,,

3

,,,, 1805102501 ⋅+⋅+⋅= [J/kg, °C] (4.1)

where


=dbinaT ,, Dry bulb temperature of ambient air [°C]

=apc , Specific heat of ambient dry air [J/kg, °C]

=w Humidity ratio of ambient air [kg water/kg dry air]


43

Figure 4.3 Cooling tower model; schematic layout and notations

Enthalpy of leaving air stream in wet mode:

In wet mode the enthalpy of the leaving air can be calculated by an approximate relation

only dependent on the wet bulb temperature of the leaving air (ASHRAE, 1999);

3

,,

2

,,,,, 98855,035,111,17865,9362 wboutawboutawboutaouta TTTh ⋅+⋅+⋅+= [J/kg, °C] (4.2)

where


=wboutaT ,, Wet bulb temperature of leaving air [°C]

In the cooling tower model, this equation is only used in the Parameter processing

section where the design UA-value is calculated. In the part of the CTM which is

calculated each timestep, there is however no need to calculate the enthalpy of the

leaving air. The original equation from ASHRAE (1999), i.e. ( )

( )wbinawbouta

inaouta

eapTT

hhc

,,,,

,,

,,−

−= ,

in where the enthalpy of the leaving air is a part, is replaced by a curve fit equation in

the CTM. The curve fit equation of the variable eapc ,, is a function of the mean wet bulb

temperature and the wet bulb temperature difference of entering and leaving moist air.

This is made to ensure a better numerical stability of the CTM. For further information,

see equations 4.10 to 4.13.

Air out

Ambient air in Liquid 1, Cold

Liquid 1, Warm Liquid 2, In

Liquid 2, Out

Cooling

tower Heat

exchanger

44

Enthalpy of leaving air stream in dry mode:

The enthalpy of air leaving the cooling tower is:

( )dboutaapdboutaouta TwcTh ,,

3

,,,, 1805102501 ⋅+⋅+⋅= [J/kg, °C] (4.3)

where


=dboutaT ,, Dry bulb temperature of ambient air [°C]

Control of supply temperature, i.e. Liquid 2 temperature, Tliq2,out

The aim of controlling Tliq2,out, is to keep the supply temperature at the requested

temperature (Tliq2,out,r). The UA-value, and consequently the leaving temperature of

Liquid 2, Tliq2,out, is controlled through sequenced control of the airflow through the

tower. When Tliq2,out +0.1 < Tliq2,out,r and rising, the first step to control Tliq2,out is by

cycling fan on-off, i.e. UA-value is cycling between zero and UA=f(min of aM& ) where

aM& is the air flow trough the cooling tower. The second step is carried out by modulate

the air flow and Liquid 1 flow between minimum flow rate and maximum flow rate to

achieve the desired temperature of Tliq2,out.

When the fan is on, the air mass flow, aM& , is controlled in a linear relation to the

control signal from the PI-controller. In figure 4.4, the relation airflow – control signal

for a fan in single speed mode is shown. In single speed mode the real life control

strategy by cycling the fan on-off is emulated as a linear relation of aM& from the design

air mass flow, daM ,

& , to 0,15 daM ,& (AHRAE/IESNA, 1999), see equation 4.4. aM&

should is in this case be considered as a mean value of the air mass flow over a time

period.

For a fan with variable speed control, the relation airflow – control signal is almost the

same as in figure 4.4, but in this case the minimum airflow is optional depending on

type of variable speed drive, see equation 4.5. The default minimum value in CTM is

0,1daM ,

& for a fan with variable speed drive.


45

Figure 4.4 Relation airflow versus control signal from PI-controller

The equation for a fan with single speed mode is as follows:

( )CtrlMCtrlMM dadaa −⋅+⋅= 115.0 ,,&&& [kg/s] (4.4)

where

=aM& Air mass flow through cooling tower [kg/s]

=daM ,& Design air mass flow through cooling tower [kg/s]

=Ctrl Signal from controller, 0< Ctrl <1 [-]

For a fan with variable speed control, equation 4.5, aM& is linear between daM ,& and a

minimum flow equal to dada MMMin ,,_ && ⋅ ( daMMin ,_ & is typically 0.1 to 0.2 depending

on type of variable speed drive).

( )CtrlMMMinCtrlMM dadadaa −⋅+⋅= 1_ ,,,&&&& [kg/s] (4.5)

where

=daMMin ,_ & Minimum fraction of air mass flow through cooling tower for fan

with variable speed drive [-]

0,0

0,2

0,4

0,6

0,8

1,0

0,0 0,2 0,4 0,6 0,8 1,0

Control signal (-)

Re

lati

ve

air

flo

w (

-)

46

The mass flow of Liquid 1, 1liqM& , is always regulated with variable speed drive.

Minimum flow is set to 10% of 1liqM& .

( )CtrlMCtrlMM dliqdliqliq −⋅+⋅= 11.0 ,1,11&&& [kg/s] (4.6)

where



The mass flow rates of air and Liquid 1 influence the heat transfer in the cooling tower,

i.e. the UA-value. In the literature many different kinds of correlations have been

presented to enable part load simulations of a cooling tower or an air-cooled heat

exchanger, e.g. Baker & Shryock (1961), Parker & Treybal (1961) and Mitzushina et al.

(1967), A common way of correlate the UA-value to variations in air flow and liquid

flow rates is shown in equation 4.7.

m

liq

n

a MMCUA && ⋅⋅= [W/°C] (4.7)

where C, n and m normally are determined empirically for a certain object. In the

literature several values of C, n and m has been presented. If the UA-value can be

determined at the design flow rates of air and liquid, UAd, the UA-value can be

expressed as follows:

m

dliq

liq

n

da

ad

M

M

M

MUAUA

⋅

⋅=

,,&

&

&

&

[W/°C] (4.8)

Lebrun and Silva (2002) has gathered and tabled values of the exponents n and m for

cooling towers from four different sources including themselves, with a total of seven

different values of n and m respectively. The mean values of n and m from those values

are

65.0=n (with min 39.0=n and max 03.1=n )

43.0=m (with min 15.0=m and max 67.0=m )

For air cooled heat exchangers (dry operation) Wetter (1999) has presented values for n

and m, see table 4.2 and below.


47

Table 4.2 Values for exponent n in UA-correlation for air-cooled heat exchangers

(Wetter, 1998)

Tube

arrangement

Reynolds number

(Re)

Exponent n

(air side)

In-line Re≤ 4·104 0.72

Staggered 1·103 ≤ Re ≤ 2·104 0.65

Staggered 2·104 ≤ Re ≤ 2·105 0.80

Staggered Re ≥ 2·105 0.95

The reference dimension in Reynolds number is the fin pitch. The exponent m for the

liquid side is only given as 0,85 and not depending on Reynolds number.

The UA-value correlation in the present model, in wet mode, is as follows:

43,0

,1

1

65,0

,

⋅

⋅=

dliq

liq

da

ad

M

M

M

MUAUA

&

&

&

&

[W/°C] (4.9)

The values of n and m in equation 4.9 are default values in the CTM but can easily be

changed by the user. The UA-value correlation with default values of n and m, as in

equation 4.9, is shown in figure 4.5. The correlation represents an almost strait line.

Figure 4.5 UA-value correlation according to equation 4.9 with default values of n

and m where 65.0=n and 43.0=m . Both aM& and 1liqM& are varied

simultaneously.

0,0

0,2

0,4

0,6

0,8

1,0

0 0,2 0,4 0,6 0,8 1

Relative flow rate (-)

Re

lati

ve

UA

-Va

lue

(-)

48

Cooling tower

Equations in this section are based on the cooling tower model in ASHRAE (1999) and

equations presented in chapter 2 Cooling Towers – Introduction and modelling, section

2.2. To supplement the ASHRAE model a dry operation mode is added as well as a

distinction between open or closed type cooling tower. The difference between open or

closed mode is however only a spray water pump adding a small amount of extra energy

for the closed type cooling tower.

A fictitious mean specific heat of moist air was introduced in section 2.2, equation 2.5.

The fictitious mean specific heat is in the CTM called the effective specific heat of

moist air, eapc ,, , complying with the terminology in ASHRAE (1999). eapc ,, is only used

in wet mode and it is written:

( )( )wbinawbouta

inaouta

eapTT

hhc

,,,,

,,

,,−

−= [J/kg °C] (4.10)

where

=eapc ,, Effective specific heat of air [J/kg, °C]



=wboutaT ,, Wet bulb temperature of leaving air [°C]

=wbinaT ,, Wet bulb temperature of entering ambient air [°C]

For dry mode operation the specific heat of dry air, apc , is used.

The original equation 4.10 is in the CTM replaced by a curve fit equation where

( )dmeap TTfc ,,, = , see equation 4.11. This is made to ensure a better numerical stability

of the CTM. The effective specific heat of air is expressed as:

30745,11961084999,1

1009962,265114,10351,01056473,0

224

44112344

,,

+⋅+

⋅++−⋅=

−

−−

dm

dmmmmeap

TT

TTTTTc (4.11)

The curve fit equation comprises two variables; mT and dT :

wbinawboutad TTT ,,,, −= (ambient wet bulb temperature difference, [°C]) (4.12)

( )

202

,,,,+

+=

wbinawbouta

m

TTT (mean ambient wet bulb temperature, [°C]) (4.13)


49

Equation 4.11 implies that the mean ambient wet bulb temperature, mT , has an offset of

20 K to place the zero value of mT on the lowest end of the allowed ambient wet bulb

temperature range (-20 < mT < 40°C). The curve fit equation has a relative error < 1%

between –15 < mT < 40°C and dT < 15°C. The relative error of the effective specific

heat, eapc ,, , is the relative difference in percentage between the original equation (4.10)

and the curve fit equation (4.11). At normal operation of a cooling tower with the

application of chilling water in a comfort cooling system, mT and dT will be well within

these limits. A graphic presentation of the accuracy of the curve fit equation is presented

in figure 4.6. A relative error of < 1% in the effective specific heat, eapc ,, , generates an

error on the leaving Liquid 2 temperature which is less than 0.1°C. Increasing to the

maximum range allowed; -20 < mT < 40 °C and dT < 30°C, the relative error in the

effective specific heat, %4,3,, <eapc .

Figure 4.6 Accuracy of the curve fit equation (4.11) compared to the original

equation (4.10) describing the effective specific heat, eapc ,, , of moist air.

Accuracy is expressed as the relative difference in percentage between

the original equation and the curve fit equation.

-4%

-3%

-2%

-1%

0%

1%

2%

3%

4%

-20 -10 0 10 20 30 40

Mean wet bulb temperature, Tm (°C)

Accu

racy (

%) 0,01

5,0

10,0

15,0

20,0

25,0

30,0

Ambient wet bulb

temperature

difference, Td [°C]

Maximum range of ambient wet bulb

temperature when operating in the northern

parts of Europe. The lower end depends on type

of cooling tower and cold weather operation

50

The name “fictitious heat transfer coefficient-area product”, which was introduced in

chapter 4.2.1, is in this chapter replaced by the expression “effective heat transfer

coefficient-area product”, eUA , to comply with the terminology in ASHRAE (1999).

ap

eap

ec

cUAUA

,

,,⋅= [W/m

2, °C] (4.14)

where

=eUA Effective heat transfer coefficient-area product [W/°C]

=UA Heat transfer coefficient-area product [W/°C]


The heat capacity flow of air, Liquid 1 and Liquid 2 is as follows:

11,1 liqliqpliqMcC &⋅= [W/°C] (4.15)

22,2 liqliqpliqMcC &⋅= [W/°C] (4.16)

aeapaMcC &⋅=

,, [W/°C] (4.17)

where







=aC Heat capacity flow of air [W/°C]


The smallest and biggest heat capacity flow Cmin and Cmax is

( )aliq CCC ,min 1min = [W/°C] (4.18)

( )aliq CCC ,max 1max = [W/°C] (4.19)


51

The relation of Cmin / Cmax is defined as

max

min

C

CCR = [-] (4.20)

The Number of transfer units, NTU, in the cooling tower can now be written

minC

UANTU e= [-] (4.21)

The effectiveness of the heat exchanger, ε, depending of configuration of the cooling

tower or air cooled heat exchanger, is written:

Counterflow configuration:

( )[ ]( )[ ]

RR

R

CNTUC

CNTU

−−⋅−

−−−=

1exp1

1exp1ε [-] (4.22)

Crossflow configuration:

( )

⋅

−⋅⋅−−=

kC

kCNTU

R

R 1expexp1ε [-] (4.23)

where

22,0−= NTUk (4.24)

The water to air heat transfer rate in the cooling tower, Q& , i.e. the cooling capacity:

Wet mode:

( )wbinawliq TTCQ ,,,1min −⋅= ε& [W] (4.25)

Dry mode:

( )dbinawliq TTCQ ,,,1min −⋅= ε& [W] (4.26)

where

=wliqT ,1 Temperature of the warmer side of Liquid 1 [°C]

Leaving air dry or wet bulb temperature from cooling tower or air cooling heat

exchanger, wboutaT ,, or dboutaT ,, , has the following expression:

52

Wet mode:

+=

a

wbinawboutaC

QTT

&

,,,, [°C] (4.27)

Dry mode:

+=

a

dbinadboutaC

QTT

&

,,,, [°C] (4.28)

Outlet Liquid 1 temperature, cliqT ,1 , i.e. the colder side of Liquid 1, from tower is

calculated as follows:

−=

1

,1,1

liq

wliqcliqC

QTT

& [°C] (4.29)

Heat exchanger

According to elementary heat exchanger theory the outlet temperature of Liquid 2,

outliqT ,2 , from heat exchanger between Liquid 1 and Liquid 2 is written as:

If Cliq2 ≤ Cliq1:

( )cliqinliqhexinliqoutliq TTTT ,1,2,2,2 −−= η [°C] (4.30)

If Cliq2 > Cliq1:

( )cliqinliq

liq

liq

hexinliqoutliq TTC

CTT ,1,2

2

1

,2,2 −

−= η [°C] (4.31)

where

=inliqT ,2 Temperature of incoming Liquid 2 [°C]

=outliqT ,2 Temperature of outgoing Liquid 2 [°C]

=cliqT ,1 Temperature of cold side Liquid 1 [°C]

=hexη Efficiency of heat exchanger [-]

The warm side temperature of Liquid 1, wliqT ,1 , from heat exchanger between Liquid 1

and Liquid 2 is calculated as:


53

( )outliqinliq

liq

liq

cliqwliq TTC

CTT ,2,2

1

2

,1,1 −

+= [°C] (4.32)

Use of energy

The speed of fan with variable speed control is written:

=

da

a

dspeedspeedM

Mnn

,

, &

&

[s-1

] (4.33)

where

=speedn Speed of fan [s-1

]

=dspeedn , Design speed of fan [s-1

]

The pressure difference in a cooling tower fan with variable speed control can, using the

fan laws, be written as:

2

,

,

∆=∆

dspeed

speed

dFFn

npp [Pa] (4.34)

For a fan with single speed mode the pressure difference is:

dFF pp ,∆=∆ [Pa] (4.35)

where

=∆ Fp Pressure difference in cooling tower fan [Pa]

=∆ dFp , Design pressure difference in cooling tower fan [Pa]

The total fan efficiency, F

η , is often assumed constant in models of energy use of a fan.

The total fan efficiency is however not constant, especially when a variable speed drive

is applied. In this model the total fan efficiency is modelled as a function of fan motor

speed. As stated in equation 4.33 the fan motor speed is direct proportional to air mass

flow through the fan. The relation for Fη with variable speed drive is chosen as a power

law expression. This is an approximation for the fact that the efficiency is a function of

air mass flow of the fan itself and the speed of the variable speed drive. The expression

for the variable speed drive is similar to a power law curve.

5.0

,

,

=

da

a

dFFM

M

&

&

ηη [-] (4.36)

54

where

=Fη Total fan efficiency [-]

=dF ,η Design total fan efficiency [-]

The exponent is chosen to 0.5 as default, but can easily be changed by the user. For a

single speed fan the total efficiency is constant:

dFF ,ηη = [-] (4.37)

The electrical power for fan(s) in a cooling tower, FP , is expressed as:

Fan with variable speed control:

Fa

aF

F

MpP

ηρ ⋅

⋅∆=

&

[W] (4.38)

Fan with single speed mode:

( )dFdadFa

Fa

F pMpMP ,,, 15.085.0

1∆⋅⋅−∆⋅

⋅⋅=

ηρ [W] (4.39)

where

=FP Fan electrical power [W]

=aρ Air density [kg/m3]

The linear relation in equation 4.39 corresponds to a relation in AHRAE/IESNA (1999)

between power and mass flow of a single speed fan. This equation emulates the relation

between mean power requirement and mean flow when a single speed fan is switching

between on or off mode. In figure 4.7, the equations 4.38 and 4.39 are visualized. The

two curves for variable speed drives are for comparing constant versus variable total fan

efficiency (eta const. vs. eta variable).


55

Figure 4.7 Electrical power requirement for a fan as a function of air mass flow in

the CTM.

The electrical power for the pump in the primary circuit (Circuit 1) between cooling

tower and heat exchanger, i.e. Liquid 1, is expressed as follows:

11

11

1

liqliq

liqliq

liq

MpP

ηρ ⋅

⋅∆=

&

[W] (4.40)

where





The pump in the primary circuit is assumed to have a variable speed drive, see equation

4.6. The total pump efficiency is however assumed to be constant in contrary to the total

fan efficiency. This simplification is done since the annual electrical use of the pump in

Circuit 1 is a relatively small part, about 10%, of the total annual electrical use of the

cooling tower or the air-cooled heat exchanger.

0,0

0,1

0,2

0,3

0,4

0,5

0,6

0,7

0,8

0,9

1,0

0,0 0,2 0,4 0,6 0,8 1,0

Relative air mass flow

Re

lati

ve

re

qu

ire

d e

lectr

ica

l p

ow

er

Single speed Variable speed, eta variable Variable speed, eta const.

56

As for the pump in circuit 1, the electrical power for pumping Liquid 2 through heat

exchanger in Circuit 2 is expressed in the same way:

22

22

2

liqliq

liqliq

liq

MpP

ηρ ⋅

⋅∆=

&

[W] (4.41)

where





When the user has chosen to simulate a wet (evaporative) closed cooling tower the

electrical power demand for a pump in the water spray sump is calculated as follows:

spray

sprayspray

spray

VpP

η

&⋅∆= [W] (4.42)

where

=sprayP Electrical power for the spray water pump [W]

=∆ sprayp Pressure difference over spray water pump [Pa]

=sprayV& Volume flow through spray water pump [m3/s]

=sprayη Total efficiency in spray water pump [-]

and

1000

018.0 dspray

QV

&&

⋅= [m

3/s] (4.43)

where

=dQ& Design cooling capacity of cooling tower [kW]

Equation 4.43 is based on an approximate relation between the volume flow of spray

water and the cooling capacity of an evaporative closed circuit cooling tower which is

assumed to be 0.018 l/s per kW cooling capacity in tower (Stoecker, 1998).


57

A coefficient of performance, COP, can be defined for a cooling tower or air-cooled

heat exchanger as:

( )( )WpoweryelectricitTotal

WcapacityCoolingCOP = ,

or

totalP

QCOP

&

= [-] (4.44)

where

=Q& Cooling capacity [W]

sprayliqliqFtotal PPPPP +++= 21 [W] (4.45)

4.2.7 Determining the design UA-value

The CTM requires a UA-value to be known at the maximum design operating point. A

design UA-value is very seldom known by the user, especially in the early stages of the

building design phase.

In the CTM some of the parameters supplied by the user in section 4.2.4 User input

data, are used to calculate a UA-value at design conditions, see table 4.3.

Table 4.3 Table showing input data for the calculation of the design UA-value of a

specific cooling tower.

Parameter

name

Symbol Description

QTower_d dQ& Design cooling capacity in tower (in kW)

Tapproach_d dAT , Design temperature difference between inlet air

(TAir_in_d) and outlet Liquid 1 (TLiq1Cold_d).

For wet mode, TAir_in_d is equal to the design

ambient wet bulb temperature.

Trange_d dRT , Design difference between inlet Liquid 1

(TLiq1Warm_d) and outlet Liquid 1

(TLiq1Cold_d).

TAir_indb_d ddbinaT ,,, Design inlet dry bulb air temperature

RelHum_d dRH Design relative humidity of ambient air [%]

MLiq1MAir_d

a

liq

M

M

&

&

Mass flow relation of Liquid 1 and Air

(MLiq1_d/MAir_d). Normally in the range of 0,5

to 2.

58

For wet mode the design wet bulb temperature, dwbina

T,,,

, is calculated from ddbina

T,,,

and

dRH by subroutines supplied with IDA Indoor Climate and Energy (IDA ICE), which is

the simulation program this model is implemented in. For more information about IDA

ICE, see chapter 5.1

According to the effectiveness – NTU method presented in chapter 2.2 the cooling

tower, as well as an air cooled heat exchanger, can be represented as a simple heat

exchanger, figure 4.8.

Figure 4.8 An evaporative cooling tower modelled as an indirect contact heat

exchanger

From given and calculated parameters the entering and leaving liquid temperatures are

easily calculated:

Wet mode:

dAdwbinadcliq TTT ,,,,,,1 += [°C] (4.46)

Dry mode:

dAddbinadcliq TTT ,,,,,,1 += [°C] (4.47)

and

dRcliqdwliq TTT ,,1,,1 += [°C] (4.48)

where

=dcliqT ,,1 Design temperature of cold side in Liquid 1 [°C]

=dwliqT ,,1 Design temperature of warm side in Liquid 1 [°C]

=dwbinaT ,,, Design wet bulb temperature of ambient air [°C]

=ddbinaT ,,, Design dry bulb temperature of ambient air [°C]

=dAT , Design temperature approach [°C]

=dRT , Design temperature range [°C]

Tliq1,c

UAe,d

Ta,in,wb

Tliq1,w

Air side

Liquid side

Ta,out,wb


59

After calculating dcliqT ,,1 and dwliqT ,,1 , one temperature is still unknown, i.e. dwboutaT ,,, in

wet mode or ddboutaT ,,, in dry mode.

In the case of dry mode ddboutaT ,,, can easily be calculated through the expression

da

dddbinaddbouta

C

QTT

,

,,,,,,

&

+= (dry mode) [°C] (4.49)

where

apdada cMC ,,, ⋅= & (dry mode) [W/°C] (4.50)

and

da

liq

dliq

da

M

M

MM

=

&

&

&& ,1

, [kg/s] (4.51)

where

=dQ& Design cooling capacity [W]

=daC , Design heat capacity flow of air [W/°C]

=daM ,& Design mass flow of air [kg/s]

=apc , Specific heat capacity of air [J/kg °C]


=

da

liq

M

M

&

&

Design mass flow relation of Liquid 1 and air [-]

The mass flow of Liquid 1 at the design point, dliq

M,1

& , is calculated as follows:

( )dcliqdwliqliqp

d

dliqTTc

QM

,,1,,11,

,1

1000

−

⋅=

&& [kg/s] (4.52)

where

=1,liqpc Specific heat capacity of Liquid 1 [J/kg °C]

60

If a wet mode is chosen, an iterative process must determine the value of dwboutaT ,,, . By

using a first guess of dwboutaT ,,, , it is possible to calculate outah , (equation 4.2) and eapc ,,

(equation 4.10) followed by aC (equation 4.17). The leaving air wet bulb temperature,

dwboutaT ,,, , can now be calculated by using equation 4.27. The iterative process is carried

out until convergence with required accuracy is reached.

The effective heat transfer coefficient-area product, deUA , , is given by the following

relationship:

ln

,T

QUA d

de∆

=&

[W/°C] (4.53)

where lnT∆ is the log-mean temperature difference defined as:

( ) ( )

−

−

−−−=∆

dinadcliq

doutadwliq

dinadcliqdoutadwliq

TT

TT

TTTTT

,,,,1

,,,,1

,,,,1,,,,1

ln

ln

[°C] (4.54)

in wet mode operation: dwboutadouta

TT,,,,,

= and dwbinadina

TT,,,,,

=

in dry mode operation: ddboutadouta

TT,,,,,

= and ddbinadina

TT,,,,,

=

Finally the design heat transfer coefficient-area product, dUA , can be calculated:

eap

ap

dedc

cUAUA

,,

,

,= [W/°C] (4.55)

The dUA -value is then used in equation 4.9 to calculate the UA -value.


61

4.3 Validation of the model

Validation of the effectiveness – NTU model in general has been discussed in section

2.2 Modelling heat and mass transfer in evaporative cooling towers, where published

material by Benton et al., (2002), Bourdouxhe et al., (1994) and Hernandez et al.,

(1994) was presented. The validation of the CTM developed within the work of this

thesis has been carried out with emphasis on the temperature of the chilled water

leaving the cooling tower, i.e. the supply cooling water. It is the same parameter which

was examined by Benton et al., (2002), Bourdouxhe et al., (1994) and Hernandez et al.,

(1994). It is the parameter which generally is the most important when applying a

cooling tower to a cooling system.

The validation has been carried out by comparing calculated data from the developed

model with relevant data from published sources and from measurements of the pilot

plant in this thesis. The published material is from Costelloe & Finn (2003) and Poppe

(1973) see Halasz (1999).

4.3.1 Validation by published data

The material from Costelloe & Finn, (2003) is based on experimental measurement data

from a test rig consisting of an open cooling tower, a primary and a secondary circuit

with an intermediate plate heat exchanger, similar to the system in figure 4.3. The

cooling tower in the test rig had a rated output of about 20 kW at design conditions. The

published data consist of 13 cases with temperatures of the primary and secondary

circuits respectively. The measurements was conducted at ambient wet bulb

temperatures in the range of 6,4 – 16,5°C, with primary approach temperatures between

0,9°C and 2,3°C and total approach temperatures between 2,2°C – 4,3°C.

The material from Poppe (1973) see Halasz (1999), which was used as a reference for

validating Halasz (1999) own model, comes from numerical calculations of “rather

accurate” differential equations describing an evaporative cooling tower (originally

published by Poppe, 1973). In this case the cooling tower is working with only a

primary circuit, hence there is only one approach and range temperature for each case.

The published material encompasses 32 values with approach temperatures from 2 to

22°C and range temperatures from 4 to 30°C. The upper values of these approach and

range temperatures are much higher than in ordinary comfort cooling applications,

where low approach and range temperatures are desirable. The value

da

liq

M

M

&

&

is in the

range from 0,25 to 8, which is a wider range than in usual applications. Typical values

are between 0,5 to 2.

62

Costelloe & Finn, (2003)

In table 4.4 the specifications of the different conditions at each case together with

findings by Costelloe & Finn, (2003) compared with results from the CTM are

presented.

Table 4.4 Specifications of the different conditions at each case together with findings

by Costelloe & Finn, (2003) compared with results from the cooling tower

model in this thesis.

Measured data by Costelloe and Finn, (2003) Results from

cooling tower model

in this thesis, i.e. the

CTM

Deviation between

results from model

and measured data

Case

no.

Load

(kW)

Ambient

wet bulb

temp.

Primary

supply

temp.

Second.

supply

temp.

Second.

return

temp.

Primary

supply

temp.

Second.

supply

temp.

Primary

supply

temp. dev.

Second.

supply

temp. dev.

1 24 6,4 8,7 10,7 14,53 8,57 10,06 -0,13 -0,64

2 24 8,9 10,8 12,5 16,33 10,70 12,10 -0,10 -0,40

3 24 9,2 11,1 12,8 16,63 10,97 12,38 -0,13 -0,42

4 24 11,1 12,8 14,6 18,43 12,71 14,14 -0,09 -0,46

5 20 8,4 9,8 11,3 14,49 9,91 11,06 0,11 -0,24

6 20 9,2 10,6 12,1 15,29 10,66 11,81 0,06 -0,29

7 20 10,2 11,3 12,9 16,09 11,57 12,70 0,27 -0,20

8 20 12,5 13,5 15,1 18,29 13,68 14,83 0,18 -0,27

9 20 16,5 17,4 19,0 22,19 17,43 18,62 0,03 -0,38

10 15 8,7 9,7 10,9 13,29 9,84 10,70 0,14 -0,20

11 15 9,3 10,6 11,7 14,09 10,45 11,36 -0,15 -0,34

12 15 9,7 11,1 12,3 14,69 10,89 11,84 -0,21 -0,46

13 15 10,6 11,6 12,8 15,19 11,65 12,54 0,05 -0,26

The temperature deviations from comparing measured data by Costelloe & Finn (2003)

with results from the cooling tower model in this thesis, the CTM, are shown in figure

4.9. The zero-level in the figure represents the data from Costelloe & Finn (2003), i.e. if

a bar is on the negative scale the calculated value from the CTM is lower than the

corresponding measured value by Costelloe & Finn (2003) and vice versa.


63

Figure 4.9 Temperature deviations from comparing measured data by Costelloe &

Finn (2003) with results from the cooling tower model in this thesis. The

zero-level represents the data from Costelloe & Finn (2003). Description

of the different cases is found in table 4.4,

The mean value, maximum value, standard deviation and the 95% confidence interval

of the deviations in table 4.4 are presented in table 4.5.

Table 4.5 Accuracy of the CTM in this thesis compared with measured data from

Costelloe & Finn (2003) based on values from table 4.4.

Deviation [°C] Primary dev. Secondary dev.

Mean value 0,00 -0,35

Maximum value 0,27 -0,64


95% Confidence interval ±0,08 ±0,07

When comparing results from the cooling tower model with data from Costelloe & Finn

(2003) there is a good accordance. The mean value of the total temperature deviation is

-0,35°C with a standard deviation of 0,12°C. Biggest total deviation is -0,64°C with all

other deviations below -0,46°C.

-0,80

-0,60

-0,40

-0,20

0,00

0,20

0,40

1 2 3 4 5 6 7 8 9 10 11 12 13

Case no.

Te

mp

era

ture

de

via

tio

n (

°C)

Primary temp dev. Second. temp. dev.

64

Halasz, (1999)

In table 4.6 the conditions at each case together with findings by Poppe (1973), see

Halasz (1999), are compared with results from the CTM.

Table 4.6 Conditions at each case together with data from Poppe (1973), see Halasz

(1999), compared with results from the cooling tower model in this

thesis. (WBT = Wet Bulb Temperature of ambient air)

Input data by Halasz, (1999) Output data,

Halasz, (1999)

Results from cooling tower model

in this thesis, i.e. the CTM

Case

no.

Return

temp.

(°C)

WBT

In

(°C)

Range

(°C)

Approach

(°C)

ma/mv Supply

temp.

(°C)

WBT

Out

(°C)

Supply

temp

(°C)

WBT

Out

(°C)

Supply

temp

dev.

WBT

out

dev

0.1 30 4 4 22 0,25 26 27,37 26,45 24,90 0,45 2,47

0.2 30 4 4 22 0,30 26 24,83 26,30 23,08 0,30 1,75

0.3 30 8 4 18 0,30 26 26,77 26,33 24,88 0,33 1,89

1.1 34 12 4 18 0,20 26 33,82 - - - -

1.2 34 12 4 18 0,25 30 30,92 30,35 29,09 0,35 1,84

1.3 34 12 4 18 0,30 30 28,68 30,22 27,32 0,22 1,37

1.4 34 16 4 14 0,30 30 30,88 30,25 29,48 0,25 1,40

2.1 34 20 4 10 0,30 30 32,92 30,33 31,65 0,33 1,27

2.2 34 20 4 10 0,35 30 31,46 30,20 30,56 0,20 0,90

2.3 34 20 4 10 0,40 30 30,31 30,13 29,60 0,13 0,71

2.4 34 24 4 6 0,40 30 33,02 30,21 32,25 0,21 0,77

3.1 34 28 4 2 0,80 30 32,27 30,01 31,98 0,01 0,29

3.2 34 28 4 2 1,00 30 31,50 30,00 31,22 0,00 0,28

3.3 34 28 4 2 1,20 30 30,97 30,00 30,70 0,00 0,28

4.1 34 12 10 12 0,50 30 33,70 - - - -

4.2 34 12 10 12 0,80 24 27,81 24,14 27,08 0,14 0,73

4.3 34 12 10 12 1,00 24 25,34 24,04 24,82 0,04 0,52

4.4 34 16 10 8 1,00 24 27,95 24,01 27,31 0,01 0,64

5.1 34 20 10 4 1,00 24 30,27 24,00 29,87 0,00 0,40

5.2 34 20 10 4 1,50 24 27,32 24,00 26,98 0,00 0,34

5.3 34 20 10 4 2,00 24 25,71 24,00 25,40 0,00 0,31

6.1 40 12 20 8 1,50 24 28,60 20,00 27,99 0,00 0,61

6.2 40 12 20 8 2,00 20 25,37 20,00 24,86 0,00 0,52

6.3 40 12 20 8 3,00 20 21,55 20,00 21,27 0,00 0,28

6.4 40 16 20 4 3,00 20 24,76 20,00 24,10 0,00 0,66

7.1 40 18 20 2 3,00 20 25,81 20,00 25,51 0,00 0,30

7.2 40 18 20 2 5,00 20 23,01 20,00 22,76 0,00 0,25

7.3 40 18 20 2 8,00 20 21,28 20,00 21,07 0,00 0,21

8.1 54 12 30 12 1,00 20 39,84 24,00 39,66 0,00 0,19

8.2 54 12 30 12 1,50 24 34,18 24,00 33,28 0,00 0,90

8.3 54 12 30 12 2,00 24 30,38 24,00 29,42 0,00 0,97

8.4 54 16 30 8 2,00 24 32,51 24,00 31,49 0,00 1,02

The temperature deviations in table 4.6 are shown in figure 4.10. The zero-level in the

figure represents the data from Halasz, i.e. if a bar is on the negative scale, the

calculated value from the cooling tower model is lower than the corresponding

reference data and vice versa.


65

Figure 4.10 Temperature deviation from comparing data from Poppe (1973), see

Halasz (1999), with results from the cooling tower model in this thesis.

Data is from table 4.6. The zero-level represents the data from Poppe.


of the temperature deviations in table 4.6 are presented in table 4.7.

-2,5 -2,0 -1,5 -1,0 -0,5 0,0 0,5

0.1

0.2

0.3

1.1

1.2

1.3

1.4

2.1

2.2

2.3

2.4

3.1

3.2

3.3

4.1

4.2

4.3

4.4

5.1

5.2

5.3

6.1

6.2

6.3

6.4

7.1

7.2

7.3

8.1

8.2

8.3

8.4

Ca

se n

o.

Temperature deviation (°C)

Supply temp. dev. Air wetbulb out dev

66

Table 4.7 Accuracy of the cooling tower model in this thesis method compared

with data from Poppe (1973), see Halasz (1999), based on data from

table 4.6.

Deviation [°C] Supply water

temperature

WBT of

discharge

air

Mean value 0,10 -0,80

Maximum value 0,45 -2,47


95% Confidence interval ±0,08 ±0,32

WBT = Wet Bulb Temperature

When comparing results from the cooling tower model with data from Poppe (1973),

see Halasz (1999), there is a good accordance. The mean value of the supply water

temperature deviation is 0,10°C with a standard deviation of 0,14°C. Biggest total

deviation is 0,45°C for the case when the cooling tower range was 4°C and the approach

was 22°C. An approach of that magnitude is exceptionally unusual in comfort cooling

applications.

Comparing with the data from section 2.2 Modelling heat and mass transfer in

evaporative cooling towers, table 2.1 and 2.2, where published data of accuracy tests on

the effectiveness – NTU model per se is presented, the results in table 4.5 and 4.7 are

approximately equal or better.

With the results in this section together with findings in section 2.2 Modelling heat and

mass transfer in evaporative cooling towers, there can be concluded that the cooling

tower model developed in this thesis, the CTM, has a good accuracy concerning the

temperature of the supply water when comparing with published data.

4.3.2 Validation by data from pilot plant

The CTM is also validated against measurement data from the pilot plant at

Kvarnberget in Göteborg. The data is from May 1 – August 31, 2007. In the previous

section the CTM is validated against a number of single operational conditions. In this

case the CTM is validated against measured data from continuous operation when

applied in IDA ICE. A presentation of the simulation tool IDA ICE is made in section

5.1 Simulation tool and description of the monitoring of the pilot plant at Kvarnberget in

Göteborg is presented in chapter 6 Monitoring of free cooling system with evaporative

cooling tower.

Figure 4.11 shows an example period during July 24 – 27, 2007 with measured supply

liquid temperature versus calculated temperature. Figure 4.12 shows all samples during

the period May 1 – August 31 when there was wet operation active, presented as

measured versus calculated supply liquid temperature.


67

10

11

12

13

14

15

16

17

18

19

20

00:00 12:00 00:00 12:00 00:00 12:00 00:00 12:00 00:00

Te

mp

era

ture

[°C

]

Time

Comparison between measured and calculated dataKvarnberget, Göteborg July 24 - 27, 2007

Measured Calculated

Figure 4.11 Comparison between measured and calculated data during an example

period between July 24 – July 27, 2007. Measured data is from

Kvarnberget, Göteborg.

10

12

14

16

18

20

22

24

10 12 14 16 18 20 22 24

Ca

lcu

late

d d

ata

Measured data

Comparison between measured and calculated dataData during wet operation, Kvarnberget Göteborg , 2007

Figure 4.12 Comparison between measured and calculated data. Measured data is

from Kvarnberget, Göteborg. Total number of measurement samples

during wet operation between May 1 – August 31, 2007

68

Table 4.8 Accuracy of the cooling tower model in this thesis method compared

with data from measurements from Kvarnberget, Göteborg, 2007

Deviation [°C] Supply water

temperature

Mean value -0,51

Maximum value -0,98

Standard deviation 0,14

95% Confidence interval ±0,04


of the deviations indicated in figure 4.12 are presented in table 4.8.

As can be seen in figure 4.11 and more so in figure 4.12 there is a deviation between

calculated and measured data. The mean value of the deviation is -0,51 °C (table 4.8).

Part of the deviation can be explained by the fact that the evaporative free cooler in the

pilot plant at Kvarnberget, Göteborg, is a mixture of an adiabatic cooler and an ordinary

cooling tower. When studying the evaporative free cooler in active operation it was

clear that the flanges was only partly wetted due to the direction of the spray water,

which was coming from underneath the cooler. This makes the evaporative free cooler

partly an adiabatic cooler, partly an ordinary cooling tower. The CTM, on the other

hand, is a model of an ordinary cooling tower. An adiabatic cooler cannot cool a

circulating liquid as efficient as a evaporative cooling tower, everything else equal,

because an adiabatic cooler is only using sensible cooling, although the ambient air is

evaporatively cooled.

69

5 Simulation of a free cooling system with evaporative cooling tower

In this chapter, the basis and the prerequisites are presented for the thermal

analysis of a cooling system with an evaporative cooling tower. The analysis is

made through simulations of a section of an office floor. In that section, comprising

ten office rooms, one room is selected to be the space where the resulting thermal

climate is logged and analysed. The primary thermal climate variable is the room

dry bulb air temperature. The simulations are first made on a base case building

and HVAC-system followed by a parameter variation analysis.

In section 5.1, the selected simulation tool is presented. In section 5.2 is the

prerequisites at base case described followed by section 5.3 where the simulation

methodology with two different simulation cases is described. Section 5.4 explains

the parameter analysis methodology including a discussion and motivation of the

considered parameter alterations.

5.1 Simulation tool

There are a number of internationally well known building simulation tools that would be suited for the required analysis in this project, e.g. TRNSYS, DOE or Energy Plus. However, another tool was selected for the task; IDA Indoor Climate and Energy (ICE). It is an advanced tool for simulation of thermal comfort, indoor air quality and energy usage in buildings. It has more than 400 commercial licenses and about 1000 registered users on a free web based application called IDA Room (a limited version of IDA ICE). The users are mostly HVAC designers but also educators and researchers. Most of the licenses are in Sweden and in Scandinavia. The first version was released in May 1998 and the latest version, 4.0, was released in June 2009. Using IDA ICE would facilitate the dissemination of the results from this project in Scandinavia. Tools like TRNSYS, DOE, Energy Plus, or any other foreign advanced building simulation tool, have a very limited number of users within Scandinavia. IDA ICE has some unique characteristics as a building performance simulation tool, especially useful in this project.

1) All mathematical models are described in terms of equations in a formal language called Neutral Model Format (NMF). This makes it easy to replace and upgrade program modules. The end user can also view all NMF-models which makes the program open and transparent. In IDA ICE mathematical models may be re-connected arbitrarily by the end user.

2) Advanced users can use IDA SE (IDA Simulation Environment) in conjunction with IDA ICE to develop and implement new models and tailored user interfaces according to their own needs. The cooling tower developed model in this thesis, the CTM, is written in NMF-language and implemented through IDA SE into IDA ICE. A complete printout of the cooling tower model in NMF source code is shown in Appendix A. It should be noted that no changes in the source code of IDA ICE is required when adding a new model.

70

3) The full system of equations, describing the building and the HVAC system, is solved with a general purpose, variable time step solver called IDA Solver. The time resolution can vary from the maximum time step set by the user, usually 30 – 60 minutes, down to fractions of minutes. The determination of the length of each time step is based on the difference between a predicted and the actual solution at the current time step.

IDA ICE is designed to handle:

� Multiple zone dynamic and simultaneous bulk air flow and heat balance, including specific contributions from sun, occupants, equipment, lights, ventilation, heating and cooling devices, surface transmissions, air leakage, large vertical openings, cold bridges and internal objects such as furniture.

� Different heating and cooling devices in each zone, e.g. hydronic or electric heating baseboard panels, floor heating, chilled ceiling panels and passive or active (ventilated) chilled beams.

� Solar influx through windows with full account for local shading devices as well as surrounding buildings and other objects. Control of shading devices dependent on solar influx and/or wind. Detailed 3D direct and diffuse shading calculations.

� Air and surface temperatures � Operating temperature at multiple arbitrary occupant locations, e.g., in the

proximity of hot or cold surfaces. Full non-linear Stephan-Bolzmann radiation with view factors is used to calculate radiation exchange between surfaces.

� Non-linear correlations for surface film coefficients

� Directed operative temperatures for estimation of asymmetric comfort conditions � Comfort indices, PPD and PMV, at multiple arbitrary occupant locations

� Daylight level at an arbitrary room location � Zone CO2 and moisture levels, both of which may be used for control of VAV

system air flow

� A model for vertical air gradient calculation is available

� Wind and buoyancy driven airflows through leaks and large vertical openings via a fully integrated airflow network model. This enables study of, e.g., temporarily open windows or doors between rooms.

� Airflow, temperature, moisture, CO2 and pressure at arbitrary locations of the air handling and distribution systems

� Power levels and temperatures for primary and secondary system components

� Total energy cost based on time-dependent prices as well detailed energy accounts As a building simulation tool with the above mentioned properties, IDA ICE can be considered as one of the most advanced building simulation tools on the market. The original development of IDA ICE was requested, specified and partly financed by a group of thirty leading Scandinavian companies within the building sector. The mathematical models were originally developed at the Royal Institute of Technology in Stockholm (KTH) and at Helsinki University of Technology. The company Bris Data, which later changed the name to Equa, developed IDA ICE. All models in the simulation tool are available as NMF source code.


71

The models are not tailored to Scandinavian needs but seek to capture the international state-of-the-art in building performance modelling. Whenever appropriate, models recommended by ASHRAE have been used. IDA ICE has been validated in several ways.

� A large number of inter-model comparisons have been made against the BRIS program (Brown 1990), which in turn has been extensively validated against measurements of thermal performance on real buildings.

� An extensive empirical validation exercise based on test cells has been carried out within the IEA Solar Heating and Cooling Task 22 (Guyon et al 1999).

� Validation with IEA Building Energy Simulation Test (BESTEST) (Ascherman 2000)

� Validation according to CEN 13791 ”Thermal Performance of Buildings – Calculation of Internal Temperatures of a Room in Summer Without Mechanical Cooling – General Criteria and Validation Procedures”, (Kropf and Zweifel 2002)

� Validation focused on radiant heating and cooling systems with RADTEST has been carried out within the IEA Solar Heating and Cooling Task 22, (Ascherman and Zweifel 2003)

5.2 Prerequisites at base case

The intention with the base case building and its technical systems is that it will reflect typical conditions of a normal office building, the activity in it and the standard and size of the technical systems. Hence, the following prerequisites are important to notice:

� The base case building does not meet the current Swedish building code concerning insulation. This is the case with the bulk part of commercial buildings in Sweden together with the rest of the building stock in the tertiary sector in the northern parts of Europe.

� The base case building is not equipped with smart, automatic and efficient solar shadings.

� The lighting is not the most energy efficient on the market, and there is no

automatic control of the lighting regarding daylight levels or presence of people in the rooms.

� The appliances in the office rooms, e.g. computer and screen, have no energy saving functions activated.

� The ventilation system has no automatic control regarding operating hours, night cooling or demand control of volume flow through variable frequency drive or other device.

� The individual temperature control device in the rooms, controlling the chilled beams, has no built-in “intelligence” with variable set point temperatures between day and night or any fuzzy-control. It is a basic PI-controller to keep the air temperature in the room constant.

If all these prerequisites were actually a part of the base case building, i.e. omit all no or not above, it would most likely lead to a lower indoor temperature during hot and sunny

72

weather, and lower energy use, than presented in this thesis. On the other hand, the results would then not reflect the conditions of a common real life commercial office building. A complementary discussion and motivation of most of the prerequisites is found in section 5.4.2 Discussion and motivation of parameter alterations. 5.2.1 Model building

The base case building consists of an imaginary multi storey non-residential building. The analysed room is situated in a floor somewhere in the middle section of the building. The floor contains ten rooms laid out as in a normal office building, se figure 5.1. The office building is located in a medium climate of the northern parts of Europe, in this case chosen to be London Gatwick. The analysed room, marked with dashed lines in the figure, has two adjacent rooms on each side and five rooms on the opposite side of the corridor and the middle section. In a real life building the middle section is normally containing service rooms such as copying and printer rooms, toilets and archives.

Figure 5.1 Original layout of the office floor The analysed room is placed in a normal office environment with adjacent rooms and a corridor to take in to account the variations of the coincident heat flow and airflow with the corridor as well as the outdoor environment. It is assumed that the doors between all the rooms and the corridor always are open and that the two corridors are connected so that airflow is possible due to differences in density or pressure between all the zones. The layout of the office floor in the base case building in figure 5.1 is simplified when modelled in IDA ICE. The simplification consists of the following:

20 m Middle section

Analysed room

Adjacent office rooms

Corridors

15 m

N


73

� The adjacent row of five office rooms on the opposite side of the floor is lumped into one room.

� The neighbouring two rooms on each side of the analysed room are lumped together respectively.

This is made of mainly two reasons. The first reason is that the exact modelling of the adjacent rooms is not necessary since their task is to exchange air and heat with the corridor so that the resulting air temperature of the corridor will be close to a real life situation. The second one is because the simulation time is linearly proportional to the number of zones in IDA ICE. Since a lot of simulations are required within this work, a short simulation time is preferred. The simplified base case building model with the office floor is shown in figure 5.2. The adjacent rooms are now double the size of the analysed room and the room on the opposite side of the corridor is five times the analysed room in size. The total size and floor area, the doorway area of the rooms in the simplified layout is however exactly the same as in the original layout. The corridor is lumped into one rectangular room with the same floor area as the two original corridors. The middle section, surrounded by the corridors in the original building, is modelled as one single imaginary inner wall placed in the corridor to simulate the thermal mass of the middle section. The wall is imaginary in the sense that it does not take up any space or hinder any air movements. The area of the single inner wall is equal to the area of the walls and slabs in the original middle section between the corridors. Figure 5.2 Simplified layout of the office floor as in IDA ICE The physical size of the rooms and the corridor is remained constant throughout the whole analysis. The construction of the different parts in the building, such as floor, ceiling, outer wall, windows and inner walls, are always the same in all the rooms. If a change is made in one or more construction parts due to parameter variation analysis, this change is made i all the rooms in the floor.

12 m

Lumped middle section to a single inner wall equivalent in area and thermal weight to middle section in figure 5.1.

Analysed room

Lumped office rooms

Lumped corridor

74

The floor and ceiling slab are made of reinforced concrete. The floor has an additional layer of linoleum carpet. There is a suspended false ceiling with acoustic tiles mounted to the slab. The inner walls are made of wood framework with single gypsum board on each side with insulation between. The outer wall is made of lightweight concrete and insulation. A top layer of plaster is added on each side. This construction type is normally categorized as medium heavy concerning thermal capacity. The U-value of the outer wall is chosen to 0,4 W/m2 K which is in accordance to Swedish building code in the mid 1970’s. The windows have double pane clear glass type with a U-value of 2,5 W/m2 K. 5.2.2 Analysed room

The analysed room is where the indoor dry bulb temperature and other variables are registered. The analysed room is 12 m2 with a room height of 2,4 m between floor slab and false ceiling. It has one external wall, including a window, facing south. The size of the window is adjusted to the required solar heat gain in the room; see section 5.2.3 Internal heat and moisture generation. The remaining three walls are inner walls facing adjacent rooms and the corridor. The corridor wall has a door, which is assumed to be always open. For details concerning construction, see section 5.2.1 Model building above. 5.2.3 Internal heat and moisture generation

The internal heat generation comprises heat from humans, lighting, office equipment and solar radiation through the window. In the analysed room, one person is present together with lighting fitting and office equipment. The adjacent and opposite office rooms have the same specific internal heat generation as the analysed room, for humans as well as equipment and lighting. Working hours is 08 – 17 (8 am. to 5 pm.), Monday to Friday all year round. Time for holidays is excluded in the simulations to avoid possible coincidence with hot and humid weather conditions and holidays, i.e. low internal heat generation. The person has in the base case an activity level equal to 1,2 Met and a clothing equal to 0,9 Clo during 1st of May to 30th of august. The rest of the year the clothing is equal to 1,1 Clo to reflect that people have more clothes during colder weather conditions. With given values of activity and clothing the generated sensible heat from a person is dependent of the indoor air temperature, mean radiation temperature and the indoor air velocity. If these two temperatures are equal to 22°C, the sensible heat is about 89 W in the summer (1st of May to 30th of august) and about 79 W the rest of the year. The indoor air velocity is set to a constant value of 0,1 m/s. The moisture generation is calculated from the values of the activity level (1,2 Met) and the partial pressure of the water vapour in the room air. The partial pressure of the water vapour is determined from the air temperature and the relative humidity in the room. The heat and moisture generation from persons is in IDA ICE calculated according to Fanger (1972). The heat and moisture generation is therefore dependent on for example, the level of activity, the clothing and the indoor air temperature as well as the mean radiant temperature. During time other than working hours, the internal heat and moisture generation from persons is zero.


75

Neither the lighting nor the office equipment is specified in detail; instead, it is given as a specific heat gain related to the floor area. The lighting is set to a specific heat gain of 12 W/m2, which equals a total heat generation of 144 W in the office rooms. In the corridor, the heat gain from lighting is set to 6 W/m2. The office equipment has a specific heat gain of 12 W/m2, which equals a total of 144 W for the room. In the corridor, no equipment load is present. In the heat transfer process the lighting is set to 30% convection – 70% radiation and office equipment to 80% convection – 20% radiation. During time other than working hours the internal heat generation from lighting in office rooms is set to 0 W/m2. Office equipment is to some extent often on during non working hours and is therefore set to 6 W/m2. In the corridor, the lighting is set to 1 W/m2 during non working hours to simulate that some lighting always is on. The sensible heat generation from humans, lighting and equipment with the numbers above is 364 W which equals a specific heat gain of 30,3 W/m2. The maximum total sensible heat gain during daytime is set to 60 W/m2 and the remaining load, about 30 W/m2 comes from solar radiation through a window in each office room. The window in each room is adjusted in size so that the maximum total sensible heat generation will be 60 W/m2. The adjustment of window size to tune the maximum solar radiation to 30 W/m2 is made during a heat wave simulation at design day conditions, where the sky is constantly clear and the insolation is at maximum. For explanation of heat wave simulation, see section 5.3.1 Heat wave simulation. The total solar radiation at 30 W/m2 is the maximum value during a day. The contribution to the total solar radiation comes from direct beam, diffuse and ground reflected radiation. The total radiation varies throughout the day with a maximum at noon. It is stressed that the maximum total sensible heat gain of 60 W/m2 includes the contribution from the solar radiation in the room. Expressing the solar radiation to the room, and the total heat gain, in specific numbers leads to a more general approach when performing the simulations. Setting the solar design heat gain in the room to 30 W/m2 implies practically almost infinite combinations of the following parameters:

� Number of window panes

� Physical size of the window

� Possible solar protective coating on one or more panes � Type of coating

� Possible indoor and/or outdoor solar shading device

� Type of solar shading devices

76

5.2.4 Cooling tower

The cooling tower, which is connected to the air-cooling coil in the air handling unit and the local cooling beams in the rooms, figure 5.4, is selected to have the following features:

� Closed circuit with steel tube bundles with smooth external wall tubes suitable for external water spraying; see schematic sketch in figure 5.3. The choice of a closed circuit tower versus an open type of cooling tower has no particular affect on the analysed variable (indoor temperature). The closed circuit has an extra pump for the spray water circulation which adds to the total energy usage of the tower. Normally the closed circuit type requires a bigger volume of the tower compared to an open type. A closed circuit tower is used more often in building related systems than an open circuit cooling tower. The reason for this emanates mainly from a desire to reduce the maintenance cost for water treatment.

� Counter flow configuration. This configuration is the most common and it also gives the lowest leaving temperature of the cooled fluid compared to a cross flow configuration with all other parameters equal.

� Wet (evaporative) type of tower. This is an obvious choice since this thesis deals with evaporative cooling towers only.

� An intermediate closed circuit with freeze protected coolant with a heat exchanger between the primary and secondary fluids, figure 5.4. This choice origin from the possible occurrence of cooling demand in a building during wintertime when the risk of freezing of the coolant is high. The heat exchanger is placed indoors where the air temperature never reaches freezing degrees. The coolant in the primary (intermediate) circuit is a 40% propylene glycol-water solution while the liquid in the secondary circuit is merely water.

Figure 5.3 Schematic illustration of a closed type cooling tower used in the

simulations.

Coolant in closed loop system

Ambient air

Fan

Spray water basin

Drift eliminator

Spray nozzles


77

Figure 5.4 The cooling tower model and its connection with the rest of a possible

cooling system. Figure 5.5 Temperature diagram of a counter flow cooling tower with intermediate

circuit and visualisation of definition of design parameters. The design parameters, explained in figure 5.5, which influences the performance of the cooling tower is chosen as follows:

Closed type cooling tower

Heat exchanger

Hydronic room cooling devices, e.g. fan coils or cooling beams

Secondary water based circuit

Intermediate freeze-protected circuit

air heating coil

supply air fan air cooling coil

Supply air to zones

Part of the AHU

System boundary for the cooling tower model


Wet bulb (Ambient air) temp. in

Area

Temperature

Liquid in secondary circuit (water)

Moist air

Secondary approach

Primary approach

Primary range

Secondary range

Total approach

Cold side

Return side

Supply side

Warm side

Wet bulb temp. out

Outdoor Indoor

78

� Primary approach: The approach in the primary circuit is set to 4°C. The value of the approach is critical concerning economical decisions in the design phase. The economical aspect covers mainly short term aspects such as investment cost due to increasing size of the cooling tower with decreasing values of the approach. In the literature the lower limit for the design approach, based on economic reasons, is almost always 3°C (approx. 5°F). A further discussion about this topic is presented in section 5.4.2 Discussion and motivation of

parameter alterations. � Secondary approach: The approach in the secondary circuit is set to 1,5°C. The

heat exchanger is a plate and frame type which normally offers excellent efficiency, i.e. low values on approach. Lower approach than about 1°C is in the literature, e.g. de Saulles (1996), not advisable due to too high investment costs.

� Primary Range: The range in the primary circuit is chosen to the same value as in the secondary circuit, i.e. 3°C. This leads to equal mass flows in the primary and secondary circuit at design conditions.

� Secondary Range: The secondary range is set equal to 3°C. This temperature difference is very common in comfort cooling systems with chilled ceilings or chilled beams.

� Mass flow quota of liquid and air (a

l

M

M

&

&

): This relation primarily state the

value of the mass flow of air since the mass flow of the liquid is given from the design cooling capacity and the range of the primary liquid. The value of the mass flow quota is normally in the range between 0,5 and 2. In the base case the mass flow quota is set to 1,0. According to Söylemez, (2003) the optimum flow quota is about 1 for mean temperature of cooling tower water at 22 – 27°C.

� Cooling capacity: The cooling capacity is set equal to the total design cooling capacity of the cooling beams together with design cooling capacity of the air cooler in the air handling unit. With 50 W/m2 in cooling capacity of the cooling beams and a specific airflow of 1,25 l/s m2 the total design cooling capacity of the tower is 7,69 kW.

� Design conditions: The ambient design conditions are set to 25°C dry bulb temperature and a relative humidity of 60% RH. This equals a wet bulb temperature of 19,5°C, which is very close to the ASHRAE design wet bulb temperature (0,4%) of London Gatwick that is 19,3°C (ASHRAE 2001a).

5.2.5 HVAC system

The HVAC system consists of a ventilation system, a heating system and a cooling system. The basic configuration is a water-air system, i.e. room units on hydronic circuits provide heating and cooling and ventilation is provided by a CAV-system to meet the requirements of indoor air quality.


79

Ventilation system

The ventilation system is a basic CAV-system with constant air volume flow and constant supply temperature. The system is supplied with air from an Air Handling Unit (AHU) comprising air supply and discharge fans, air-to-air heat exchanger for heat recovery, air heating coil and air-cooling coil see figure 5.6. The air-cooling coil is supplied with chilled water from the cooling tower, figure 5.4. Figure 5.6 Basic scheme of the Air Handling Unit (AHU) in the base case In the office floor, the office rooms are supplied with air at a rate equivalent to 1,25 l/s m2 and the supply air has a set point temperature of 17°C. The air flow rate is maintained constant throughout the year, only during workdays though. The actual supply air temperature is however not constant during warm weather conditions, due to the natural limitations of the cooling tower capacity to cool the water. In the office rooms, supply air diffusion is of mixing ventilation type. The supply air in the office rooms is transported to the corridor where the exhaust air is transported from the corridor. The ventilation system is operating constantly during workdays (Monday 06:00 - Friday 18:00) all year and is shut down during weekends. The operating scheme is not optimal concerning minimal energy usage for the ventilation system (heating and electricity). This is however not an issue in this thesis since the energy and temperature performance of the cooling system, i.e. the cooling tower, is in focus. The primary objective of the ventilation system in this case is to remove odour and pollutants, thus a moderate airflow. A secondary objective is to provide cool air to supplement the cooling from the chilled beams. About 20% of the design cooling capacity in the room is supplied from the ventilation at design conditions. The operating schedule therefore includes 24-hour operation during workdays to utilize nighttime cooling. No control function is added to switch nighttime cooling on or off based on some criteria.

air heating coil

return air fan

supply air fan air cooling coil

Air-to-air heat exchanger

supply air to building

return air from building

Ambient air

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Heating system

A heating system is added to the office floor to provide a minimum indoor air temperature during colder periods. Ordinary hydronic baseboard heaters placed under the windows supply the heating to the office rooms. The heat capacity of the heaters will be sufficient to keep the set point temperature at the coldest day. The baseboard heaters are controlled by an ordinary proportional thermostat with a proportional band at 2°C and a set point temperature of 20°C. The corridor has no heating device. The primary heating source is assumed to be district heating. Cooling system

The cooling system consists of a cooling tower, as described in section 5.2.4 Cooling

tower, a distribution system of pipe work, an air cooling coil in the AHU and room cooling devices, see figure 5.4. The rooms cooling devices is in this case active cooling beams. An active cooling beam combines cooling and ventilation functions into a single unit. Heating is also possible with this type of beam, but is not included in this thesis. To minimize the simulation time, every office room has only one cooling beam. In a real life situation an office room may have two or several cooling beams depending on room size, cooling loads and other design factors. To lump cooling beams into one single does not affect the simulation result since the sum of a number of cooling beams has the same cooling capacity and characteristics as one lumped beam. In the heat transfer process, a cooling beam uses nearly 100% convective heat to remove excess heat in contrast to a chilled ceiling or a cooling panel, which use approximately 50% radiation and 50% convection. For a more in depth discussion about room cooling devices in general see Nilsson (ed) (2003) and chilled ceilings and chilled beams i detail see Novoselac & Srebric (2002), Mott et al. (2001) and Sodec (1999) The sizing of the cooling beams in all office rooms is based on the following:

� The design room air temperature is 25°C

� The design mean cooling water temperature is 17,5°C with a supply temperature of 16°C and a return temperature of 19°C

� The design temperature difference between the room air temperature and the mean cooling water temperature is 7,5°C

� The design secondary temperature range is 3°C (difference between cooling water supply and return temperature), see figure 5.5.

� The design cooling capacity is 600 W for an office room at 12m2, equal to 50 W/m2.

In a room with the size of 12 m2, as in the base case, two modern active cooling beams with the length of 1,8 m, will suffice providing the design cooling capacity. The cooling beam in each room is controlled by a PI-controller with a set point temperature at 22°C. The cooling system will be operating 24 hours a day all year to ensure cooling whenever the indoor air temperature exceeds 22°C, i.e. set point temperature.


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5.3 Simulation methodology

One objective with this thesis is to investigate the indoor climate when a building is cooled solely by a cooling tower. The primary variables to be analysed is therefore dry bulb temperature and relative humidity of the indoor air. These variables are chosen simply because they are widely used as indicators of indoor thermal climate and they are also easy to measure. In section 5.4 Parameter analysis methodology, different sets of parameters that will be used in the sensitivity analysis are discussed. The analysis of the primary variables for each set of parameters is made in two different ways. The first way is to make simulation runs with the building and the cooling system exposed to a heat wave with a duration of five working days. This is a normal way of making simulations in the design process to either calculate the resulting indoor temperature or the necessary cooling capacity to meet the indoor temperature criterion. The other way of making the parameter analysis is to make full year simulations to examine the variation and duration of the indoor temperature during a normal year. The normal year in this case is a Typical Meteorological Year (TMY-year), issued by ASHRAE (2001b). Secondary variables such as energy usage and COP for the cooling tower, system temperatures etc. will be logged, however only with full year simulations and for a limited set of parameter variations. 5.3.1 Heat wave simulation

Testing and sizing the cooling system together with the building against a simulated heat wave is a design strategy, which is normal among HVAC consultants. It is therefore obvious that this design strategy should be applied in this thesis to evaluate the performance of the cooling system and building together. The heat wave can either be chosen from real meteorological data or created by a synthetic 24 hour periodic sinus wave of ambient temperature. In this thesis the latter is chosen since it is the most commonly used during the design phase. The duration of the heat wave is chosen to five days, Monday to Friday, to coincide with the working days and consequently maximum cooling loads, and with a 24-hour periodic outdoor temperature curve. The maximum and minimum outdoor dry and wet bulb temperature is taken from ASHRAE (2001a). Design conditions are chosen based on the wet bulb temperature together with the mean coincident dry bulb temperature. In normal design of an air conditioning system the design value is the dry bulb temperature. In this case it is preferred to use the wet bulb temperature as the design value since the performance of the cooling tower normally has a greater effect on the indoor air temperature than the design dry bulb temperature has. The 0,4% value is chosen as design value. The 0,4% value represent a wet bulb temperature value that is exceeded on average by 0,4% of the total number of hours in a year (8760 h/year). 0,4% of a year equals 35 h/year. 5.3.2 Full year simulation

A full year simulation, together with the heat wave simulation, gives a more complete picture of how the cooling system and the indoor parameters will perform. The full year

82

simulation is made with climate data from a Typical Meteorological Year (TMY), issued by ASHRAE (2001b). A TMY consists of twelve typical Meteorological Months’ (TMM) selected from the calendar months in a weather database comprising a period of normally 10-30 years. For example, January of 1983 may be selected as the first TMM, February of 1989 as the second TMM, and so on. All the twelve selected months will then be combined to form the TMY. The TMY’s used in this thesis are based on climate data for the years between 1982 and 1993. Selection criteria of a TMM are based on statistical analysis and evaluation of four weather parameters: Global Solar Radiation, Dry Bulb Temperature, Dew Point Temperature and Wind Speed. From these weather parameter a set of nine indices are selected and given different weightings. Of the nine indices, the ones which have the highest weighting are daily mean values of dry bulb temperature, dew point temperature and daily total of global solar radiation. Extreme values have lower weighting in the selection process. The selection process of TMM leads to a TMY, which contains a distribution of weather data reflecting a “normal” year rather than a year containing extreme values. This in turn leads to an indoor temperature distribution over a year that reflects a mean indoor condition occurring over a longer time span.

5.4 Parameter analysis methodology

When a building, a cooling tower and an HVAC-system as described in section 5.2, is coupled in a building simulation program it is obvious that the resulting temperature and humidity of the indoor air is dependent on many parameters. Some parameters have stronger and some has weaker influence on the resulting values of the primary variables described in section 5.3. To sort the parameters which have an influence on the primary variables, a single parameter sensitivity analysis will be done at first. In this case, it means that a number of parameters will, one by one, be changed from an original value stated at the base case. The change will be done in both directions for each parameter, i.e. a parameter will be changed to both a higher and a lower value. When all the parameters have been investigated, they can be sorted in strength according to their impact on the primary variables. 5.4.1 List of parameter alterations

In table 5.1 a complete description of parameters, which are to be altered, is displayed. Values of both base case and the higher and lower values are presented. In section 5.4.2 Discussion and motivation of parameter alterations, each parameter is discussed and the high and low values are motivated.


83

Table 5.1 List of parameter alterations

Parameter Lower

indoor

temp.

Base

case

Higher

indoor

temp.

Cooling Tower 1. Design primary Approach

3°C

4°C

5°C

2. Design range (both primary and secondary) 2°C 3°C 4°C

3. Design mass flow quota of liquid and air, (Ml/Ma)

2,0 1,0 0,5

4. Design cooling capacity, CTQ& CTQ&⋅25,1 CTQ& CTQ&⋅75,0

Building 5. Thermal capacity (weight)

”heavier”

”medium”

”lighter”

6. Location in building Ground floor Middle floor

Top floor

7. U-value in external walls U-value in windows

0,2 W/m2 K 2,0 W/m2 K

0,4 W/m2 K 2,5 W/m2 K

0,6 W/m2 K 3,0W/m2 K

8. Internal heat gain; magnitude (design maximum value, including solar radiation)

50 W/m2 60 W/m2 70 W/m2

9. Internal heat gain; working hours 8 h (08-16) 9 h (08-17) 10 h (08-18)

10. Geographical location Stockholm/ Arlanda

London/ Gatwick

Berlin

11. Facade direction East South West

HVAC-system

12. Design temp. difference Beams: waterroom TT −

5°C

7,5°C

10°C

13. Design cooling capacity cooling beams 60 W/m2 50 W/m2 40 W/m2

14. Design ventilation rate 1,67 l/s m2 1,25 l/s m2 0,83 l/s m2

15. Supply air temperature set point 15°C 17°C 19°C

16. Design secondary approach (heat exchanger) 0°C 1,5°C 3°C

17. Temperature gradient in office room 2°C 1°C 0°C

18. Set point temperature cooling control 21°C 22°C 23°C

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The reader should be aware that when changing the value of a parameter to a lower value, the indoor air temperature might rise and vice versa. Therefore all parameter changes causing the indoor air temperature to decrease is placed in the column “Lower indoor temperature”. Consequently, all parameter changes causing the indoor air temperature to increase are placed in the column “Higher indoor temperature”. This arrangement makes it easier to display the results of the sensitivity analysis. 5.4.2 Discussion and motivation of parameter alterations

1. Design Primary Approach

The design primary approach is a vital parameter for determining the ability of the cooling tower to chill the cooling water as close as possible to the ambient air wet bulb temperature. The parameter is however influencing size and investment cost of the cooling tower. In figure 5.7 the relation between the design UA-value and Primary Approach for three different values of Primary Range, 2, 4, and 6°C, of a cooling tower. The UA-value can be considered as proportional to the size of the cooling tower and consequently the investment cost. Figure 5.7 The relation between UA-value and Primary approach for three different

Primary ranges As can be seen the UA-value rises quickly when the Primary Approach reach under 4°C. The Primary Range also has an influence on the UA-value so that the lower the Range the higher the UA-value. In the literature, there is a unanimous agreement that the lowest practical and economical limit for the Primary Approach is 3°C (5°F). Based on this information and on figure 5.7, it is logical to set the parameter variations to 3, 4 and 5°C.

0,0

0,2

0,4

0,6

0,8

1,0

2,0 3,0 4,0 5,0 6,0 7,0 8,0

Primary Approach (°C)

Re

lati

ve

de

sig

n U

A-v

alu

e (

-)

Primary Range 2,0 Primary Range 4,0 Primary Range 6,0


85

2. Design Primary Range

The influence on the UA-value from the design Primary Range is discussed above. The values of the Primary Range are chosen as low as practically possible to make the mean temperature in the circuit as low as possible. The values of the Primary Range are also synchronized with the Secondary Range and are therefore chosen to 2, 3 and 4°C. These values are common temperature differences, Ranges, in the secondary circuit when connected to chilled ceilings or chilled beams.

3. Design mass flow quota of liquid and air, (a

l

M

M

&

&

)

The design flow quota between the air and water mass flow, in most literature called L/G (Liguid/Gas), basically determines the air mass flow rate. Normally the liquid mass flow is known from the choice of design Range and the design cooling capacity, and the air mass flow is the given through the design mass flow quota of liquid and air (L/G). In figure 5.8, this parameters relative impact on the UA-value of a cooling tower is shown. The impact on the UA-value is relative moderate in the displayed range of L/G values. This parameter has more importance regarding the energy use since the major part of the energy use of a cooling tower is due to the fan, typically about 80% of total use of electricity. Figure 5.8 The relation between UA-value and the design mass flow quota of liquid

and air (L/G). In the literature, a variety of L/G values can bee found. The differences between them are not very big and most values of L/G are in the range of 0,5 to 2. In Söylemez (2004) the optimum values of L/G is in the range from 0,82 to 2,45 for a forced draft counter flow cooling tower, depending on mean water temperature and air pressure, i.e. geographical elevation. The values of the design flow quota are chosen to be 0,5, 1,0 and 2.

0,0

0,2

0,4

0,6

0,8

1,0

0,5 0,75 1 1,25 1,5 1,75 2

Design flow quota, L/G (-)

Re

lati

ve

de

sig

n U

A-v

alu

e (

-)

86

4. Design cooling capacity, CTQ& (cooling tower)

The design cooling capacity of the cooling tower, CTQ& , i chosen to the same value as

the total design cooling capacity of the cooling beams together with design cooling capacity of the air cooler in the air-handling unit. With 50 W/m2 in design cooling capacity of the cooling beams and a specific airflow of 1,25 l/s m2 at base case, the total design cooling capacity of the tower is 7,69 kW. The parameter range is simply chosen to ± 25% of CTQ& .

The design values of the outdoor conditions at which the design cooling capacity is determined is 25°C dry bulb temperature and a relative humidity of 60%. This equals a wet bulb temperature of 19,5°C, which is very close to the ASHRAE design wet bulb temperature (0,4%) of London Gatwick that is 19,3°C (ASHRAE 2001a). The design outdoor condition has an influence on the size of the cooling tower. In figure 5.9 the relation between the UA-value and the design outdoor condition is shown. Figure 5.9 The relation between UA-value and the design outdoor wet bulb

temperature and an approach of 4°C. 5. Thermal capacity

The thermal capacity of the rooms and the building structure is a well known factor for playing an important role of the resulting indoor thermal climate. In this case, no numerical values are stated since the most common way to describe the thermal capacity of a room is in terms of ‘light’, ‘medium’ and ‘heavy’. In this context, the terms are instead chosen to be ‘lighter’ ‘medium’ and ‘heavier’. This is because the term ‘light’ is normally connected to a construction merely consisting of wood joist frames, wooden floor and ceiling and all the walls made of wood framework and gypsum boards. Commercial buildings with this type of construction are extremely unusual. In this case, all the variations of thermal capacity assume a building based on a construction with concrete slabs. In figure 5.10 – 5.12 the three different variations of constructions with different thermal capacity are presented.

0,0

0,2

0,4

0,6

0,8

1,0

10 12 14 16 18 20 22

Design ambient wet bulb temperature (°C)

Re

lati

ve

de

sig

n U

A-v

alu

e (

-)


87

First layer on each construction is always closest to the room.

Façade Ceiling slab Floor slab Inner walls

Concrete 0,15 m Insulation 0,08 m Concrete 0,08 m Plaster 0,01 m

Concrete 0,15 m Light concrete 0,02 m Plastic sheeting 0,005 m

Plastic sheeting 0,005 m Light concrete 0,02m Concrete 0,15 m

Gypsum board 0,026 m Insulation 0,03 m (frame work 0,095 m) Gypsum board 0,026 m

Figure 5.10 Building construction with Thermal capacity ‘heavier’


Figure 5.11 Building construction with Thermal capacity ‘medium’


Light concrete 0,125 m Insulation 0,02 m Light concrete 0,125 m Plaster 0,01 m

False ceiling 0,05 m Concrete 0,15 m Light concrete 0,02 m Plastic sheeting 0,005 m

Plastic sheeting 0,005 m Light concrete 0,02 m Concrete 0,15 m False ceiling 0,05 m


Ceiling slab

Inner wall

Floor slab

Facade

Window

False ceiling

Thermal capacity ’heavier’

Thermal capacity ’medium’

88


Figure 5.12 Building construction with Thermal capacity ‘lighter’ 6. Location in building

This parameter examines the impact of the location of the office floor in the building. Two obvious locations are either on the ground floor, utilising contact with the colder temperature of the ground, and on the top floor, with the extra heat load from the roof. A middle floor location is normally considered as a base case location. 7. U-values in wall and windows

The U-values of the base case building have been discussed in section 5.2.1 Model

building. The parameter alteration “low temperature” is set to a U-value of 0,2 W/m2 °C for the outer wall and 2,0 W/m2 °C for the window. These values are equal to a modern Swedish building with well insulated walls and three pane windows. The parameter alteration “high temperature” is set to a U-value of 0,6 W/m2 °C for the outer wall and 3,0 W/m2 °C for the window. These values are equal to an older Swedish building (before 1960) with poorly insulated walls and two pane windows.


Gypsum board 0,026 m Frame work and insulation 0,1 m Gypsum board 0,009 m Wood panel 0,025 m

False ceiling 0,05 m Concrete 0,15 m Light concrete 0,02 m Wooden parquetry 0,01 m

Wooden parquetry 0,01 m Light concrete 0,02 m Concrete 0,15 m False ceiling 0,05 m


Wooden parquetry

Thermal capacity ’lighter’


89

8. Internal heat gain; magnitude

The internal heat gains from people, lighting and equipment of the base case building have been discussed in section 5.2.3 Internal heat gain and moisture generation. The basic parameters governing these internal heat sources are kept constant through the parameter variations. The parameter variations in the total heat gain come from changes in the solar radiation to the room, due to changes of the size of the window. It should be noted that the values of the specific internal heat generation in table 5.1, section 5.4.1 List of parameter alterations, includes solar radiation in the room. In the literature, there are very few published values of normal ranges in commercial buildings, e.g. office buildings, which are based on actual measurements. Based on results from a series of test calculations the “low temperature value” is set to 50 W/m2. This value is close to what can be expected in a normal office room with the heat gains from people, lighting and equipment, as discussed in section 5.2.3 Internal heat gain and moisture

generation, together with solar radiation from a smaller window without solar shading or a larger window with good solar protection. The “high temperature” value of 70 W/m2 is close to what can be expected in a normal office room with a larger window with no or less efficient solar protection. Based on this discussion it is assumed that the range of 50 – 70 W/m2 covers a large part of normal office buildings. 9. Internal heat gain; working hours

The internal heat gains from people, lighting and equipment of the base case building are not only a matter of the magnitude of heat gain. It is also a matter of the daily duration of the internal heat gain. The working hours has been chosen to alter with ± one hour. It can be argued that working hours can be much longer in some commercial buildings. This is true, but they are very seldom much shorter than 08 – 16 as it is in the “low temperature” parameter setting. Since the working hours of the base case is from 08-17 the “high temperature” setting is consequently 08-18. It is also quite rare that working hours start earlier than 8 o’clock in the morning. 10. Geographical location

Geographical location is a difficult parameter to determine. It is not simply a matter of finding locations with evenly distributed latitudes. A geographical location parameter is more of a climate parameter. The climate part of this parameter has bigger influence than merely the latitude since the cooling tower performance is highly dependent of the ambient air wet bulb temperature. At an early stage of this project it was decided to analyse climates in the northern Europe, i.e. north of latitude 48 – 49°N, mainly to cover geographical areas where cooling system applications with chilled ceiling or chilled beams are common. The available climate data, in the format of TMY-tears, issued by ASHRAE (2001b) have been used to determine suitable locations. The data set in a TMY-year does not contain wet bulb temperature per se, but by processing all climate data in IDA ICE it was possible to get hourly wet bulb temperatures for all locations.

90

The selection started with the “low temperature” parameter in order to involve a Swedish location. The choice fell on Stockholm/Arlanda as the “low” parameter. From that climate, the duration curves of the ambient wet bulb temperature for all locations in figure 5.13 and the bar graph in figure 5.14 were used to determine the two other climates. In figure 5.14 all the locations used are identified. The base case location was selected to be London/Gatwick and the “high temperature” parameter was chosen to be Berlin. As a complement to the single parameter variation a limited multi parameter variation is performed. In this parameter variation, there are five locations used; Östersund and Stockholm/Arlanda (Sweden), London/Gatwick (UK), Berlin (Germany) and Paris/Orly (France). These locations cover the climatic interval in figure 5.14 very well. Figure 5.13 Duration curves of ambient wet bulb temperature (WBT) of 25 locations

in northern Europe. The duration displayed is the 1000 hours with the highest values of WBT. See figure 5.14 for a listing of the locations used.

10

12

14

16

18

20

22

24

7760 7960 8160 8360 8560 8760

Last 1000 hours of one year duration (8760 h)

We

t b

ulb

te

mp

era

ture

(°C

)

Berlin

London/Gatwick

Stockholm/Arlanda


91

Figure 5.14 Design values of ambient wet bulb temperature (WBT) of 25 locations in

northern Europe according to ASHRAE (2001a). The design values are those were the wet bulb temperature is exceeded 0,4%, i.e. 35 h/year, during a normal year.

11. Façade direction

This parameter gives three obvious parameter alterations, east as the “low” parameter, south as the base case and west as the “high” parameter. 12. Design temperature difference waterroom TT −

This parameter determines the characteristics of the cooling capacity of the cooling beams. The cooling capacity of an active chilled beam is dependent of the difference of the mean cooling water temperature and the room air temperature. This relation is showed in figure 5.15. The design mean water temperatures is chosen to 15, 17,5 and 20°C and the design indoor air temperature is set to 25°C in all cases. This leads to design temperature differences of 10, 7,5 and 5°C. Design temperature differences between 10 to 7,5°C represent fairly normal design values. A design temperature of 5°C is less usual due to the increase in the size, and hence investment costs, of the room cooling devices. The

14 15 16 17 18 19 20 21 22

Paris/Orly

Krakow

Warzaw

Brussels

Frankfurt

Düsseldorf

Amsterdam

Köln

Berlin

Bremen

Stuttgart

Hamburg

Prague

München

Luxenburg

London/Gatwick

Birmingham

Karlstad

Stockholm/Arlanda

Köpenhamn

Dublin

Göteborg/Landvetter

Aberdeen/Dyce

Östersund

Kiruna

Design wet bulb temperature (°C)

92

smallest design temperature difference represents the steepest line, i.e. cooling characteristics, in figure 5.15 and vice versa. The steeper the line, the more rapid increase in cooling capacity when the room temperature increases. A steeper line, i.e. a smaller design temperature difference, also leads to more or longer cooling beams in a room and thus higher investment costs as mentioned above. Figure 5.15 Cooling capacity characteristics of cooling beams at design mean water

temperatures 15, 17,5 and 20°C and a design indoor temperature of 25°C. 13. Design cooling capacity cooling beams

There are several ways to determine the required cooling capacity for a room. In this case, the base case cooling capacity is simply chosen to the “rule of thumb”-value common among Swedish HVAC engineers, namely 50 W/m2. In a room with the size of 12 m2, as in the base case, two modern active cooling beams with the length of 1,8 m, will suffice to provide the design cooling capacity. The variation is set to ± 10 W/m2 to cover realistic variations of design cooling capacities in commercial office buildings. 14. Design ventilation rate

The HVAC-system in this project is an air-water system. This type of system uses the ventilation primarily to supply the room with fresh air and extract contaminated air. The cooling or heating effect from the ventilation is always secondary. In the base case the specific ventilation rate 1,25 l/s m2 is used. This equals a ventilation rate of 15 l/s in an office room of 12 m2 with one person present. 15 l/s is a normally used figure which ensures, with sufficient marginal, that the CO2 level never exceeds 1000 ppm.

Ventilation rates below 10 l/s in a one-person office room are very rare. Based on this, a parameter variation of ± 5 l/s seems reasonable.

0

20

40

60

80

100

120

140

160

180

200

10 12,5 15 17,5 20 22,5 25 27,5 30

Indoor temperature (°C)

Re

lati

ve

co

oli

ng

ca

pa

cit

y (

%)

Design temp. diff. 5°C

Design temp. diff. 10°C


93

15. Supply air temperature set point

A supply air temperature set point near 17°C is fairly normal in air-water HVAC systems with mixing ventilation. The ventilation is not primarily used for cooling purposes, but it is quite normal to cool the air to gain some additional cooling from the ventilation, especially during warmer periods of the year. A supply temperature below 15°C may involve increased problems with thermal discomfort due to cold draft. A parameter variation of ± 2°C is hence chosen. 16. Design secondary approach

The design secondary approach is a temperature difference between the primary circuit cold fluid entering temperature, Tliq1_cold, and the secondary circuit chilled fluid

leaving temperature, Tliq2_out, see figure 5.16.

Figure 5.16 Definition of secondary approach The secondary approach, together with the design cooling capacity, defines the size, i.e. the UA-value, of the heat exchanger. In this case, a plate type heat exchanger is used. A plate type heat exchanger with a secondary approach of 1,5°C has a good performance but is not extremely costly (de Saulles, 1996). The “low” parameter alteration, 0°C, is the same as no heat exchanger between the two fluids. This is an option that, in some cases, can be feasible when the operation of a cooling tower in wet mode during freezing conditions can be handled with other solutions. The “high” parameter alteration is consequently 3°C. 17. Temperature gradient in office room

A temperature gradient is almost always generated in rooms occupied by people and with other internal heat sources. The size of the temperature stratification is determined not only by internal heat sources but also by e.g. type of ventilation, room height, solar radiation and type of heating and cooling devices. The reason why the temperature gradient can influence the resulting indoor air temperature may not be obvious. The cooling system with chilled ceilings in a room consists of the beam/beams, a temperature sensor and a control device, the supply and return pipes with control valve and actuator. The sensor is normally placed on an inner

Tliq1_cold Tliq2_out

Secondary Approach = Tliq2_out - Tliq1_cold

Primary circuit Secondary circuit

94

wall at a height of about 1,4 m above floor. The chilled beams are placed close to the ceiling, or the false ceiling whenever applicable, i.e. 2,4 – 2,5 m above the floor. The sensor, together with the control device, will strive for maintaining a temperature equal to the set point temperature. In the office room model the control device is a PI-controller, thus there will be no temperature offset. If there is a temperature stratification in the room, the chilled beams will be in a position with higher temperature than the temperature sensor. As mentioned above and at parameter #12. Design temperature difference waterroom TT − , the design temperature difference

( waterroom TT − ) determines the available cooling capacity of a chilled beam. roomT is in

this case the temperature in the close vicinity of the chilled beams. In the room model in IDA ICE the temperature sensor and the chilled beams are placed on different height in the room, hence the available cooling capacity of the chilled beams will vary depending on the magnitude of the temperature gradient but with the same set point temperature. The temperature gradient can in IDA ICE be either calculated or given. In this case, it is a given parameter and it is assumed linear from floor to ceiling. Novocelac and Srebric (2002) report a list of measured or simulated temperature gradients from six different references. The gradients were in the range 0-2 °C/m. It should be mentioned that these gradient were in cases with displacement ventilation and in this case mixing ventilation is assumed. The figures in table 5.1 represent the total temperature difference between floor and ceiling. The base case value of 1°C is hence equivalent to a gradient of 0,4°C/m. The decided parameter variations of 0 and 2°C are equivalent to a gradient of 0 and 0,8°C/m. 18. Set point cooling control

The air temperature in an office room has a typical diurnal variation with lower temperature during night time and higher during daytime. When the temperature rises in the morning, it will at times with high internal heat gain reach the cooling set point temperature. The cooling devices in the room will then start to cool the room acting as a “brake” on the rising temperature, thus hindering the temperature rise. If the set point is at a lower value, the cooling will start earlier and rise of the temperature will be less. The temperature will then level in the afternoon on a lower value than with a higher set point temperature. The person working in an office room normally sets the set point temperature on a local control device. In reality the set point temperature may vary from room to room depending on the thermal climate preferences, the awareness and knowledge of the heating and cooling system of each person together with the function of the control device and cooling device in each room. This situation is almost impossible to emulate in building simulation. The decided values of this parameter should therefore be considered as estimated mean set point temperatures The base case set point temperature is set at 22°C. The “low” value is seldom lower than 21°C and hence the “high” value is consequently set to 23°C.

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6 Monitoring of free cooling system with evaporative cooling device

In this chapter a free cooling pilot plant is described. The pilot plant comprises an

evaporative cooling device connected to an existing hydronic cooling system with

chilled beams in a real office building. The pilot plant has been monitored during

the summer period of 2007.

6.1 Introduction

To test the theoretical results from the simulations, a project was initiated to build and monitor a free cooling system with a evaporative cooling connected to an existing hydronic cooling system with chilled beams in a real office building. The free cooling system was sized as a pilot plant and thus only serving a smaller part of the office building to fit the budget of the project. The pilot plant together with monitoring and evaluation was financed by BELOK, Vasakronan and CIT Energy Management. The pilot plant was designed and built during 2006 and monitoring was conducted during the summer period of 2007. The pilot plant is located adjacent to a commercial office building at Kvarnberget at the city centre of Gothenburg, Sweden. The design work was performed by an HVAC-consultant (ÅF-konsult) in cooperation with a representative of the building owner (Vasakronan) and the project manager (the author of this thesis). 6.2 Pilot plant system lay-out

A sketch of the free cooling system connected to an existing hydronic comfort cooling system with chilled beams is outlined in figure 6.1. The connected system is a branch of the entire comfort cooling system of the building. The branch is serving part of the first floor of the building and the connection to the existing chiller is maintained through pipefitting with motor-operated shut-off valves. The shut-off valves are normally closed and the connection to the chiller is only opened if the indoor temperature rises above a predefined level or if the free cooling system is malfunctioning. The supply air is chilled by an existing air cooling coil in the central air conditioning unit connected to the existing chiller. The temperature in the existing supply air system is held constant. The supplemental heater in the supply air to the office space which is connected to the free cooling system has a special purpose. The heater is fitted to the existing duct system and its purpose is to emulate the somewhat poorer cooling capacity an air cooling coil would have connected to the free cooling system. The heater is controlled so that the supply air temperature is held at a temperature similar to what an evaporative free cooling system would have had at a given ambient air condition.

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Figure 6.1 Simplified sketch of the pilot plant at Kvarnberget, Gothenburg. Thin lines indicate existing systems.

The evaporative cooling device in the system is not a cooling tower in a traditional meaning, but an air cooled heat exchanger equipped with a spray-water device which gives the heat exchanger a function similar to an evaporative cooling tower, see figure 6.2. The heat exchanger is wetted by fresh water, i.e. once through spray water system, and thus there is no risk for growth of legionella bacteria in the spray water. In order to avoid misinterpretation the evaporative cooler in the pilot plant is hereafter named free cooler. The reason for choosing this kind of evaporative cooler is partly cost related and partly market supply related. The investment cost for an air cooled heat exchanger is significantly lower than for a closed loop cooling tower with the same cooling capacity. The market supply of closed loop cooling towers with the cooling capacity required in this case (< 20 kW) is very limited. The dominating market is power plants and industry with typical requirements of cooling capacities from a few hundred kW up to hundreds of MW. Air cooled heat exchangers, without additional water spray device, are occasionally used as free cooling sources in conventional comfort cooling systems. For this reason it is also interesting to analyze an established product on the market, i.e. an air cooled heat

Free cooling device

Freeze-protected circuit (primary)

Plate heat exchanger

Room cooling devices (chilled beams)

Connection to conventional cooling system with a chiller. Shut-off valves is normally closed

Supplementary heater to emulate a slightly poorer cooling than with conventional air conditioning

Conventional ventilation system with air conditioning connected to a chiller. (only supply part of system is shown)

Chilled water circuit (secondary)

6 Monitoring of free cooling system with evaporative cooling device

97

exchanger with an evaporative utility, to investigate its potential as a cooling device in a hydronic comfort cooling system. Detailed data about the pilot plant is presented in appendix B.

Figure 6.2 The free cooler in the pilot plant. A traditional air cooled heat exchanger

equipped with a spray water devise for wetting of the heat exchanger The free cooler in figure 6.2 chills a primary circuit, containing a propylene glycol solution (40%) as freeze protection. This facilitates operation in ambient temperatures below 0°C. The primary circuit transfer heat from a secondary circuit through a plate heat exchanger. The plate heat exchanger has a design approach/grädigkeit of 1°C and a design temperature difference between supply and return, a.k.a. range, of 3°C in the primary circuit as well as in the secondary ditto. In the secondary circuit there are chilled beams which have a design ∆Tm of 8°C and a design specific cooling capacity of 50 W/m2 floor area. The circulation in the primary and secondary circuits is maintained by pumps equipped with variable frequency drives (VFD). The control system strives to maintain a constant supply temperature in the secondary circuit. The constant supply temperature is maintained through sequential control where, in order and at increasing cooling demand, the pump in the primary circuit increase its speed, the fans in the free cooler increase their speed and finally the spray water is turned on. The control system also includes a function which disconnects the free

98

cooling system and connects the conventional comfort cooling system, through a set of shut-off valves, if the indoor temperature exceeds 25°C. This function was applied through demand from both the building owner Vasakronan and the tenant ÅF-konsult. The part of the building at which the pilot plant is installed is erected during the 1970’s with corresponding building standard. The office space which the free cooling system is serving has a floor area of 450 m2 with an outer wall facing south east. The people density is nominally 14 m2/person and at normal day-time operation the density is about 16 – 20 m2/person. Design power density for lighting is 15 W/m2. There are no external sun shading devices. 6.1 Monitoring system

Monitoring of the pilot plant was conducted during the summer period of 2007, i.e. may – august 2007. Each measurement point has been sampled every 10 minutes and the samples have been processed into one hour average figures. The following points have been monitored: Outside

- Air temperature (one point) - Relative humidity (one point) - Solar radiation (one point, measurement of both diffuse and total radiation)

Inside

- Air temperature (three points distributed in the office area) - Relative humidity (one point) - Air temperature gradient (three points at different heights above floor level) - Electric energy in office space (one point in the distribution box) - Liquid temperature in the secondary circuit (two points, supply and return) - Liquid flow rate in the secondary circuit (one point)

Free cooling system

- Liquid temperature in the secondary circuit (two points, supply and return) - Liquid temperature in the primary circuit (two points, supply and return) - Liquid flow rate in the secondary circuit (one point) - Liquid flow rate in the primary circuit (one point) - Liquid flow rate in the spray water supply (one point) - Cooling capacity in the secondary circuit (one point) - Cooling capacity in the primary circuit (one point) - Electric energy in pumps (two points, primary and secondary circuits) - Electric energy in fans (one point, fans in free cooler)

Further information about the monitoring system can be read in appendix C.

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7 Results and analysis of simulations

The results from the simulations described in chapter 5 “Simulation of a cooling

system with an evaporative cooling tower” are presented and analyzed in this

chapter. The results are separated into sections including Cooling tower

performance, Indoor thermal climate and Energy use. The chapter also contains

relevant findings from other researchers.

The results in section 7.1 Cooling tower performance are solely based on the prerequisites according to the base case conditions, see chapter 5 Simulation of a

cooling system with evaporative cooling. In section 7.2 Indoor thermal climate, the results are based on prerequisites in line with the base case conditions as well as all the parameter variations. In section 7.3 Energy use, the results are based on the same presumptions as in section 7.2, but with a limited number of parameter variations. The results presented in diagrams based on a full year simulation have a resolution of 15 minutes, i.e. the presented variable is logged every 15 minutes during a year. Some diagrams show one or several variables during three days in August; Monday – Wednesday the 18th – 20th of August. This period in the TMY year of London/Gatwick, which is the base case climate/location, is the one with the highest wet bulb temperature together with one of the warmest periods of the summer, i.e. the outdoor climate is at, or close at, ambient design conditions, see figure 7.1. The presented indoor temperatures during day time at this period are therefore close to, or at, their maximum. One exception is however the approach temperatures and the COP which are at their minimum during this period. Figure 7.1 Ambient conditions during August 18 – August 20, showing dry bulb and

wet bulb temperature.

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7.1 Cooling tower performance

The performance of the cooling tower in this section is focused on how close to the ambient wet bulb temperature (WBT) the supply water temperature in the secondary circuit can be brought, i.e. the size of the total approach. The other type of performance discussed in this chapter, namely the energy use, is presented in section 7.3 Energy use. As mentioned in chapter 5 Simulation of a cooling system with evaporative cooling

tower, the primary approach is the difference between the chilled primary circuit temperature and the ambient wet bulb temperature. The total approach is the difference between the supply temperatures in the secondary circuit and the ambient wet bulb temperature, see figure 7.2. Figure 7.2 Temperature diagram of a counter flow cooling tower with intermediate

circuit and visualisation of definition of design parameters. The size of the total approach temperatures is one of the most important factors for the ability to provide sufficient cooling to a building. The total approach is of course only important during the warmer part of the year, when high cooling loads and high wet bulb temperatures occur simultaneously. At cooler weather conditions, the cooling tower will normally operate at part load. One important question to answer is how well the cooling tower model can calculate the two approach temperatures during the influence of varying climate conditions and variations of mass flows and temperatures in the cooling system due to varying cooling loads and temperatures in the building. This is discussed in section 4.3 Validation of the

model. In this section, the question is further examined by running full year simulations at base case conditions.


Wet bulb (Ambient air) temp. in

Area

Temperature

Liquid in secondary circuit (water)

Moist air

Secondary approach

Primary approach

Primary range

Secondary range

Total approach

Cold side

Return side

Supply side

Warm side

Wet bulb temp. out

7 Results and analysis of simulation

101

Variations of the ambient wet bulb temperature, the chilled primary circuit temperature and the leaving temperature in the secondary circuit during three days in August are shown in figure 7.3. The figure visualizes how these temperatures fluctuate during this period. The ambient WBT varies from between 12 – 14 °C during night time up to peaks between 18 – 20 °C during daytime. The amplitude of the WBT is approximately 6 °C all three days, which is less than the amplitude for the ambient dry bulb temperature. The maximum dry bulb temperature was 28,8°C and the amplitude for the ambient dry bulb temperature was between 13-14 °C during Monday and Tuesday and around 9 °C on Wednesday, see figure 7.1. The chilled primary circuit temperature is close to equal with the ambient WBT during night time. Due to absence of cooling load during this time of night, the mass flows in both secondary and primary circuits are close to zero. The mass flows in both circuits are never equal to zero in the model of numerical reasons. The chilled primary circuit temperature is therefore very close to the ambient WBT at night time. During daytime the chilled primary circuit temperature is between 1 – 2°C above the WBT. These numbers are the primary approach temperature, which is shown in figure 7.4. The supply water temperature in the secondary circuit is not nearing the ambient WBT during night time as the chilled primary circuit temperature does. Both the primary and secondary circuit mass flows should go down to zero during the night, since there is no cooling load in the rooms, and consequently the temperature differences should reach zero, but they do not. This is mainly caused by different minimum limits in the cooling tower model, due to numerical reasons, for the mass flows in the two circuits. The supply water temperature in the secondary circuit is fluctuating between 1 – 3 °C above the WBT during this three-day period. These numbers are the total approach temperature, which is shown in figure 7.4. The fluctuations are mainly dependent on variations in mass flow in the secondary circuit as a result of changes in the indoor temperature over the day due to the changes in cooling load. Interesting to notice is that during Tuesday, August 19th, the leaving temperature in the secondary circuit, i.e. the supply temperature to the cooling beams, is between 20 – 21°C during working hours and the indoor temperature still only peaks at slightly above 25°C (figure 7.9, section 7.2 Indoor thermal climate). Bearing in mind that the design cooling heat gain is 60 W/m2 and the design cooling capacity of the chilled beams is 50 W/m2, it is indeed remarkable that a cooling system based only on a cooling tower as a “chiller” gives a resulting indoor temperature of about 25°C. It should be noted that the design cooling capacity of the chilled beams at base case conditions is defined at a design temperature difference of 7,5°C between mean cooling water temperature and the room air temperature, ∆Tair-liq. In this case, ∆Tair-liq is merely in the range 2 -3°C!

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Figure 7.3 Ambient WBT, primary circuit cold side temperature and secondary

circuit supply temperature between August 18 – August 20. Base case conditions.

Figure 7.4 Primary and total approach temperatures between August18 –

August 20. Base case conditions.

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In figure 7.5 the primary and total approach temperatures during a full year simulation is shown plotted against the ambient WBT. A few values below -5 °C WBT are omitted in the diagram. The values of the primary and total approach temperatures are very scattered but conglomerations of data form a linear, or almost linear, relationship with the WBT. The two approach temperatures are decreasing with increasing WBT. In figure 7.6 the primary approach temperature values and relation with the WBT is zoomed in between 5 – 20 °C WBT. In figure 7.7, the same thing is made with the total approach temperature. The primary approach temperature dependency with the WBT at higher WBT is less steep as with lower values of WBT. The change in inclination takes place at an ambient wet bulb temperature of about 13°C. This is due to that the set point temperature for the chilled water supply temperature in the secondary circuit is 16°C at base case. This set point temperature is achievable with a maximum ambient wet bulb temperature of 13°C together with a total approach of about 3°C. With lower ambient wet bulb temperatures than 13°C, the cooling tower starts to decrease the fan speed and thus the cooling capacity. Most of the primary approach values are in the range of 1 – 2 °C when ambient wet bulb temperatures are higher than 13°C. In figure 7.7, the total approach temperature shows the same dependency with the WBT as in figure 7.6. Most of these values are in the range of 1,5 – 3 °C when ambient wet bulb temperatures are higher than 13°C. Figure 7.5 Primary and total approach temperatures in a full year simulation plotted

against the ambient wet bulb temperature. Base case conditions.

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-5 -4 -3 -2 -1 0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20


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Figure 7.6 Primary approach temperatures in a full year simulation plotted against the

ambient wet bulb temperature. Base case conditions.

Figure 7.7 Total approach temperatures in a full year simulation plotted against the

ambient wet bulb temperature. Base case conditions.

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The presented temperatures in figures 7.5 to 7.7 apply well with published research of Costelloe and Finn (2002), Hasan and Gan (2002) and Facao and Oliveira (2000). Costelloe and Finn (2002) have made measurements on a prototype open type cooling tower test rig with a primary and secondary circuit (described in chapter 1.3 Previous

work). They found that the primary approach temperatures was in the range of 0,9 – 2,3°C and total approach temperatures in the range of 2,2 – 4,3 °C, with ambient WBT between 6,4 – 16,5 °C. Their findings are plotted in figure 7.6 and 7.7. Hasan and Gan (2002) have made measurements on a closed wet cooling tower for the generation of cooling water for use with chilled ceilings. Their findings are plotted in figure 7.6. The approach temperature is calculated by the author from published data by the Hasan and Gan (2002). Facao and Oliveira (2000) have also made measurements on a closed wet cooling tower for the generation of cooling water for use with chilled ceilings. They reported an approximate 8% increase in cooling tower thermal efficiency, ε, when the ambient WBT rises from 10°C to 20°C (increased thermal efficiency, ε, leads to lower approach temperatures). The increase is linear and the influence is similar at different air and water mass flow rates. Together with the validation of the cooling tower model in chapter 4.3 Validation of the

model against published data on cooling towers and the results from the full year simulations in figure 7.5 to 7.7, it seems that the cooling tower model can calculate accurate values of primary and total approach temperatures. This is an important result since the total approach gives the available cooling water supply temperature. As shown in chapter 5.4.2 Discussion and motivation of parameter

alterations, the cooling water temperature is of great importance for the available cooling capacity of the chilled beams and hence, the resulting indoor temperature. In the next chapter, the indoor thermal climate with this cooling system applied to an office building will be discussed.

7.2 Indoor thermal climate

In this section, results concerning the indoor thermal climate are presented. The air dry bulb temperature as well as the operative temperature is studied. The air temperature is presented as a mean value of all temperatures in the room since temperature stratification in the analysed room is set as a constant parameter value in all simulation cases (between 0 - 2°C from floor to ceiling). The operative temperature is calculated as for a seated person at 0,6 m above floor level in the middle of the analysed room. When calculating the operative temperature the air temperature and mean radiant temperature at 0,6 m above floor level is used. However, more focus will be put on the air temperature in the presentations of the results since most indoor thermal climate requirements in building projects is formulated with the air temperature in mind.

106

The relative humidity is also of interest since all cooling, of supply air as well as in the building, is sensible. No dehumidification occurs, which means that the relative humidity increases when the air temperature decreases, everything else constant. Hence, there might be concerns for unpleasantly high relative humidity at times. All results in this chapter are from the analysed room. For an explanation of the analysed room and its prerequisites, see chapter 5.2.1 Model building. The presentation of the results starts with indoor conditions at base case, followed by results of a parameter variation analysis. 7.3.1 Base case

The presentation of the results of the thermal indoor climate for the base case starts with a period containing the two last days of a five day heat wave followed by three days in August, as described earlier, from the full year simulation. Variations of the indoor air temperature, the operative temperature and the indoor relative humidity during the last two days in a five-day heat wave simulation are shown in figure 7.8, and during three days in August in a full year simulation are shown in figure 7.9. Figure 7.8 Indoor air temperature, operative temperature and indoor relative humidity

during the last two days in a five day heat wave simulation.

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Figure 7.9 Indoor air temperature, operative temperature and indoor relative humidity

between August 18 – August 20, full year simulation. In figure 7.8, the indoor air temperature range is about 21,5 – 25.5 °C. The operative temperature is slightly higher except during morning hours when the two temperatures are equal. The indoor relative humidity (RH) is in the range of 65 – 75 %. Interesting to notice is a slight decrease in the RH during daytime although one person is present and producing moisture. The decrease in RH can be explained with the increase in air temperature during working hours, which compensates for the moisture emitted from the person present. The ventilation rate is constant throughout the day. Figure 7.9 show a similar temperature pattern as figure 7.8. The indoor air temperature range is however slightly smaller, around 21 – 25 °C. The difference between the air temperature and the operative temperature is about the same as in figure 7.8. The largest difference is in the afternoon and is mainly explained by increasing surface temperatures in the room, mostly due to solar radiation. The solar radiation is quite high during Monday and Tuesday but lower on Wednesday. Consequently, the air temperature is lower on Wednesday as well as the difference between the air temperature and the operative temperature. The indoor relative humidity in figure 7.9 varies a bit more and is in the range 55 – 75 %. Interesting to notice is that the same decrease, after a small increase, in the relative humidity during daytime as in figure 7.8 can be found here as well. Figure 7.10 and 7.11 show a comparison between indoor air temperature with cooling and without cooling. The latter is when the cooling tower is shut down. Without cooling, the indoor temperature would reach levels which most people find unpleasant. In the morning around 8 am, the temperature is 23 - 24 °C and quickly rises during the day to reach 28 – 29,5 °C in the afternoon. With the cooling system operating, it

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manages to lower the indoor temperature by about 3,5 – 4°C during the afternoon, which is a significant reduction. It is significant in the sense that it makes a big difference in the thermal sensation of a person being exposed to temperatures around 24 - 25 °C compared to 28 – 29,5 °C. According to ISO 7730 (1994) the PPD index for a person is 8% at 24 °C, 13% at 25 °C and 55% at 29 °C, at 0,9 Clo and 1,2 Met as in the base case during the summer. A more detailed discussion about thermal sensation and indoor thermal climate can be found in ASHRAE (2001a) or Nilsson (ed) (2003).

Figure 7.10 Indoor air temperatures with and without cooling during the last two days

in a five day heat wave simulation. Figure 7.11 Indoor air temperatures with and without cooling between August 18 –

August 20, full year simulation.

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In figure 7.12, the indoor temperatures during working hours throughout a year, with and without cooling, are shown in the form of duration curves. With cooling, the system manages to hold the indoor temperature below e.g. 24 °C during 97,9% of working hours in a year. The same figure without cooling is 60%. See also table 7.2. Figure 7.12 Duration of indoor air temperature with and without cooling during one

year of working hours (08-17), full year simulation. This figure shows that the cooling system can keep the indoor temperature relatively constant at 22 °C at working hours during most of the year. It should be noted that the data behind the curves are not coincident in time, i.e. the temperatures on the two curves, for example at 90%, do not necessarily occur at the same day and hour of the year. The other indoor variable investigated is the indoor relative humidity. Figure 7.13 shows the indoor relative humidity plotted against indoor air temperature during one year of working hours (08-17) at a full year simulation. The data is for the case when cooling is activated. Interesting to observe is that the maximum relative humidity, 79%, occurs at indoor temperatures around 23°C. At temperatures above 23,5°C the relative humidity decreases and is about 60% at the highest indoor temperatures. The same tendency can be seen at the other two climates, which have been investigated, Stockholm and Berlin. Table 7.1 shows the maximum indoor relative humidity with coincident temperature and maximum indoor temperature with coincident relative humidity for the three climates.

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Figure 7.13 Indoor relative humidity plotted against indoor air temperature with

cooling during one year of working hours (08-17), full year simulation. Table 7.1 Maximum indoor relative humidity with coincident indoor temperature

and maximum indoor temperature with coincident indoor relative humidity. Full year simulation.

Climate Maximum indoor relative humidity (%)

Coincident indoor temperature (°C)

Maximum indoor temperature (°C)

Coincident indoor relative humidity (%)

Stockholm/Arlanda 85 23,4 24,8 59

London/Gatwick 80 22,7 25,3 61

Berlin 78 21,4 26,1 70

In table 7.2 different durations concerning indoor temperature and relative humidity are listed. The values are taken from the figures 7.12 and 7.13. The durations are expressed as percentage of working hours (08 – 17) when a limit value of either the indoor air temperature or the indoor relative humidity is exceeded.

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Table 7.2 Percentage of working hours (8 – 17) which a limit value is exceeded of either the indoor air temperature or the indoor relative humidity. Full year simulation.

Climate Percentage of working hours the indoor air temperature exceeds 24°C

Percentage of working hours the indoor air temperature exceeds 25°C

Percentage of working hours the indoor air relative humidity exceeds 70%

Percentage of working hours the indoor air relative humidity exceeds 80%

Stockholm/Arlanda 1,4% 0,0% 2,7% 0,8%

London/Gatwick 2,1% 0,3% 3,8% 0,0%

Berlin 5,6% 2,1% 4,1% 0,0%

7.3.2 Single parameter variations – indoor temperature

The indoor air temperature in the analysed room is a result of the performance of many parts in the cooling system, the building and the activity in it. The performance of the different parts is influenced by many parameter values. An obvious question in this context is; how “much” is each parameter contributing to the resulting indoor air temperature. To answer the question, parameter variations are made to examine the potential of each parameter regarding its influence on the indoor air temperature. The variations for each parameter are made with one higher and one lower value. To make the variations consistent the variations are called “lower indoor temperature” and “higher indoor temperature”. This is made since lowering the actual value of some parameters raises the indoor temperature and vice versa. All variations that cause a drop in the indoor temperature are gathered in the column labelled “lower indoor temperature” and all variations that cause a raise in the indoor temperature are in the column “higher indoor temperature”. A total of 18 parameters have been altered. A list of all parameter alterations can be seen in table 7.3. A more detailed discussion of the parameters can be found in chapter 5 Simulation of a cooling system with evaporative

cooling tower.

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Table 7.3 List of all parameter alterations

Parameter Lower indoor temp.

Base case

Higher indoor temp.

Cooling Tower 1. Design primary approach

3°C

4°C

5°C

2. Design primary range (both primary and secondary)

2°C 3°C 4°C

3. Design mass flow quota of liquid and air,

(a

l

M

M

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&

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2,0 1,0 0,5


Building 5. Thermal capacity (weight)

”heavy”

”medium”

”lighter”

6. Location in building Ground floor Middle floor

Top floor

7. U-value in external walls U-value in windows

0,2 W/m2 K 2,0 W/m2 K

0,4 W/m2 K 2,5 W/m2 K

0,6 W/m2 K 3,0W/m2 K

8. Internal heat gain; magnitude (maximum ) 50 W/m2 60 W/m2 70 W/m2

9. Internal heat gain; working hours 8 h (08-16) 9 h (08-17) 10 h (08-18)

10. Climate (Geographical location) Stockholm/ Arlanda

London/ Gatwick

Berlin

11. Facade direction East South West

HVAC-system

12. Design temp. difference Beams: waterroom TT −

5°C

7,5°C

10°C

13. Design cooling capacity cooling beams 60 W/m2 50 W/m2 40 W/m2

14. Design ventilation rate 1,67 l/s m2 1,25 l/s m2 0,83 l/s m2

15. Supply air temperature set point 15°C 17°C 19°C

16. Design secondary approach 0°C 1,5°C 3°C

17. Temperature gradient in office room 2°C 1°C 0°C

18. Set point temperature cooling control 21°C 22°C 23°C

Figure 7.15 shows results from parameter variations with heat wave simulations and figure 7.16 shows results from parameter variations with full year simulations. Both figure 7.15 and 7.16 show the differences in maximum indoor temperature compared to the results from the base case. For example, when varying the parameter climate to a value equal to “lower indoor temperature” in table 7.3, using heat wave simulations, fig 7.15, the maximum indoor air temperature is 0,76°C lower than for the base case. For heat wave simulations, figure 7.15, no single parameter can affect the maximum indoor temperature more than 0,76°C. 12 parameters out of 18 could not change the maximum indoor temperature more than 0,4°C. Interesting to notice is that all cooling tower parameters are in the range of ±0,2°C. The two most powerful parameters are climate and internal heat gain. Facade direction - “higher temperature” and location in


113

building - “higher temperature” as well as thermal capacity – “lower temperature” and design temperature difference cooling beams – “lower temperature” parameter variation also generates a relatively big change. In fig 7.16, full year simulation, the pattern is almost the same as with figure 7.14. The parameters working hours and ventilation rate give a bit more change with the full year simulation than with the heat wave simulation. For the rest of the parameters the differences are rather small compared to figure 7.15. Not surprisingly, the climate parameter is powerful. A warmer climate raises the cooling load at the same time as the cooling capacity of the cooling tower is reduced. The magnitude of the indoor temperature variation when varying climate locations from “higher temperature” to “lower temperature” is purely a result from the choice of climate range. If the variations in climate had been from Kiruna in the north of Sweden to Palermo in Sicily, Italy, the variation in indoor air temperature would naturally have been larger. The variation in internal heat gain is ±10 W/m2. This range, from 50 W/m2 to 70 W/m2 (sensible heat), probably gather a large part of existing maximum cooling loads in office buildings. Since the internal heat gain directly affects the indoor air temperature, it is expected that the variation in indoor air temperature in figure 7.15 and 7.16 is relatively high. The facade direction west gives high temperatures in the afternoon and the location on the top floor leads to a higher transmission of heat through the ceiling, which adds to the internal heat gain and raises the temperature relatively high. High thermal capacity is a well-known factor for reducing indoor temperatures. The parameter variation “lower temperature” has a bigger influence than the “higher temperature” one. This parameter variation is not a result of an equal change up or down of the value of thermal capacity. The variation is more a result of the different building constructions, especially concerning the fabrics of the inner layers on the floor or ceiling. The parameter “lower temperature” implies large areas of exposed concrete slab or wall, which gives this alternative a stronger impact than the parameter alternative “higher temperature” does. A detailed description of the different building constructions for the different parameter variations is found in chapter 5.4.1 List of parameter alterations. A smaller design temperature difference for the cooling beams directly influences the size of the cooling beam, i.e. they become bigger. A consequence of a smaller design temperature difference is that the cooling capacity rises faster when the indoor air temperature increases. This leads to lower amplitudes in the daily indoor temperature swing. A more detailed discussion about this is presented in chapter 5.4.2 Discussion and motivation of

parameter alterations.

114

Figure 7.15 Differences in maximum indoor air temperature, compared with the results

of the base case, for parameter alterations in table 7.3. Heat wave simulations.

-1,0 -0,8 -0,6 -0,4 -0,2 0,0 0,2 0,4 0,6 0,8 1,0

Prim Approach

Prim Range

Flow quota

Cooling capacity tower

Thermal capacity

Location in building

U-value

Internal heat gain

Working hours

Climate

Facade direction

Design temp. diff. Beams

Cooling capacity Beams

Ventilation rate

Supply air temp. set point

Secondary Approach

Temperature gradient

Set point Cooling

temperature difference (°C)

Lower temperature Higher temperatue

-0,76°C


115

Figure 7.16 Differences in maximum indoor air temperature, compared with the results

of the base case, for parameter alterations in table 7.3. Full year simulations

7.3.3 Multi parameter variations

Since it is not practically possible to make a full multi factorial analysis with 18 parameters (this would require 262 144 simulation runs) a limited parameter variation is made with three of the most dominant parameters (see figure 7.16); climate/location, thermal capacity and design internal heat gain (including solar radiation). In this case, there are five locations; Östersund and Stockholm (Sweden), London (UK), Berlin (Germany) and Paris (France), two thermal capacities; medium and heavy, and three design heat gains; 50, 60 and 70 W/m2, in the parameter variations. The design heat gain includes maximum solar radiation in the room. Combining all possible parameter combinations involves 30 full year simulation runs. The result is shown in figure 7.17 – 7.20

-1,0 -0,8 -0,6 -0,4 -0,2 0,0 0,2 0,4 0,6 0,8 1,0

Prim Approach

Prim Range

Flow quota

Cooling capacity tower

Thermal capacity

Location in building

U-value

Internal heat gain

Working hours

Climate

Facade direction

Design temp. diff. beams

Cooling capacity Beams

Ventilation rate

Supply Air Temperature

Secondary Approach

Temperature gradient

Set point Cooling

Temperature difference (°C)

Higher temperature Lower temperatue

116

Figure 7.17 Maximum indoor air temperature, at five different design wet bulb

temperatures, i.e. climate/location and for three different internal total heat gains. Medium weight thermal capacity. Full year simulations.

Figure 7.18 Percentage of working hours when indoor temperature exceeds 24°C, at

five different design wet bulb temperatures, i.e. climate/location and for three different internal heat gains. Medium weight thermal capacity. Full year simulations.

R2 = 0,8332

R2 = 0,8763

R2 = 0,7875

22

23

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25

26

27

28

15 16 17 18 19 20 21 22

Design wet bulb temperature, ASHRAE 0,4%, (°C)

Ma

xim

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in

do

or

air

te

mp

era

ture

(°C

)

50 W/m2 60 W/m2 70 W/m2

Linjär (50 W/m2) Linjär (60 W/m2) Linjär (70 W/m2)

Medium thermal capacity

R2 = 0,7131

R2 = 0,7945

R2 = 0,8369

0%

1%

2%

3%

4%

5%

6%

7%

8%

9%

10%

15 16 17 18 19 20 21 22

Design wet bulb temperature, (ASHRAE 0,4%) (°C)

Pe

rce

nt

of

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rkin

g h

ou

rs r

oo

m

tem

pe

ratu

re e

xce

ed

s 24 °

C

50 W/m2 60 W/m2 70 W/m2


Medium thermal capacity


117

Figure 7.19 Maximum indoor air temperature, at five different design wet bulb

temperatures, i.e. climate/location and for three different internal heat gains. Heavy weight thermal capacity. Full year simulations.

Figure 7.20 Percentage of working hours when indoor temperature exceeds 24°C, at

five different design wet bulb temperatures, i.e. climate/location and for three different internal heat gains. Heavy weight thermal capacity. Full year simulations.

R2 = 0,8566

R2 = 0,7973

R2 = 0,8637

22

23

24

25

26

27

28

15 16 17 18 19 20 21 22

Design wet bulb temperature, ASHRAE 0,4% (°C)

Ma

xim

um

in

do

or

air

te

mp

era

ture

(°C

)

50 W/m2 60 W/m2 70 W/m2


Heavy thermal capacity

R2 = 0,6049

R2 = 0,706

R2 = 0,7816

0%

1%

2%

3%

4%

5%

6%

7%

8%

9%

10%

15 16 17 18 19 20 21 22

Design wet bulb temperature, (ASHRAE 0,4%) (°C)

Pe

rce

nt

of

wo

rkin

g h

ou

rs r

oo

m

tem

pe

ratu

re e

xce

ed

24 °

C

50 W/m2 60 W/m2 70 W/m2


Heavy thermal capacity

118

The five locations are represented by their respective design wet bulb temperature in the figures 7.17 – 7.20. The design condition for the locations is the design ambient wet bulb temperature from ASHRAE (2001a). This design condition is statistically exceeded 0,4% of the year, i.e. 35 h/year. Locations in northern Europe with their respective design condition are presented in table 7.4 below. Table 7.4 Design wet bulb temperature at different locations in northern Europe

Location Design wet bulb temperature (ASHRAE 0,4%)

Coincident dry bulb temperature (ASHRAE 0,4%)

Kiruna 14,2 18,9

Östersund 15,9 21,7

Aberdeen 17,5 21,0

Göteborg 17,7 23,5

Dublin 17,9 20,5

Copenhagen 18,2 23,2

Oslo/Fornebu 18,4 24,1

Stockholm 18,4 23,6

Birmingham 18,5 23,8

Helsinki 18,7 23,6

Tallin 19,0 23,0

London/Gatwick 19,3 25,0

Luxemburg 19,5 25,4

Riga 19,6 23,7

Munich 19,6 26,7

Prague 19,7 26,2

Hamburg 19,9 25,7

Stuttgart 19,9 27,3

Berlin 20,1 27,0

Amsterdam 20,3 24,8

Koln 20,3 27,1

Frankfurt 20,5 27,8

Brussels 20,7 26,5

Warzaw 21,0 27,6

Vienna 21,1 28,4

Krakow 21,2 27,9

Paris/Orly 21,4 27,9

Budapest 21,4 30,5


119

For a given design internal heat gain and thermal capacity, the maximum room air temperature shows a fair linear correlation with the design wet bulb temperature. The linear correlation has an R2 value between 0,78 and 0,89 The same plotting as above can be made with the variable “the percentage of working hours when the indoor temperature exceeds 24°C”, named %24°C, against the design wet bulb temperature. The plotting is made at the same given design internal heat gains and thermal capacities as above. In this case, a linear correlation can be found between %24°C and the design wet bulb temperature. Here the linear correlation is somewhat weaker with an R2 value between 0,60 and 0,84. The range of design wet bulb temperature scale in the figures 7.17 – 7.20, i.e. 15 – 22°C, cover entirely the wet bulb design conditions in the following countries in Europe (in alphabetical order); Austria, Belgium, Czech Republic, Denmark, Estonia, Finland, Germany, Iceland, Ireland, Latvia, Lithuania, Luxembourg, Netherlands, Norway, Poland, Slovakia, Sweden and United Kingdom. The range 15 – 22°C also covers the northern parts of France, parts of Switzerland and the northern parts of Hungary. Parameter variations are also made to study the upper and lower limits of the indoor air temperature when all parameters are set to either “lower indoor temperature”, i.e. best case, or “higher indoor temperature”, i.e. worst case. Examining the best and worst case gives an image of the span the indoor air temperature can vary within, when altering the parameters within the given limits in table 7.3. Figure 7.21 Duration of the indoor air temperature on three different locations with

and without cooling at best case parameter settings. Full year simulations.

20

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30

32

34

36

0% 10% 20% 30% 40% 50% 60% 70% 80% 90% 100%

working hours during a year (%)

Ind

oo

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mp

era

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(°C

)

Berlin

London

Stockholm

Berlin

London

Stockholm

Without cooling

With cooling

Without cooling

With cooling

120

In figure 7.21, the annual duration of the indoor air temperature on three different locations, Stockholm, London and Berlin, with and without cooling at best case parameter settings are presented. In the case with cooling, the analysed system performs as well as any conventional vapour compression cooling system. Most of the year, the indoor temperature is held constant close to the cooling set point, 21°C, at all three locations. The maximum indoor temperature in Berlin is 23,3°C when cooling is activated. Even at no cooling, the indoor conditions are not extreme. Stockholm and London has maximum indoor temperatures slightly above 26°C and Berlin above 28°C. When conditions are turned to worst case, as in figure 7.22, the indoor temperatures are of course higher. With cooling, Berlin peaks at 30,1°C, London at 29,1°C and Stockholm at 28,4°C. Without cooling the maximum temperatures are between 32°C to 35°C. Note that the set point temperature for the cooling beams is 23°C at the worst case. Figure 7.22 Duration of the indoor air temperature at three different locations with and

without cooling at worst case parameter settings. Full year simulations. Interesting to notice is that the difference between maximum indoor temperatures with and without cooling seems to bee rather conform. The difference between maximum indoor temperatures with and without cooling is also called temperature depression. In table 7.5 is the temperature depression shown for different locations and different cases. For Stockholm and London the temperature differences is in the range 3,5° to 4,9°C. For Berlin it is somewhat higher 4,8°C to 5,4°C. The results in table 7.5 indicate that a

20

22

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26

28

30

32

34

36

0% 10% 20% 30% 40% 50% 60% 70% 80% 90% 100%

working hours during a year (%)

Ind

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era

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(°C

)

Berlin

London

Stockholm

Berlin

London

Stockholm

Without cooling

With cooling

With cooling

Without cooling


121

temperature depression of somewhere between 3,5°C and 5,5°C between the maximum indoor temperature with and without cooling can be expected using this type of cooling system. Table 7.5 Indoor temperature depression, i.e. difference in maximum indoor air

temperature between cases with and without cooling for Base case, Best case and Worst case. Full year simulation

Climate Base case Best case Worst case

Stockholm/Arlanda 4,9°C 3,6°C

London/Gatwick 4,2°C 3,8°C 3,5°C

Berlin 5,4°C 4,8°C

In table 7.6 the maximum indoor air temperatures for best case, worst case with cooling and the difference between them are shown. The difference is very consistent ranging from 6,5°C to 6,8°C. Table 7.6 Maximum indoor air temperature for Best case, Worst case with cooling

and the difference between them. Full year simulation

Climate Best case Worst case Difference

Stockholm/Arlanda 21,8°C 28,4°C 6,6°C

London/Gatwick 22,6°C 29,1°C 6,5°C

Berlin 23,3°C 30,1°C 6,8°C

In table 7.7 the percentage of working hours which the indoor temperature exceeds 24, 25 and 26 °C shown, in best case and worst case, with and without cooling. The figures are calculated from the results shown in figure 7.21 and 7.22. Table 7.7 Percentage of working hours which the indoor temperature exceeds 24, 25

and 26 °C in best case and worst case, with and without cooling.

Climate

With cooling Without cooling

Best case Worst case Best case Worst case

24°C 25°C 26°C 24°C 25°C 26°C 24°C 25°C 26°C 24°C 25°C 26°C

Stockholm/Arlanda 0,0% 0,0% 0,0% 11,5% 5,0% 1,9% 15,1% 4,1% 0,6% 31,2% 23,5% 15,9%

London/Gatwick 0,0% 0,0% 0,0% 12,9% 6,7% 2,4% 16,4% 5,1% 0,3% 33,7% 23,2% 15,9%

Berlin 0,0% 0,0% 0,0% 16,5% 9,8% 6,5% 24,3% 12,3% 5,5% 38,0% 29,9% 22,2%

122

At best case with cooling, the cooling capacity is much oversized and the indoor temperature is never exceeding 24°C in any climate location. At worst case with cooling, the numbers for Stockholm and London are relatively equal, from around 2% at 26°C up to 11-13% at 24°C. For Berlin, the result is 3-4% units above the other two locations. Bearing in mind that these are numbers for the worst case, the time each temperature limit is exceeded is close to where normal and acceptable indoor summer temperature design limits according to presented requirements in table 7.8. Table 7.8 Recommended peak summer conditions for natural ventilated or air

conditioned buildings in design guides (Lawrence Race, 2003)

Lawrence Race (2003) has made a compilation of requirements based upon durations from published design guides. As can be seen, there are big differences in the requirements but also in the criteria. The requirements and criteria from BRE and MOD are exclusively for naturally ventilated buildings. The conclusion of the parameter variation is that climate location and internal heat gains play a relatively big role influencing the indoor temperature. Other parameters characterized as medium potent are; location in building, facade direction, thermal capacity and design temperature difference of cooling beams. The difference between best and worst case considering the maximum temperatures, in a full year simulation, is very consistent ranging from 6,5°C to 6,8°C. The percentage of working hours which the indoor temperature exceeds 24, 25 and 26 °C in is zero for the best case and for the worst case close to where normal and acceptable indoor summer temperature design limits are for many buildings. From the analysis of the results it can be concluded that the analysed cooling system is capable of cooling a building with relatively high internal heat gains, still maintaining an acceptable indoor air temperature.

Organisation Requirement Criteria (°C)

CIBSE Guide A (1999) Not exceeded for more than 5% of annual occupied period

25 resultant temp (≈ operative temp)

BRE Environmental Design manual – Summer conditions in naturally ventilated offices (1988)

“Satisfactory” “Intermediate” “Min acceptable” Acceptable design risk for these conditions not to be exceeded for more than 30 days in 10 year

23±2 24±2 25±2

MOD Defence Works Functional Standard, DMG 07:1996

Not exceeded for more than 2,5% of a month

30 resultant temp (≈ operative temp)

Energy Efficiency Office – ‘Performance specification for the energy efficient office of the future’ – Best Practice Programme Report 30 (1995). Applies to air-conditioned and naturally ventilated buildings

No more than 1% of year No more than 5% of year

>28 >25 resultant temp (≈ operative temp)


123

7.3 Energy use

The energy use of the cooling system is presented and analysed in this chapter. The energy use, i.e. use of electricity, is important since it generates both operating costs and contributes to pollution of the environment. The results are presented at parameter settings corresponding to the base case if not otherwise mentioned. 7.3.1 Base case

At base case the COP and the cooling capacity of the cooling tower is presented during three days, 18th to 20th of August in figure 7.25. The COP is previously defined in chapter 4 Simulation model development, equation 4.44.

Figure 7.25 COP and cooling capacity of cooling tower between August 18 –

August-20. Full year simulation. The COP is ranging between 4 – 8,5 during daytime whereas it drops down to zero during night time. The drop in COP down to zero comes from the drop of cooling capacity during night time when there is no need for cooling in the rooms and the ambient temperature is below the set point temperature of the supply air. The cooling capacity of the cooling tower, during day time, is in the range 2 – 3 kW. The COP during a full year is plotted against the ambient wet bulb temperature in figure 7.26. Very little is found in literature concerning COP of a cooling tower applied to chilled ceilings or chilled beams. Costelloe and Finn (2001) have published data from measurements of the COP on a cooling tower at different ambient air wet bulb temperature (WBT), with the purpose of cooling water for chilled ceilings or chilled

0

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00:00 06:00 12:00 18:00 00:00 06:00 12:00 18:00 00:00 06:00 12:00 18:00 00:00

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pa

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COP Cooling capacity


124

beams. Their findings are incorporated in figure 7.26. As can be seen, their data is well in accordance with the results from the simulations in this thesis. The simulation result gives a yearly mean COP of 7,0 based on the quota between yearly total cooling energy and yearly total electrical energy for pumps and fans in the system.

Figure 7.26 COP cooling tower in relation to the ambient air wet bulb temperature.

Full year simulation. The COP of the cooling tower increases when the WBT drops under 12-13°C. This is due to reduction of fan speed, and thus reduction of energy use, when the required cooling tower capacity is lowered to maintain a constant supply temperature in the chilled liquid. The electricity to the fan stands for the major part of the energy use of the cooling tower, hence the rise in COP. In table 7.9, the energy use on a yearly basis of the different devices in the cooling tower is presented.

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Simulation results Costelloe & Finn 2001


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Table 7.9 Energy use of the different devices in the cooling tower and comfort cooling system at base case conditions. Full year simulation

Device Use of electricity

(kWh/year) Use of electricity

(%)

Pump for spray water 39 6,8

Pump in primary circuit

52 9,1

Pump in secondary circuit

21 3,7

Cooling tower fan 459 80,4

Total use 571 100,0

7.3.2 Single parameter variations – energy use

The energy use of the cooling tower and the cooling system is a result of the performance of many parts in the cooling system, the building and the activity in it. The performance of the different parts is influenced by many parameter values. An obvious question in this context is; how “much” is each parameter contributing with to the resulting energy use. To answer this question parameter variations in this chapter are made to examine the potential of some parameters regarding its influence on the energy use. The parameters used are the examined cooling tower parameters, see chapter 7.2.2 Single parameter variations – indoor temperature, and two other parameters being presumed as giving a large affect on the cooling tower energy use. Table 7.10 shows the selected parameters. Table 7.10 List of parameter alterations

Parameter Lower indoor temperature

Base case

Higher indoor temperature

Cooling Tower 4. Design primary approach

3°C

4°C

5°C

5. Design primary range (both primary and secondary)

2°C 3°C 4°C

6. Design mass flow quota of liquid and

air, (a

l

M

M

&

&

)

2,0 1,0 0,5


Building 19. Internal heat gain; magnitude

(maximum )

50 W/m2

60 W/m2

70 W/m2

10. Geographical location/ Climate Stockholm/ Arlanda

London/ Gatwick

Berlin

126

The variations for each parameter are made with one higher and one lower value. To make the labelling of the variations consistent with the variations in chapter 7.2.2 Single

parameter variations – indoor temperature, they are called “lower indoor temperature” and “higher indoor temperature”. They are referred to these names since lowering the actual value of some parameters raises the indoor temperature and vice versa. All variations that cause a drop in the indoor temperature are gathered in the column labelled “lower indoor temperature” and all variations that cause a raise in the indoor temperature are in the column “higher indoor temperature”. A more detailed discussion of the parameters can be found in chapter 5 Simulation of a free cooling system with

evaporative cooling tower. The variables examined are the COP (Coefficient of Performance) of the cooling system and the use of electrical energy in the cooling system.

Figure 7.27 Differences in cooling system energy use in comparison to base case for

different parameter alterations. For an explanation of the labels “higher temperature” and “lower temperature” see above. Full year simulation.

In figure 7.27 the variations in cooling tower annual energy use is presented due to parameter variations. Two parameters are dominating; Primary range and the Flow quota. The latter is quite often referred to as the L/G, i.e. the mass flow relation Liquid/Gas, in cooling tower literature. These two parameters directly influence the air mass flow. The air mass flow is in turn directly influencing the energy use of the fan which is the dominating part of the contributors to the total energy use, see table 7.9. The primary range influences the mass flow of liquid at a given cooling capacity of the

-60%

-40%

-20%

0%

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40%

60%

80%

100%

Prim

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Prim

Range

Flow quota Cooling

capacity

CT

Internal

heat gain

Climate

Dif

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%)

Higher temperature Lower temperature


127

cooling tower. The primary range, which is inverse proportional to the liquid flow, directly influences the air mass flow through the flow quota (L/G), hence the total use of energy in a cooling tower. The mass flow quota (L/G) has an inverse proportionality to the annual energy use, the higher the L/G the lower the design air mass flow; hence a lower use of energy is obtained. The flow quota has the strongest impact on the energy use, with a change of up to 75% of the annual use of energy going from a flow quota of 1,0 to 0,5. The other parameters in figure 7.27 have a more indirect and a bit weaker influence on the annual energy use. The influence is within ±20% of the energy use in the base case with the exception of one climate alteration.

Figure 7.28 Differences in cooling system annual COP in comparison to base case for

different parameter alterations. For an explanation of the labels “higher temperature” and “lower temperature” see above. Full year simulation.

The variations of the coefficient of performance (COP) for the cooling tower and the comfort cooling system together are shown in figure 7.28. The bars in the figure are more or less a mirror image of the bars in figure 7.27. However, there are some discrepancies. The most obvious is the Primary range but also the Primary approach.

-60%

-40%

-20%

0%

20%

40%

60%

80%

100%

Prim

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Flow quota Cooling

capacity

CT

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(%

)

Higher temperature Lower temperature

128

7.4 Findings from other sources

There are very few findings from other published sources where the thermal indoor climate is analysed in a building with a cooling system which is the same as, or equal to, the one presented in this thesis. Facão (n.d.) presents outcome from the EcoCool project and Bohler et al. (2002) presents findings from simulations using the simulation program ConsoClim. In figure 7.22 duration curves from Facão (n.d.) show the outcome of a simulation for a whole cooling season in Zurich, using the simulation program TRNSYS. The simulation was performed on an office building with a typical construction and utilization pattern together with a cooling system containing a closed wet cooling tower and chilled ceilings. During working hours, indoor temperature reaches a maximum of 27.5ºC, and is above 26ºC less than 40 hours, which is about 1% of the annual working hours. The design cooling capacity for the chilled ceiling is unfortunately not mentioned. Bearing in mind that the climate in figure 7.22 is Zurich, the duration curves of Facão (n.d.) comply fairly well with the duration curve in figure 7.11, except for at low ambient temperatures (there seems to be no heater in the room).

Figure 7.22 Diagram from Facão (n.d.) showing duration of indoor air temperature

together with temperatures of ambient air and chilled water and cooling tower performance. Outcome from TRNSYS simulations.

The diagrams 7.23 and 7.24 show findings from Bohler et al. (2002). They made simulations with the simulation tool ConsoClim (a French program) on an office room of 15 m2 equipped with chilled ceilings connected to an open cooling tower with an intermediate heat exchanger to avoid fouling in the chilled water secondary circuit. There was no conventional chiller attached. The research work included analysis of the operative temperature in an office at a number of different prerequisites, i.e. three design cooling capacities of the chilled ceilings, two building inertia, two solar gains, two internal heat gains, two orientations (east and west) and three different climates

Diagram from Facão (n.d.) with results from the EcoCool project

Room air temperature

Ambient air dry bulb temperature

Ambient air wet bulb temperature


129

(locations of Trappes, Nice and Carpentras). In figure 7.23 and 7.24, made by the author with data from Bohler et al. (2002), the figures from simulations using the location Trappes is shown. Trappes is located close to Paris. The room cooling load contains both latent and sensible loads as well as effects from thermal storage in the building fabrics. The maximum operative temperature is the mean operative temperature on the three hottest hours of occupation. The three different cooling capacities, 30, 40 and 50 W/m2 is the design cooling capacity at a temperature difference between the room air temperature and the mean cooling water temperature of 6ºC. Figure 7.23 Outcome of simulations from Bohler et al. (2002). Location is Trappes

near Paris and with medium thermal weight in the room. *) The maximum operative temperature is the mean operative temperature on the three hottest hours of occupation.

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Room cooling load (W/m2)

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30W/m2

40 W/m2

50 W/m2

Medium thermal weight

Cooling capacity chilled ceiling

130

Figure 7.24 Outcome of simulations from Bohler et al. (2002). Location is Trappes

near Paris and with heavy thermal weight in the room. *) The maximum operative temperature is the mean operative temperature on the three hottest hours of occupation.

It is difficult to translate the room cooling load, including both latent and sensible heat, as used by Bohler et al. (2002) to the total internal sensible heat gain, including solar radiation, used in this thesis. The total internal sensible heat gain in a room is however almost always higher than the sensible cooling load. The difference between them can vary considerably depending on several factors, e.g. the thermal capacity of the building fabric. Considering this, the figures from Bohler et al. (2002) apply well to the results in this thesis.

20

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Room cooling load (W/m2)

Ma

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)

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40 W/m2

50 W/m2

Heavy thermal weight

Cooling capacity chilled ceiling

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8 Results and analysis of measurements

The results from the monitoring described in chapter 6 “Monitoring of free cooling

system with evaporative cooling device” are presented and analyzed in this

chapter. The results are separated into sections including Ambient conditions,

Indoor thermal climate, Cooling tower performance and energy use.

8.1 Ambient conditions

The ambient conditions during the summer period, i.e. May 01 – August 31 year 2007,

was relative normal according to statistical data from the Swedish Meteorological and

Hydrological Institute (SMHI). When comparing the monthly mean values of the

measured outdoor temperature with the normal temperatures for Gothenburg from

SMHI during the period of 1961 – 1990 (VVS 2000, 2003), the difference are relatively

small for the whole period of May – August, see table 8.1. In table 8.1 the measured

monthly mean values are between 0,3°C - 2°C higher than the normal temperatures for

May, June and August. In July the measured monthly value is 1°C lower. For the whole

period, May – August, the mean value of the measured temperatures is 0,7°C above the

statistically normal temperature.

Table 8.1 Comparison between measured monthly outdoor mean temperatures and

statistically normal temperatures (SMHI; Gothenburg 1961 – 1990)

Period

Monthly mean outdoor temperature [°C]

Measured values,

Kvarnberget Gbg

Normal values, Gothenburg

(1961-1990) SMHI

May 11,8 11,5

June 17,6 15,6

July 16,0 17,0

August 17,6 16,2

May - August 15,8 15,1

In figure 8.1 the measured outdoor temperature is shown during the period of May 1 –

August 31, 2007. During about a week in the beginning of June, the only real heat wave

occurred with temperatures above 30°C during several days. In the first half of August

there was a shorter period of warm and sunny weather. As for the rest, the whole period

is characterized by varying temperatures between 10 – 25°C.

132

Figure 8.1 Measured outdoor temperature at Kvarnberget, Gothenburg

during the period of May 1 – August 31, 2007.

A duration curve of the measured outdoor temperature during May – August is shown

in figure 8.2. The outdoor temperature was below 20°C during 85% of the time.

Figure 8.2 A duration curve of the measured outdoor temperature at

Kvarnberget, Gothenburg during the period of May 1 – August 31, 2007.

0

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40

Ou

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Outdoor temperature 01 may - 31 augKvarnberget, Göteborg 2007

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Duration outdoor temperature 1 may - 31 augKvarnberget, Göteborg 2007


133

In the next chapter the indoor thermal climate in the office building at Kvarnberget in

Gothenburg, with the described pilot plant system applied, will be discussed.

8.2 Indoor thermal climate

In this section, results concerning the indoor thermal climate in the office building at

Kvarnberget are presented.

The resulting indoor thermal climate is not solely dependent on the outdoor climate

such as air temperature and solar radiation. It is also, to a fairly high degree, dependent

of the magnitude of the indoor heat generation from people, lighting and electrical

appliances as well as the ability of the building structure to store excess heat.

Furthermore the indoor thermal climate is of course dependent on the ability of the

comfort cooling system to remove excess heat.

The indoor air temperature in relation to the corresponding outdoor air temperature is

shown in figure 8.3. The dark (blue) dots are representing the indoor temperature at

times when the free cooling system is running. The lighter (orange) dots represent the

indoor temperature when the conventional comfort cooling system with a vapour

compression chiller is running.

As described in chapter 6.2 Pilot plant system lay-out there is a function in the control

system which disconnects the free cooling system if the indoor air temperature exceeds

25°C. Thus there are no registered indoor air temperature readings above 25°C when the

free cooling system is running. Besides that, the free cooling system has experienced

malfunctions at a few occasions and then the conventional comfort cooling system has

taken over the indoor climate control.

However, as figure 8.3 indicates, the free cooling system seems to be able to keep the

indoor air temperature below 25°C at outdoor temperatures up to 27°C.

134

Figure 8.3 Indoor air temperatures in relation to the outdoor air

temperature at Kvarnberget, Gothenburg, during the period May 1 – August 31,

2007.

The duration of the indoor air temperature, when the free cooling system has been

active and running, is shown in figure 8.4. The free cooling system has been running

active during 2585 h of the period May 1 – August 31, which is equivalent to 88% of

this period. The time scale 0 – 100% in figure 8.4 is similar to 0 – 2585 h.

The duration of the indoor relative humidity during the period of May 1 – August 31,

2007 is shown in figure 8.5. In this case the relative humidity for all hours throughout

the period of May 1 – August 31 is shown (a total of 2952 h). This is because the supply

air to the office space involved is delivered from the central air conditioning unit where

the air is cooled and also dehumidified. The relative humidity has been in the interval

30 – 70% during the period in question.

18

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IIn

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Outdoor temperature [°C]

Indoor temperature vs outdoor temperatureKvarnberget Gothenburg, May 1 - Aug 31 2007

With free cooling With ordinary chiller


135

Figure 8.4 Duration of the indoor air temperature when the free cooling system is

active and running. Kvarnberget, Göteborg for the period May 1 – August

31, 2007

Figure 8.5 Duration of the indoor relative humidity for the period of May 1 – August

31, 2007 at Kvarnberget, Göteborg.

18

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0% 10% 20% 30% 40% 50% 60% 70% 80% 90% 100%

IIn

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Duration of indoor temperature when free cooling system is activeKvarnberget Göteborg, May 1 - August 31 2007

136

8.3 Performance and energy use

The resulting indoor climate when applying evaporative cooling is of prime interest in

this thesis. The energy use and the performance of the free cooling system is however

also of great interest. Few, if any, building owners can neglect the running costs and the

environmental impact from a cooling system. One performance factor of interest is the

supply temperature in the secondary circuit, the circuit connected to the chilled beams in

the building (see figure 5.5 and 6.1).

Figure 8.6 Secondary supply temperature vs. outdoor temperature at evaporative

cooling mode for the period of May 1 – August 31, 2007 at Kvarnberget,

Göteborg

Figure 8.6 shows the secondary supply temperature (SST) in the system with chilled

beams vs. the outdoor drybulb temperature. In area A the set point temperature is 12°C.

During the period the samples in area B was measured, the outdoor temperature was

above 12°C. There is a distribution of SST in area A around the diagonal line. The

diagonal line indicate where the SST is equal to the outdoor temperature. The SST:s

which are above the diagonal line, i.e. is higher than corresponding outdoor

temperature, occurs mostly during night time or early morning when the outdoor

relative humidity is close to 100% . The SST:s which are below the diagonal line, i.e. is

lower than corresponding outdoor temperature, occurs mostly during daytime when the

relative humidity is below 100%. During these circumstances the wet bulb temperature

is several degrees below the dry bulb temperature. Consequently it is possible to

achieve SST:s to a temperature below the outdoor dry bulb temperature.

0

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Seco

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su

pp

ly t

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[°C

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Outdoor drybulb temperature[°C]

Secondary supply temp. vs outdoor temp.evaporative cooling

Kvarnberget, Göteborg 1 maj - 31 aug 2007

B

A


137

In area B the set point temperature is 17°C in order to examine the energy use at a

higher set point temperature.

0

1

2

3

4

5

6

7

8

0 5 10 15 20 25 30 35

CO

P [

-]

Outdoor dry bulb temperature [°C]

COP vs outdoor temperatureevaporative cooling

Kvarnberget, Göteborg, May 1 - August 31 2007

Figure 8.7 COP of the free cooling system vs. outdoor dry bulb temperature at

evaporative cooling mode for the period of May 1 – August 31, 2007 at

Kvarnberget, Göteborg

In figure 8.7 the COP of the evaporative cooler vs. the outdoor temperature is shown.

There is a fairly large variation of COP at a given outdoor temperature. A reason for this

is a variation in the cooling demand in the building while the evaporative cooler is

running at full speed to cool the liquid to the set-point temperature.

The variation in COP is not only from differences in cooling demand. COP for an

evaporative cooler is usually increases with decreasing outdoor temperature. The basic

reason comes from the reduction of fan speed at lower temperatures since the set-point

temperature can be reached with lower air flow through the evaporative cooler.

Simultaneously the cooling demand can be fairly high at lower temperatures. This two

factors together lead normally to an increasing COP at lower outdoor temperatures. At

lower temperatures the COP can reach about 10 and even reach 20 – 30 and more at

outdoor temperatures below 5 - 10°C.

Regarding the evaporative cooler in the pilot plant, the COP is relatively low. Most of

the time the COP is equal to or lower than the COP for a conventional chiller. However

it is important to bear in mind that the COP varies with the outdoor wet bulb

temperature and is lower at high wet bulb temperatures. The yearly mean COP, often

called SPF (Seasonal Performance Factor), is therefore usually higher than the values

showed in figure 8.7.

138

There are basically two main reasons for the low COP of the evaporativ cooler in the

pilot plant. The first is a fairly low L/G-value (see chapter 5.4.2 Discussion and

motivation of parameter alterations for explanation of L/G). In this case L/G is 0,15 at

design rate. Compared to normal range of L/G of 0,5 – 2 for cooling towers it can be

considered low. A low L/G implies high air flow rates, which in turn leads to a low

COP.

The second reason involves the design of the evaporative cooler. As mentioned earlier,

the evaporative cooler in the pilot plant is a type of standard air cooled heat exchanger

equipped with a spray water device for the evaporative function. It is primarily designed

for cooling liquid with outdoor air without the spray water function. When the spray

water function is active the water is sprayed from underneath, the whole heat exchanger

area is not completely wetted due to a quite narrow spacing of the flanges of the heat

exchanger. Due to the narrow spacing the water tends to adsorb to the flanges and also

between them, due to the surface tension of the water. This leads to a more narrow

space for the air stream to pass, which both influence the heat exchange and the pressure

drop of the air in a negative way.

Together, this influences the COP of the evaporative cooler in a negative way.

Standard air cooled heat exchangers, sometimes labeled fluid cooler, is primarily

designed for high cooling capacity at a low price, not to have the lowest life cycle cost.

Among the biggest producers of air cooled heat exchangers, there are none, who declare

COP for their range of products.

139

9 Conclusions and discussion

In this chapter, conclusions from previous research and the results in this thesis

are presented. Conclusions are followed by a discussion concerning various

aspects of comfort cooling in general and the comfort cooling system examined in

the thesis in particular.

9.1 Conclusions

Research on hydronic comfort cooling systems with cooling towers as a sole free

cooling source in conjunction with chilled ceilings or chilled beams have been in

progress since the late 90’s. The amount of published research is relatively limited.

However, the collected experience from the published material gives clear indications of

the following:

� It is possible to maintain cooling water supply temperature below 18 – 20°C

during a large part of the year in the northern parts of Europe. The northern parts

are approximately north of latitude 48 - 49° N. For most climates north of latitude

48 - 49° N the availability is above 90%. Costelloe and Finn (2002) reports annual

availability of chilled water from a cooling tower with a fixed approach of 3°C at

different supply water temperatures for the locations Dublin and Milan. For

cooling water supply temperature below 18 – 20°C the availability is 97 – 99% of

the year for Dublin and 67 – 78% for Milan.

� The cooling water temperature obviously exceeds its design temperature for short

periods during warm weather. Despite this, the maximum indoor temperature in

most cases is in the range 25 - 27°C, based on measurements or outcome from

building simulations. However, the maximum temperatures occur only during a

limited time. There is limited information about the design prerequisites

concerning building, internal heat gains and cooling capacity in the published

material. Published findings concerning duration of indoor temperature during for

example a year have not been found.

� The COP of a cooling tower, with the mentioned application, is significantly

higher than of a conventional chiller. A measured yearly average COP of around

33 has been reported (Costelloe & Finn, 2002). Other sources, e.g. Hasan et.al.

(2007), indicate however lower values, around 7 - 8. There is however limited

information regarding measured COP of cooling towers in literature.

The following can conclude the results from the research presented in this thesis:

1. A cooling tower model (CTM) is developed for the building simulation program

IDA ICE. The CTM has been validated against both published data, which is

based on measurement data and data calculated from an accurate model, and

measured data from a pilot plant. The effectiveness – NTU method forms a basis

for the CTM. The effectiveness – NTU method is also used in other cooling tower

models in internationally recognized building simulation programs like TRNSYS

and EnergyPlus.

140

The CTM can be used to simulate air heat exchangers as well as most common

types of cooling towers in conjunction with a hydronic cooling system. The CTM

can be configured as an open tower or one with a closed circuit. The heat

exchange configuration can be either counter flow or cross flow in the model. The

cooling tower fan in the CTM can be either single speed type or equipped with

variable speed drive. The main results from CTM are mass flow and temperature

of supply cooling liquid, energy use of fans and pumps in the system and COP of

the cooling tower during the simulated period.

2. Results from the simulations of a base case building gives a maximum indoor air

temperature about 25,5°C, whether during a heat wave or a full year simulation.

When varying the design total internal heat gain, including solar radiation,

between 50 – 70 W/m2 the maximum indoor air temperature is in the range

25,7 – 26,7°C at a design wet bulb temperature (dWBT) of 21°C

24,9 – 26,0°C at a dWBT of 19°C

24,0 – 25,2°C at a dWBT of 17°C

All figures are for a building with medium thermal capacity and prerequisites

according to the base case. For a building with heavy thermal capacity the

maximum temperature figures are 0,5 – 0,7°C lower than those with medium

thermal capacity. Design wet bulb temperatures of 17, 19 and 21°C is in

accordance with ASHRAE climate data (0,4%) and represent climates in the

northern Europe above latitude 48 - 49°N. The design total internal heat gain

includes design heat gains from humans, lighting, office equipment and

transmitted solar radiation.

The duration of the indoor air temperature during working hours has also been

presented. The percentage of indoor air temperature during working hours

exceeding 24°C is in the range 3 – 8% with total internal heat gains between 50 –

70 W/m2 at a dWBT of 21°C and medium thermal capacity. At heavy thermal

capacity, the range is 2 – 5%. For a dWBT of 17°C the range has decreased to 0 –

2% and 0 – 1% respectively.

3. Results concerning the indoor relative humidity (RH) from the simulations of the

base case building, and during a year, give a maximum value of 80% RH at

coincident indoor air temperature of 22,7°C. At maximum indoor air temperature,

25,3°C, the relative humidity is 61% RH.

When varying the climate parameter, the same pattern is observed, i.e. the

maximum relative humidity is in the range 78 -85% RH at coincident indoor

temperatures of 21,4 – 23,4°C and at maximum indoor air temperatures, 24,8 –

26,1°C, the coincident relative humidity is 59 – 70% RH.

4. The annual coefficient of performance (COP) for the cooling tower at base case is

7,0. The annual coefficient of performance comes from the quota of the annual

cooling energy and the annual electric energy to pumps and fans in the system.

The variation of the COP over a year is from close to zero to about 100. The

annual total electrical energy use is about 2 kWh/(m2, year), where the area is the


141

total area of the simulated office floor including office rooms, corridors and

middle section. The cooling tower fan uses the major part of the total energy,

about 80% and is in this thesis always equipped with a variable speed drive.

When varying the design parameters of the cooling tower, i.e. primary approach,

primary range and the mass flow quota, the annual COP differ between 10 – 70%

from the base case COP, i.e. a COP of 7. The annual electric energy use varies

between a few percent up to 75%.

5. The single parameter variations gives information of the relative impact each

parameter has concerning indoor air temperature. The parameters with the highest

impact are “Climate/location”, “Total internal heat gain” and “Thermal capacity

of the building fabric”. The parameters ”Location in building” (especially when at

top floor) and ”Design temperature difference of chilled beams” also have a fairly

high impact. Interesting to notice is that all cooling tower related design

parameters as well as the “Secondary range”-parameter has a relatively low

impact on the indoor air temperature, i.e. less than about 0,2°C.

The results from simulation in this thesis concerning maximum indoor air temperature

and the COP of a cooling tower largely confirm the findings in earlier published

research. The results from the monitoring of the pilot plant fairly well confirm the

results from the simulations regarding indoor climate. The COP of the evaporative

cooler in the pilot plant is however lower than compared to findings in the literature and

results from simulations in this thesis. The reason is mainly a low L/G-value together

with constrained airflow through the partly wetted flanges of the heat exchanger.

This thesis brings the following new information:

� A validated cooling tower model in NMF-language.

� The impact each of 18 different design parameters in the cooling tower, the

building and the HVAC-system, have on the thermal indoor environment.

� The indoor thermal climate in a commercial building with the examined comfort

cooling system over a large geographical span, i.e. climates equal to the northern

parts of Europe, which is approximately north of latitude 48 - 49°N.

� The annual duration, i.e. the percentage of working hours when a given indoor

temperature limit, e.g. 24, 25 and 26°C, is exceeded, during a whole year.

� The indoor relative humidity during a whole year in a normal office room with the

examined comfort cooling system applied.

� The impact each of six different design parameters, in the cooling tower and the

building, have on the use of electric energy in the cooling tower and the cooling

system.

� Operational experience from a pilot plant; the evaporative cooler could keep a

good indoor thermal climate in a normally occupied office with chilled beams

based on normal design criteria. The energy efficiency of the evaporative cooler

was however poor. It had a relatively low COP during the measured period (May

1 – Aug 31, 2007) ranging from close to zero up to about 4.

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9.2 Discussion

In this thesis, a hydronic comfort cooling system with a cooling tower as the sole free

cooling source is discussed. Although hydronic comfort cooling systems with chilled

ceilings, chilled beams or fan coils and cooling towers by themselves are well-

established techniques, the combination of them can be considered as novel.

New and unknown techniques are often treated with scepticism, especially in the

building industry. Sometimes the scepticism is of good reasons but at times, it can be

derived from lack of knowledge. For this reason, it is important to gain and disseminate

knowledge of the comfort cooling system in this thesis, and other low energy and CFC-

free systems, to give design teams more options concerning technical systems with low

environmental impact to chose from when selecting solutions for new or refurbished

buildings.

In the following sections, various aspects of comfort cooling in general and the comfort

cooling system examined in thesis in particular are discussed.

9.2.1 Design

Design aspects of a cooling tower system

In a normal design procedure of an arbitrary heat exchanger, the media which is to be

heated or chilled, e.g. air or water, and the heating or chilling media are assumed to

have constant design temperatures. The two mass flows are also normally assumed

constant. These assumptions make it easy to select a proper size of an arbitrary heat

exchanger from either design charts or design computer programs issued by a producer.

Normal design of a cooling tower in, for example, an industrial process with high

temperatures is straightforward. After determined the design wet bulb temperature

(dWBT) for the actual location, the difference between the dWBT and the required

supply temperature of the cooling water settles the approach temperature. The required

cooling capacity determines the mass flow and the range temperature. After settling a

proper L/G, the mass flow of air in the tower is determined.

The matter of designing a cooling tower as the sole cooling source in a hydronic

comfort cooling system is however not obvious. Figure 9.2 illustrates the dilemma.


143

Figure 9.2 Example of the design dilemma when designing a cooling tower as the

sole cooling source in a hydronic comfort cooling system

Assuming the dWBT is 19°C and the cooling tower has a primary approach of 4°C, the

supply water from a cooling tower has a design temperature of 23°C. If there is a

secondary circuit, an additional 1 – 2°C must be added to the supply water design

temperature to make up for the secondary approach of the intermediate heat exchanger.

The chilled ceilings or beams have in this example a design supply temperature of 14°C.

The design temperature can of course be raised up to 18 – 20°C, with a following

increase in size of the chilled ceilings or beams, but going as far as 23°C or more is

practically and economically impossible since a normal design indoor maximum

temperature is in the range 24 - 26°C. A major part of the room would be filled with

chilled ceilings or beams at these conditions!

How can this dilemma be solved? There are principally two ways to deal with this

problem:

1. The static approach: Lower the dWBT until temperature 1 and 3 meets as in

normal static design of heat exchangers. Temperature 1 can possibly be raised up

to 18 – 20°C to minimize the decrease of the dWBT. Lowering the dWBT will

increase the size and investment cost of the cooling tower. In section 7.4.2

Discussion and motivation of parameter alterations, figure 7.9 and adjacent text,

this is discussed. On the other hand, a rise in the design supply temperature of the

chilled ceilings or beams will increase the size and investment cost of them. It is

easy to see that there is an interesting optimization problem here; at what dWBT

will the total investment cost for the cooling tower and the chilled ceiling or

beams be at minimum? This has however not been investigated in this thesis. Still

it is an important question which should be looked into more thoroughly.

The result of a low dWBT would be an oversized cooling tower, i.e. the cooling

tower can produce supply cooling water at a smaller approach than a tower

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Design wet bulb

temperature

Design supply temperature

from cooling tower

Ordinary design supply

temperature for chilled beams

3

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normally equal!

Temperature [°C]

144

designed at a higher dWBT when the actual WBT is above the (low) dWBT. This

is of course positive but not necessarily needed. If it is necessary depends on the

indoor thermal requirements, the cooling load and the design conditions of the

chilled ceilings or beams. A matter of fact is however that the “oversized” cooling

tower generates higher investment costs.

2. The dynamic approach: This approach involves an acceptance of the dilemma,

i.e. leave the design temperatures where they are. This is the approach which has

been applied in this thesis. For the base case, the cooling tower is designed with a

dWBT of 19,5°C with a primary approach of 4°C and a secondary ditto of 1,5°C.

At these design conditions, a secondary supply water temperature would be 25°C.

This temperature equals exactly the design indoor temperature. Designing the

chilled beams at this temperature would imply an infinite size of the beams! The

design supply temperature in the base case is however at a moderate level of 16°C

with a secondary range of 3°C. With a design heat gain of 60 W/m2 and a design

cooling capacity of 50 W/m2 the resulting maximum temperature in the analysed

room is 25,3°C for the base case.

How can the resulting indoor temperature be almost equal to the design indoor

temperature when the design temperature condition of the cooling water system

obviously is not met? The answer is found in the dynamic heat balance of the

room. There is no heat flux or temperature inside or outside a normal room that is

static, especially not during the warmer part of the year.

An example can illustrate this: When the maximum indoor temperature at base

case, 25.3°C, was registered there was a few days of heat wave with maximum

outdoor dry bulb temperatures about 28.5°C. However the WBT reached the

design wet bulb temperature (dWBT=19.3°C) only for a very short period during

one day. The rest of the days during the warm period when the maximum indoor

temperature occurred, the WBT was often several degrees below the dWBT

(about 1 - 4°C). The chilled beams could however not reach the design cooling

capacity of 600 W, i.e. 50 W/m2. Instead, the cooling capacity of the chilled

beams varied throughout the day between 150 W and 260 W. This is mainly due

to variations in the WBT, which gives variable supply temperatures, together with

variations in the indoor air temperature. The indoor air temperature varied during

the day from about 22°C in the morning to about 25°C in the afternoon. Another

important factor, apart from the cooling system, is the heat storage in the building

fabric. Its maximum was more than double the cooling capacity of the chilled

beams.

The lesson to learn from this is that none of the heat fluxes in a room can be regarded as

constant. Not even if a conventional chiller provides constant supply cooling water

temperature. The required indoor temperature can still be obtained even if all the

temperatures fluctuate, including the cooling water supply temperature. It is important

to realize that the design conditions in a building hardly ever occur, whereas for

example the design conditions for a heat exchanger in an industrial process can be ever

present. In the former case, the design conditions are merely a way to size a cooling

device. The matter of greatest importance is normally how and when the different heat


145

fluxes in a room interact, i.e. the dynamic pattern of the heat fluxes. In general, the

pattern itself has the greatest influence on the resulting temperature. In the latter case,

the design conditions are crucial since these conditions always occur and there are no

other heat fluxes influencing the process.

In this discussion, it is essential to stress that the capacity of a cooling device in a room

is not unimportant, it is just less important compared to the dynamic pattern of all heat

fluxes. For the industrial heat exchanger example, the design heat (or cooling) capacity

is however crucial.

A metaphor of the dynamic pattern of heat fluxes can be two teams playing

soccer. One team is the “heating” team and the other is the “cooling” team. The

outcome of the game is the resulting indoor air temperature. If one or two team

members in the cooling team have a minor injury or a period of tiredness, the

team still can perform well (and win!). The performance will of course decrease if

the team plays with reduced number of players or if they all play bad. In team

sports as well as in the dynamic heat balance of a room, the result is the collected

effort of each member in each team.

The only way to find out if the design cooling capacity of the room cooling devices,

when connected to a cooling tower, is sufficient to keep the indoor air temperature

below the maximum allowed room temperature is to run simulations with a qualified

building simulation tool. Only then can all the dynamic heat fluxes and fluctuating

temperatures be taken into account.

This design approach may feel unusual for an HVAC engineer, but the difference from

traditional design is not very big. The only actual difference is that the supply cooling

water temperature is varying due to the variable wet bulb temperature during the day.

All other temperatures and heat fluxes varies anyway, whether the supply cooling

temperature is constant or not.

Risk of condensation on chilled beams

One argument against chilled ceilings or beams is the risk of condensation with

subsequent risk of indoor precipitation. This is a risk which is apparent at combinations

of warm and humid climate, limited or no dehumidification of supply air, and low

cooling system temperatures. There is however several ways of eliminating this risk in

systems with normal constant supply cooling water temperatures in the range of 12 –

14°C. For example, humidity sensors can be used either in an exhaust air duct or in

zones with increased risk for condensation. If the humidity rises above a given limit, a

control device actuates a rise in the supply cooling water temperature above the dew

point temperature to eliminate the condensation risk. There is also a special sensor on

the market which detects condensation on single room cooling devices and simply

closes a valve to the cooling device, thus eliminating the risk of condensation locally.

For a hydronic cooling system with chilled ceiling or beams connected to a cooling

tower as the sole provider of chilled water, the situation is different. In this case, the

cooling water supply temperature has a natural minimum limit in the wet bulb

146

temperature. In practical terms, the cooling water supply temperature is a few degrees

Celsius above the WBT. In figure 9.3, an example of possible conditions in a room is

visualized. The room is chilled with a cooling system with only sensible cooling, e.g. a

hydronic cooling system with chilled ceiling or beams connected to a cooling tower as

the sole provider of chilled water.

In the example in figure 9.3, the outdoor condition is 25°C with a humidity ratio of

0,011 kg water/kg dry air, i.e. close to 60% relative humidity. In the building, this air is

chilled to 22°C. The vertical thick line illustrates a case with no humidification in the

room or with very high air change rates, i.e. ventilation rates, together with a limited

moisture load. At this case, the dew point temperature is about 15,5°C. The cases with

dotted lines indicate the resultant condition in the room with various combinations of

moisture loads and air change rates.

The two cases which have a resulting relative humidity of 70% and 80% RH represent

cases with normal air change rates and moisture loads in commercial buildings. In these

cases, the dew point temperatures are well below the upper limit dew point temperature.

In the case with a resulting relative humidity of about 95% RH, the actual dew point

temperature is above the upper limit dew point temperature, hence there is a risk for

condensation on the room cooling devices. This case has a very high moisture load or a

combination of high moisture load and low air change rate, which is unusual in most

commercial buildings.

Figure 9.3 Enthalpy – humidity diagram (Mollier chart) with an illustrated example of

temperature and humidity conditions in a room with chilled ceiling or

beams in a hydronic cooling system connected to a cooling tower as the

sole provider of chilled water. Sensible cooling only.

humidity (kg/kg)

100%RF

-20

-15

-10

-5

0

5

10

15

20

25

0,000 0,005 0,010 0,015 0,020

Outdoor wet bulb

temperature

Approach

Dew point temperature at

no humidification in room

Upper limit dew

point temperature

70% RH 80% RH 90% RH 100% RH Outdoor air

condition

Possible indoor air

conditions in room


147

As shown in section 7.2 Indoor thermal climate, the highest values of indoor relative

humidity, 78 – 85% RH, occurs at indoor air temperatures between 22 and 24°C. Above

that, the indoor relative humidity decreases. Based on the results in this thesis

concerning the indoor thermal climate and the discussion above, it can be considered as

extremely unusual to experience condensation on room cooling devices connected to a

cooling tower as the sole provider of chilled water.

Design cooling capacity of chilled beams

There are several ways to determine the required cooling capacity for a room. Usually

some form of simulation program is used. If it is a qualified simulation tool, it considers

all important factors in a dynamic heat balance to determine the required cooling

capacity in a room. The required cooling capacity is determined with a given indoor

thermal requirement. In some cases, the design cooling capacity is chosen from a “rule

of thumb” figure. In this thesis, the base case cooling capacity is chosen to the “rule of

thumb”-value common among Swedish HVAC engineers, namely 50 W/m2. This is

done simply to obtain an even number and also to apply with normal cooling capacities

in Swedish buildings. In a room with base case conditions and with a base case design

temperature difference between room air temperature and the mean cooling water

temperature (∆Ta-l) of 7,5°C, two modern active cooling beams with the length of 1,8

m, will suffice to provide the design cooling capacity. A ∆Ta-l of 7,5°C is a bit lower

than standard ∆Ta-l which is in the range 8 - 10°C. If a mean standard ∆Ta-l is about 9°C

the chilled beams in the base case are about 15% bigger than with a ∆Ta-l of 9°C.

9.2.2 Indoor climate

Thermal climate

The results in section 7.2 Indoor thermal climate indicate that the indoor climate can be

kept at conditions where most people find it comfortable at a high or very high duration

of normal working hours, i.e. the indoor air temperature does not exceed 24°C or 25°C

more than between fractions of a percent up to 10 percent. This conclusion is valid for

normal commercial buildings in climates similar to those in the northern parts of

Europe, i.e. north of latitude 48 – 49°N.

A cooling system comprising a cooling tower in conjunction with chilled ceilings or

beams can however provide comfortable indoor conditions in warmer climates than

examined in this thesis. The design internal heat gains including solar radiation must

then be kept at moderate values, i.e. below about 40 – 45 W/m2 at medium thermal

capacity or below about 45 – 50 W/m2 at higher thermal capacity. The warmer the

climate the lower the allowed internal heat gain. An exception from this statement may

be in climates with combinations of hot and humid conditions. At climates with design

wet bulb temperatures higher than about 25 - 27°C the permitted internal heat gains are

probably at such low values which would be practically impossible to reach in

commercial buildings with reasonable efforts. The upper values of plausible design wet

bulb temperatures for this kind of comfort cooling system has however not been

investigated in this thesis.

148

Parameter variations

In section 7.2.3 Multi parameter variations, variations with more than one parameter at

a time have been performed. The number of parameter combinations have however

been kept at reasonable levels to limit the number of simulation cases.

One parameter analysis includes full variation of five locations/climates, two thermal

capacities and three levels of design internal heat gain. The locations are translated to

design wet bulb temperatures, thus allowing the maximum room air temperature to be

plotted against the design wet bulb temperature at different combinations of design

internal heat gain and thermal capacity.

The diagrams discussed above and presented in section 7.2.3 Multi parameter

variations can be used in early stages of the design phase of a building to get a first

indication of probable maximum indoor air temperatures, alternately the probable

percentage of working hours when the indoor temperature exceeds 24°C at given

conditions, named %24°C. Alternately, the diagrams can be used to get indications of at

which conditions a required maximum indoor temperature or maximum value of %24°C

can be achieved.

The other diagrams in section 7.2.3 Multi parameter variations, are presenting annual

duration curves of the indoor air temperature at three different locations;

Stockholm/Arlanda, London/Gatwick and Berlin at the conditions “Best case” and

“Worst case”. Best case is defined as when all the examined parameters are adjusted to

values equal to “Lower temperature”. Worst case is defined as when all the examined

parameters are adjusted to values equal to “Higher temperature”.

These diagrams give indications of the span of the annual duration of the indoor air

temperature between the conditions “Best case” and “Worst case”. At “Best case” the

indoor air temperature are kept at the set point temperature, i.e. 21°C, for a long period

of time and maximum temperatures were kept below about 23°C. This is a performance

equal to a conventional cooling system with high capacity and fairly strict temperature

control.

At “Worst case”, the annual duration of the indoor air temperature is of course at

somewhat higher level. However, the results in this case are in accordance with normal

and usually acceptable design limits concerning indoor temperature duration, i.e. a

duration of the indoor air temperature where the percentage of working hours when the

indoor temperature exceeds 25°C is between 5 to 10%. The maximum occurring indoor

temperatures where between 28 - 30°C, however occurring for a very limited time.

9.2.3 Investments and annual costs

An analysis of investment costs for different types of cooling towers compared with

investment costs for conventional chiller has not been made in this thesis. Sources in the

literature discussing investment costs for different types of cooling towers are extremely

rare to find.


149

However in one reference, Facão (n.d.), which is a (unofficial) homepage where the

EcoCool project is presented, a table with investments and annual costs is presented, see

table 9.1. The EcoCool project is described in section 1.3 Previous work.

According to Facão (n.d.), the first two column in the table presents investment costs for

cooling equipment, i.e. chiller or cooling tower, and the chilled ceiling system including

distribution system and room cooling equipment. The third column is the sum of the

first two. The last column presents total annual costs, which includes investment costs

with a depreciation period of 5 years together with annual costs for energy and

maintenance. The data is from Switzerland and from about year 2000. There is no

explanation to what area the numbers in the table are related to; [Euro/m2]

Table 9.1 Comparison between investment costs for different kinds of comfort

cooling systems in the EcoCool project, Facão (n.d.).

System

Cooling

equipment

[Euro/m2]

Chilled

ceiling

[Euro/m2]

Investment

cost

[Euro/m2]

Total annual

costs

[Euro/m2]

1) Ecocool

(closed cooling

tower)

35 155 190 45

2) Ecocool (open

cooling tower) 17 155 172 41

3) Chilled ceil.

with refrig. mach. 22 155 177 43

4) Conventional

AC (all-air) - - 255 61

The total investment cost for the EcoCool system number one are 7% above that of a

system with chilled ceiling and a conventional chiller (refrigerant machine), i.e. system

number 3, and EcoCool system number 2 are 3% below the mentioned system.

The distribution system and room cooling equipment for the first three systems in table

9.1 are equal, hence the investment costs for that part of the entire cooling system are

equal. The investment cost for a conventional all-air air conditioning system is

considerably higher than the first three systems, i.e. about 40% higher. When comparing

the first three systems, the total annual costs show no significant differences.

The figures in table 9.1 indicate that there are small differences in total investment costs

for hydronic systems whether they are chilled with a conventional chiller or a cooling

tower (open or closed type). The investment cost for the open type cooling tower is

however about 50% of the investment cost for a closed type tower. The relative share of

the investment cost for the cooling towers is about 10 – 20% of the total investment

cost.

150

It should be noted that the cooling towers in the EcoCool project were prototypes, hence

the investment costs are most likely on the higher end. If the towers were mass

fabricated the price would have been significantly lower.

9.2.4 Applicability

Each low energy cooling system has it advantages and disadvantages. This is of course

true also for the cooling system investigated in this thesis. Among the advantages for

hydronic cooling systems with a cooling tower there is one which stands out; namely

the applicability of the system. Some of the alternative cooling systems mentioned in

section 2.2 Low energy cooling alternatives can only be applied to a new building, such

as hollow core slab cooling with air or slab cooling with water. A few of the alternatives

require access to the ground or an ambient water resource in the vicinity. This

requirement can often be difficult to fulfil in the central parts of a city where most

commercial buildings are located.

A hydronic cooling system with a cooling tower can be applied to both new and

refurbished buildings. In Sweden, it is common to replace an old air conditioning

system distributing chilled air with a hydronic cooling system equipped with chilled

beams. Hydronic systems with chilled beams are now the predominant cooling system

in Sweden and all actors in the building industry consider this system as well

established and accepted.

Using this type of system in conjunction with a cooling tower, instead of a conventional

chiller, requires only access to the ambient air. There is no need for access to the ground

or any ambient water resources. The space required for the location of the cooling

tower, normally on roofs or adjacent to the building, equals approximately the space

required for the cooling of condenser heat from a conventional chiller. All parts of this

cooling system represent well-established techniques. No parts of the system are new or

unproven on the market.

All these advantages, together with moderate investment costs, low annual costs and the

environmental advantages, makes this cooling system an attractive alternative to

conventional cooling.

9.2.5 Miscellaneous

Legionella

The minimal risk of growth of legionella bacteria in cooling towers applied as a sole

free cooling source in a hydronic cooling system is discussed in section 4.4 Legionella

in evaporative cooling towers.

There is however a great concern for the threat of the legionella bacteria from cooling

towers among both the public and HVAC engineers. In that context, it may be daring to

introduce a cooling tower in a low energy comfort cooling system where the tower

would be located adjacent to or on the roof of commercial buildings. If this system is


151

introduced in several buildings in city centres, there will be several cooling towers close

to where numerous people are working and living.

The risk for growth of legionella bacteria is close to zero in this application of a cooling

tower as shown in section 4.4 Legionella in evaporative cooling towers. It is equal to

the risk of legionella growth in public outdoor swimming pools, shallow ponds or

public fountains, which also are close to where people live or work. Despite this, there

still might be concerns only because it is a cooling tower. To many people, cooling

towers are dangerous per definition. This is a concern which must be addressed.

Cooling tower size

When discussion prejudiced thinking, there might be people associating cooling towers

with those in huge power plants at heights of 30 meters and more, exhausting enormous

fog clouds. A small contribution against this thinking, when applying cooling towers to

hydronic comfort cooling systems in commercial buildings, can be made by presenting

pictures of the cooling tower used in the EcoCool project, se figure 9.4.

Figure 9.4 Picture of no. 2 prototype cooling tower in the EcoCool project, Facão

(n.d.).

The cooling tower in figure 9.4 has a design cooling capacity of 10 kW. The cross

section of the cooling tower measures 0,5 m × 1,23 m with a height of 1,74 m. Required

cooling capacities of normal commercial buildings are of course normally higher than

10 kW, usually in the range from 10 to about 300 kW. The intention with figure 9.4 is

that the reader will get an idea of the magnitude of a cooling tower with cooling

capacities close to normal for commercial buildings. In most cases, a cooling tower for

cooling a commercial building is however somewhat bigger than in figure 9.4, but not

more than about a few times bigger.

152

9.3 Further research

Although a fair amount of new knowledge and information is presented in this thesis,

several topics or questions can be further investigated;

� There is very little information from real life buildings with hydronic comfort

cooling systems in conjunction with a cooling tower. A pilot plant has been

evaluated in this thesis but further measurements on a real full size object would

give important information about the performance of this kind of cooling system

together with the indoor climate it can provide.

� The described comfort cooling system could possibly benefit from using thermal

storage in the system, e.g. by using water tanks to store chilled water. The chilled

water can be produced during nighttime when the ambient wet bulb temperature is

lower. This would most certainly reduce the maximum indoor air temperature. A

thermal storage can reduce the size of the cooling tower hence reduce investment

costs.

� The economics of a hydronic comfort cooling system with a cooling tower needs

to be further investigated. The different types of cooling towers, open and closed,

have different investment costs and life cycle costs. How big are the differences?

How do variations in cooling tower design parameters affect investment costs and

life cycle costs? Is there an optimum combination of design wet bulb temperature,

total approach and design supply and return temperature of chilled water, i.e. the

size of the chilled beams? Is it an important question or could we care less?

153

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NMF version 3.02, ASHRAE RP – 839, report from Building Sciences, KTH, Stockholm, Sweden. 1996.

Santamouris M. (coord.), (1995), Natural cooling techniques – Design methodology

and application to southern Europe, Pascool Final Report, European Commission: Directorate General XII for Science Research and Development, 1995

Santamouris M., Asimakopoulos D., (ed), (1996), Passive cooling of Buildings, James & James (Science Publishers), London UK, 1996, ISBN: 1 873936 47 8

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SCANVAC, VVS-tekniska föreningen, Förlags AB VVS, Stockholm 2000. ISBN: 91-973834-3-0 (in Swedish)

Sodec, F. (1999), Economic viability of cooling ceiling systems, Energy and Buildings, 30 (1999) pp 195 – 201.

Sprecher P., Tillenkamp F., (2003), Energy saving systems in building technology

based on concrete-core-cooling, International Journal of Ambient Energy 24 (1) (2003) pp 29 – 34.

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160

161

APPENDIX A

The Cooling Tower Model (CTM) in NMF code

In this appendix the cooling tower model in NMF code is presented. For information about

the Neutral Model Format, see Sahlin et al. (1989) and Sahlin (1996).

CONTINUOUS_MODEL Cooling_Tower

ABSTRACT "Cooling Tower

INTRODUCTION

This model describes a general device for cooling a liquid with ambient air.

The model can be used for simulation of cooling towers, open and closed

circuit type, wet or dry type but also any air/liquid heat exchanger with

above mentioned purpose. The model implies that cooling is supported by one

or several fans inducing forced draft.

The model is based on the cooling tower model described in ASHRAE Primary HVAC

Toolkit [1]. The ASHRAE model is based on the effectiveness-NTU method which makes

it possible to simulate any heat exchanging device with the above mentioned

purpose. The ASHRAE model is supplemented with different modes and different

control strategies for part load control. This makes this model more general for

different applications for cooling liquids with ambient air in HVAC systems. In

the following text the cooling device is called 'cooling tower' for practical reasons.

GENERAL DESCRIPTION

The model includes a freeze protected circuit, Circuit 1 (primary circuit),

which is a closed loop between the cooling tower and a heat exchanger. The heat

exchanger separates Circuit 1 from Circuit 2. Circuit 2 (secondary circuit)

is the one supplying chilled liquid (water) to the air cooler in AHU and

the cooling device(s) in the zone(s).

The liquid in Circuit 1 is called Liquid 1 and subsequently the liquid

in Circuit 2 is called Liquid 2.

If the user wants to omit the freeze protected circuit the following parameters

is set to:

cpLiq1 = 4187 (heat capacity for water, which is normally used)

rhoLiq1 = 999 (density for water, which is normally used)

TApproachS = 0 (the same as no heat exchanger or one with infinite capacity)

MODES

The cooling tower can be set to different modes by the user.

- Mode_WetDry; The user can choose between wet or dry operation mode.

The wet mode implies the liquid to be cooled (liquid 1) is sprayed over a

cooling tower fill (open mode) or water spraying over the closed circuit

tube bundles containing the liquid to be cooled (closed mode).

- Mode_Fan; if the fan(s)is/are single speed or equipped with

variable speed drive. The single speed mode is applicable when cooling

tower has only one fan (or maybe two fans) with single speed. If the

cooling tower has a two speed fan or several fans with single or two

speed operation the user should choose the mode with variable speed drive.

This mode simulates the case with multiple fans cycling on-off in sequence

fairly well. If cooling tower is equipped with variable speed drive the

choice of mode is obvious.

- Mode_ClOp; for choice of open circuit or closed circuit cooling tower. The

mode controls if spray water is on or off. This mode is added because the

162

combination of Closed circuit mode and Wet mode use spray water. All other

combinations of Open/Closed and Wet/Dry modes use no additional spray water.

- Mode_Tower; lets the user choose between a counterflow configuration or a

cross flow configuration. The cross flow configuration assumes both streams unmixed.

CONTROL

The cooling tower has three different ways of control.

- First way of control is a simple on/off control of the whole tower via an

external signal from a time schedule through a link (Tower_control).

- Secondly the leaving liquid 2 temperature is controlled in two sequences;

First by modulating the massflow of air through the fan(s). This is done

via an external signal from an external controller through a link (Fancontrol).

In single speed mode the real life cycling on-off is simulated as a linear

relation of the mass flow from a minimum flow (15% of max.) up to maximum flow.

In variable speed mode the fan laws are applied controlling the mass flow.

- Last step is to cycle the fan(s) on and off when the ambient air temperature

is below the limit for modulating the fan(s).

USER INPUT

The user must give the following design stage parameters as input:

Input parameters affecting size and cooling capacity of the cooling tower:

QTower_d 'Design cooling capacity in tower (in kW)'

T_ApproachP 'Temperature difference between inlet air (TAir_in_d) and outlet liquid 1 (TLiq1Cold_d) in primary circuit

at design stage'

TRange_P 'Difference between inlet liquid 1 (TLiq1Warm_d) and outlet liquid 1 (TLiq1Cold_d) in primary circuit at

design stage'

TAir_indb_d 'Design inlet dry bulb air temperature'

RelHum_d 'Design relative humidity of ambient air (%)'

MLiq1MAir_d 'Massflow relation of liquid 1 and Air (MLiq1_d/MAir_d).Normally in range 0.5 to 2'

Input parameters primarily affecting energy efficiency of the cooling tower:

EtaFan_d 'Fan efficiency (in tower) at design air flow rate'

dpFan_d 'Pressure difference over fan at design rate massflow of air in tower'

dpLiq1 Pressure difference in circuit 1 (liquid 1), assumed constant'

dpLiq2 Pressure difference in circuit 2 (liquid 2 passing heat exchanger), assumed constant'

dpSpray 'Pressure difference in spray water circuit (open) assumed constant'

Mode_Fan 'Control mode; 0=Fan with variable speed control, 1=Single speed fan'

EtaLiq1 'Total pump efficiency in circuit with liquid 1'

EtaLiq2 'Total pump efficiency in circuit with liquid 2'

EtaSpray 'Total pump efficiency in (open) circuit with spray water'

Miscellaneous input parameters:

EtaHex 'Heat exchanger efficiency (between circuit 1 and 2)'

Min_MAir 'Min. airflow in fraction of design rate massflow of air in tower. When using variable speed drive

Min_MAir is normally in range 0.1 to 0.2'

cpLiq1 'Liquid 1 specific heat (default 3685= 40% propylene glycol at 15°C)'

rhoLiq1 'Liquid 1 density (default 1039= 40% propylene glycol at 15°C)'

Mode_WetDry 'Control mode; 0=Wet mode, 1=Dry mode'

Mode_Spray 'Control mode; 0=Closed circuit tower, 1=Open circuit tower'

Mode_Tower 'Control Mode; 0=Counterflow tower, 1=Crossflow tower'

LIMITATIONS

- The model does not calculate the consumption of makeup water for evaporation,

drift losses and blow down.

- Hybrid cooling towers with combination of, or alteration between, wet and dry mode

Appendix A

163

can not be simulated in this model.

- Natural draft cooling towers can not be simulated in this model.

- The model works with good accuracy between -20 to 40 C wet bulb temperature

of ambient air. It should not be used outside this temperature span.

Refereces

[1] ASHRAE Primary HVAC Toolkit

[2] ASHRAE/IESNA Standard 90.1-1999

[3] Simulation model - Finned water-to-air coil without condensation, LBNL-42355, 1998 (model in SPARK)

[4] Stoecker W.F, Industrial refrigeration handbook, p. 266, McGraw-Hill 1998

[5] Lebrun J., Silva C., Cooling Tower model and experimental validation, ASHRAE Trans. 2002

Date: June 19, 2003

By: Bengt Bergsten

CIT Energy Management

Call: Enthal from PSYCHRO1.NMF

Humrat from PSYCHRO2.NMF

Wetbulb from PSYCHRO3.NMF

Revisions:

030625 BB Mode with both counterflow and cross flow configuration added

030826 BB Equation for calculating hAir_out changed for better numerical stability when calculating cpAir_e

030828 BB Original equation for cpAir_e replaced by curve fit equation. Original equation not numerically stable

around 0 deg C.

031012 BB A tank (thermal mass) is added. Event Fan_on removed. Better stability.

080909 BB Minor changes in Eff-NTU equations + thermal mass in MAir. ”

EQUATIONS

/**************** Psychrometics of ambient air ***********************/

/* calculate wet and dry bulb temperatures of entering air stream */

TAir_wb := Wetbulb(TAir, HumAir) ;

TAir_inwb := IF LINEARIZE (1)

THEN TAir - 3

ELSE /* limits due to cpAir_e regression equation */

IF TAir_wb < -20

THEN -20

ELSE_IF TAir_wb > 40

THEN 40

ELSE TAir_wb

END_IF

END_IF ;

TAir_indb := TAir ;

/********************* Control 1 ******************************/

/* Control of leaving liquid 2 temperature, TOut, via air flow in cooling tower fan.

The control is managed by an external controller in IDA ICE Primary systems */

TLiq2Out_r = TLiq2Out_set ;

164

/* Further control equations is placed under the heading "Control 2" */

Appendix A

165

/* Limiting the incoming control signal */

Ctrl := IF LINEARIZE (1)

THEN 0.5

ELSE

IF Ctrl_In > 1

THEN 1

ELSE_IF Ctrl_In < 0

THEN 0

ELSE Ctrl_In

END_IF

END_IF ;

/* When Fan is on, the air flow is regulated as below depending if fan is in variable speed mode

or single speed mode. In single speed mode the real life control strategy by cycling on-off is simulated

as a linear relation of MAir from MAir_d to 15% of MAir_d [2]. For fan with variable

speed control MAir is linear between MAir_d and a minimum flow equal to

Min_MAir * MAir_d (Min_MAir is typically 0.1 to 0.2) */

MAir + 60*MAir' = IF NINT(TowerOn) == 0 /* Tower is off */

THEN MAir_0 * MAir_d /* A minimum level for numerical stability reasons */

ELSE MAir_d * Ctrl + MAir_0 * MAir_d * (1 - Ctrl)

END_IF ;

/* MLiq1 is regulated in the same way as MAir. Minimum flow is set to 30%

of MLiq1_d. For an open cooling tower the minimum flow must not be to small

so the distribution of the cooling water is sufficient on the cooling tower fill. */

MLiq1 := IF NINT(TowerOn) == 0 /* Tower is off */

THEN MLiq1_0 * MLiq1_d /* A minimum level for numerical stability reasons */

ELSE MLiq1_d * Ctrl + MLiq1_0 * MLiq1_d * (1-Ctrl)

END_IF ;

/********************** Miscellaneous ***********************/

/* Pressure difference. Inlet pressure is given as reference pressure for massflow

circuit. This corresponds to the grounding of an electrical circuit, or an

expansion vessel for a fluid circuit. The outlet pressure is at present a

S_P parameter (dpLiq2) */

pLiqOut = dpLiq2 ;

pLiqIn = 0 ;

/* total massflow in liquid 2 (help variable) going through the heat exchanger. */

MLiq2 := M1 + M2 ;

/* Entering liquid 2 average temp. */

TLiq2In := IF MLiq2 <= 0

THEN (TIn1 + TIn2)/2

ELSE (M1 * TIn1 + M2 * TIn2) / MLiq2

END_IF ;

166

/************************** Cooling tower ******************************/

/* Equations based on model in ref. [1]. A dry operation mode is added */

/* Help variables for equation cpAir_e */

Td := ABS(TAir_outwb - TAir_inwb) ; /* difference of in- and outgoing wet bulb air temp */

Tm := ((TAir_outwb + TAir_inwb)/2) + 21 ; /* mean value of in- and outgoing wet bulb air temp */

/* The effective specific heat of air in cooling tower. */

/* The original equation from ref.[1] cpAir_e = (hAirOut - hAirIn)/(TAir_outwb - TAir_inwb)

is replaced by a curve fit equation where cpAir_e = f(Tm, Td) of numerical reasons.

The curve fit equation has a relative error <1% between -15<Tm<40 °C and Td<15°C.

This error generates an error on the leaving liquid 2 temperature which is less than 0.1°C.

Increasing to the maximum range allowed; -20<Tm<40 °C and Td<30°C the relative error is <3.3%. */

cpAir_e := IF Mode_WetDry >= 0.5

THEN cpAir /* Cooling Tower in dry mode */

ELSE_IF LINEARIZE (1)

THEN cpAir + 1810

ELSE

6.56473E-4*ABS(Tm)*Tm**3 - 0.0351*Tm**3 + 1.65114*Tm**2 + 1196.30745 +

1.84999E-4*(Td**2)*(Tm**2) + 2.09962E-11*(ABS(Td)*Td**3)*(ABS(Tm)*Tm**3)

/* Cooling Tower in wet mode */

END_IF ;

/* Calculating the UA value at part load */

/*The power law relation between UA and MAir/MAir_d and MLiq1 / MLiq1_d with

corresponding exponents n2 and n3 is a correlation based on ref. [3] and [5]

The presented values of n2 and n3 from other references in [3] and [5] shows

a big variation. With n2=0.65 and n3=0.43 as default values they represent

mean values of n2 and n3 from the ones presented in [3] and [5]. */

UA := IF LINEARIZE (1)

THEN UA_d * (MAir / MAir_d) * (MLiq1 / MLiq1_d)

ELSE_IF MAir <= 1E-3

THEN UA_d * (MAir / MAir_d) * (MLiq1 / MLiq1_d)

ELSE UA_d * (MAir / MAir_d)**n2 * (MLiq1 / MLiq1_d)**n3

END_IF ;

/* The effective heat transfer coefficient-area product */

UA_e := UA * cpAir_e / cpAir ;

/* Heat capacity flow of air, liquid 1 and liquid 2 */

CLiq1 := cpLiq1 * MLiq1 ;

CLiq2 := cpLiq2 * MLiq2 ;

CAir := cpAir_e * Mair ;

/* Smallest and biggest heat capacity flow */

CMin := max(1E-6, min(Cliq1, CAir));

CMax := max(CLiq1, CAir) ;

/* Relation Cmin/Cmax */

CRatio := IF CMax < 1E-6

THEN 1

ELSE max(1E-4, CMin) / CMax

END_IF ;

Appendix A

167

/* Number of transfer units in tower */

NTU := IF CMin <= 0

THEN 1E-4

ELSE UA_e / CMin

END_IF ;

/* The effectiveness of the heat exchanger */

k := NTU**(-0.22) ; /* help variable in cross flow equation */

Eff := IF LINEARIZE (1)

THEN 0.5 * NTU

ELSE_IF Mode_Tower < 0.5

/* Counterflow configured cooling tower: */

THEN (1 - EXP(- NTU * (1 - CRatio))) / (1 - CRatio * EXP(- NTU * (1 - CRatio)))

/* Crossflow configured cooling tower: */

ELSE 1 - EXP((EXP(-NTU * CRatio * k) - 1 ) / (CRatio * k))

END_IF ;

/* Water-air heat transfer rate in cooling tower */

QTower := IF NINT(TowerOn) == 0 /* Cooling Tower is off */

THEN 0

ELSE_IF Mode_WetDry < 0.5

/* Cooling Tower is on and in wet mode: */

THEN Eff * CMin * (TLiq1Warm - TAir_inwb) / 1000

/* Cooling Tower is on and in dry mode: */

ELSE Eff * CMin * (TLiq1Warm - TAir_indb) / 1000

END_IF ;

/* Outlet air dry/wet bulb temperature from tower */

TAir_outdb = IF LINEARIZE (1)

THEN TAir + 2

ELSE_IF Mode_WetDry < 0.5 /* Cooling Tower is in wet mode */

THEN TAir_outwb /* is good approx. when state of air is close to saturated */

ELSE TAir_indb + (QTower * 1000 / CAir) /* Cooling Tower is in dry mode */

END_IF ;

TAir_outwb = IF LINEARIZE (1)

THEN TAir + 2

ELSE_IF Mode_WetDry < 0.5 /* Cooling Tower is in wet mode */

THEN TAir_inwb + (QTower * 1000 / CAir)

ELSE Wetbulb(TAir_outdb, HumAir) /* Cooling Tower is in dry mode */

END_IF ;

/* Outlet liquid 1 (cold side) temperature from tower */

TLiq1Cold := TLiq1Warm - (QTower * 1000 / CLiq1) ;

/********************* Heat exchanger and tank ******************************/

/* Outlet temperature of liquid 2 (help variable) from heat exchanger between liquid 1 and liquid 2 */

TLiq2Out = IF LINEARIZE (1)

THEN TAir + 4

ELSE

IF NINT(TowerOn) == 0 /* Cooling Tower is off */

THEN TLiq2In

ELSE_IF CLiq2 <= CLiq1

THEN TLiq2In - EtaHex * (TLiq2In - TLiq1Cold)

ELSE TLiq2In - (CLiq1 / CLiq2) * EtaHex * (TLiq2In - TLiq1Cold)

END_IF

END_IF ;

168

/* Outlet temperature of liquid 1 (TLiq1Warm) from heat exchanger between liquid 1 and liquid 2 */

TLiq1Warm = IF LINEARIZE (1)

THEN TAir + 3

ELSE TLiq1Cold + (CLiq2 / CLiq1) * (TLiq2In - TLiq2Out)

END_IF ;

/* Tank after heat exchanger in secondary circuit */

T_TankIn := IF CTRL_T_TankIn < 0.5

THEN

IF TLiq2Out <= TLiq2Out_set

THEN TLiq2Out_set

ELSE TLiq2Out

END_IF

ELSE

IF NINT(Warm) == 0 /* check for not heating the liquid: */

THEN TLiq2In

ELSE_IF NINT(Cold) == 0 /* check the liquid is not below set point temp: */

THEN min(TLiq2Out_set, TLiq2In)

ELSE TLiq2Out /* TLiq2Out <= TLiq2In OR TLiq2Out >= TLiq2Out_set: OK! */

END_IF

END_IF ;

M_Tank * T_Tank' = MLiq2 * (T_TankIn - T_Tank) ;

/********************* Use of energy ******************************/

/* Speed of fan with variable speed control */

n_speed := IF MAir < MAir_0 * MAir_d * 1.01

THEN 0 /* Fan(s) assumed to be off */

ELSE_IF MAir < Min_MAir * MAir_d

THEN n_speed_d * Min_MAir * MAir_d / Mair_d

ELSE n_speed_d * MAir / Mair_d

END_IF ;

/* Pressure difference in cooling tower fan (through fan laws) */

dpFan := IF Mode_Fan > 0.5

/* Fan with single speed mode: */

THEN dpFan_d

/* Fan with variable speed control mode: */

ELSE_IF LINEARIZE (1) OR (n_speed / n_speed_d) < 1E-3

THEN dpFan_d * (n_speed / n_speed_d)

ELSE dpFan_d * (n_speed / n_speed_d)**2 /* according to fan laws */

END_IF ;

/* Total fan efficiency as a function of fan speed. */

/* The relation fan efficiency - fan speed is chosen as a power law expression which is an approximate

relation for the fact that total fan efficiency is not constant when a variable speed drive is used.

The total energy usage in the tower is dominated by energy supplied to the fan(s). For single speed fan

'EtaFan' is constant */

EtaFan := IF Mode_Fan > 0.5 /* Fan with single speed mode: */

THEN EtaFan_d /* Fan with variable speed control mode: */

ELSE_IF LINEARIZE (1)

THEN EtaFan_d * (MAir / Mair_d)

ELSE_IF MAir < Min_MAir * MAir_d

THEN EtaFan_d * (Min_MAir * MAir_d / Mair_d)**n1

ELSE EtaFan_d * (MAir / Mair_d)**n1

END_IF ;

Appendix A

169

/* Electrical power for fan in cooling tower */

PFan := IF NINT(TowerOn) == 0 OR MAir < MAir_0 * MAir_d * 1.01 /* Cooling Tower is off OR

Fan(s) is off */

THEN 0

ELSE_IF Mode_Fan < 0.5

/* Fan with variable speed control mode: */

THEN dpFan * MAir / (EtaFan * rhoAir)

/* Fan with single speed mode: */

ELSE_IF MAir < 0.15 * MAir_d

THEN 0

ELSE

(MAir * dpFan_d / (0.85 * EtaFan * rhoAir)) - (0.15 * Mair_d * dpFan_d / (0.85 * EtaFan * rhoAir))

/* This linear relation corresponds to relation in [2] */

END_IF ;

/* Electrical power for pump in circuit 1 between cooling tower and heat exchanger, i.e. liquid 1 */

PLiq1 := IF NINT(TowerOn) == 0 OR MLiq1 < MLiq1_0 * MLiq1_d * 1.01 /* Cooling Tower is off */

THEN 0

ELSE dpLiq1 * MLiq1 / (EtaLiq1 * rhoLiq1)

END_IF ;

/* Electrical power for pumping liquid 2 through heat exchanger in circuit 2 */

PLiq2 := IF NINT(TowerOn) == 0 /* Cooling Tower is off */

THEN 0

ELSE dpLiq2 * MLiq2 / (EtaLiq2 * rhoLiq2)

END_IF ;

/* Electrical power for pump in water spray sump. */

/* Equation is based on approximate relation between MSpray and QTower [4]

which is 0.018 l/s per kW cooling capacity in tower */

VSpray := IF Mode_WetDry < 0.5 AND Mode_ClOp < 0.5

/* Cooling Tower is in wet mode AND closed circuit tower mode */

THEN (0.018 / 1000) * QTower_d /* QTower_d in kW */

ELSE 0

/* All other combinations of Mode_WetDry and Mode_ClOp */

END_IF ;

PSpray := IF NINT(TowerOn) == 0 /* Cooling Tower is off */

THEN 0

ELSE dpSpray * VSpray / EtaSpray

END_IF ;

/* COP for cooling tower */

COP := IF NINT(TowerOn) == 0 /* Cooling Tower is off */

THEN 0

ELSE QTower * 1000 / (PFan + PLiq1 + PLiq2 + PSpray)

END_IF ;

/*Total electrical power demand for cooling tower */

PTot := IF NINT(TowerOn) == 0 /* Cooling Tower is off */

THEN 0

ELSE (PFan + PLiq1 + PLiq2 + PSpray)

END_IF ;

170

/********************* Control 2 ******************************/

/* Operating mode TowerOn */

TowerOn := IF Event(G2, TowerCtrl) > 0.5

THEN 1

ELSE 0

END_IF ;

/* Control of leaving liquid 2 temperature, TOut, via air flow in cooling tower fan */

TOut = IF LINEARIZE (1)

THEN TAir + 4

ELSE T_Tank

END_IF ;

/* check for not heating the liquid at T_TankIn : */

Warm := IF Event(G1, TLiq2Out - TLiq2In) > 0

THEN 0 /* TLiq2Out > TLiq2In : Not OK! */

ELSE 1

END_IF ;

/* check for not cooling the liquid beneath set point temp. at T_TankIn: */

Cold := IF Event(G3, TLiq2Out - TLiq2Out_set) < 0

THEN 0 /* TLiq2Out < TLiq2Out_set : Not OK! */

ELSE 1

END_IF ;

/* The rest of control equations are placed under the heading "Control 1" */

/****************************** End of Equations ************************************************/

LINKS

/* type name variables .... */

T T_Ambient TAir ; /* ambient outdoor air temperature */

W W_Ambient HumAir ; /* ambient outdoor absolute humidity */

PMT Inlet1 PLiqIn, POS_IN M1, TIn1 ; /* from AHU */

PMT Inlet2 PLiqIn, POS_IN M2, TIn2 ; /* from Zone(s) */

PMT Outlet1 PLiqOut, POS_OUT M1Out, TOut ; /* to AHU */

PMT Outlet2 PLiqOut, POS_OUT M2Out, TOut ; /* to Zone(s) */

ControlLink Fan_Control Ctrl_In ;

ControlLink Tower_Control TowerCtrl ;

T Temp_Setpoint TLiq2Out_r ; /* Liquid 2 leaving temperature set point */

T Temp_Measure TLiq2Out ; /* Liquid 2 leaving temperature */

VARIABLES

/* type name role def min max description */

Pressure pLiqOut OUT 15000 0 BIG "Outlet liquid 1 pressure"

Pressure pLiqIn OUT 0 0 BIG "Inlet liquid 1 pressure"

Pressure dpFan LOC 75 SMALL BIG "Pressure difference over fan"

MassFlow M1 IN 0.30 0 BIG "Inlet massflow of liquid from air cooler in

AHU"

MassFlow M2 IN 0.15 0 BIG "Inlet massflow of liquid from room cooler(s)

in zone(s)"

Appendix A

171

MassFlow M1Out IN 0.30 0 BIG "Inlet massflow of liquid to air cooler in

AHU"

MassFlow M2Out IN 0.15 0 BIG "Inlet massflow of liquid to room cooler(s) in

zone(s)"

MassFlow MLiq1 LOC 0.45 SMALL BIG "Massflow of liquid 1"

MassFlow MLiq2 LOC 0.45 SMALL BIG "Total massflow of liquid 2, help variable"

MassFlow Mair LOC 0.89 SMALL BIG "Actual massflow of air in tower"

VolFlow VSpray LOC 1.4E-4 SMALL BIG "Volume flow of spray water used in wet and

closed mode"

Temp TAir IN 15 -30 50 "Ambient air dry bulb temperature"

Temp TOut OUT 19 -30 50 "Final outlet liquid 2 temperature"

Temp TIn1 IN 22 -30 50 "Inlet temp. of liquid from air cooler"

Temp TIn2 IN 22 -30 100 "Inlet massflow of liquid from room cooler"

Temp TLiq2Out OUT 19 -30 100 "Outlet 2 liquid temperature"

Temp TLiq2In LOC 22 -30 100 "Average temp. of inlet liquid 2"

Temp TLiq1Cold LOC 17.5 -30 100 "Cold side liquid 1 temp."

Temp TLiq1Warm OUT 20.5 -30 100 "Warm side liquid 1 temp."

Temp TAir_indb LOC 15 -30 50 "Inlet dry bulb temp. of amb. air (=TAir)"

Temp TAir_outdb OUT 16 -30 50 "Outlet dry bulb temp. from tower"

Temp TAir_inwb LOC 10 -20 40 "Inlet wet bulb temp of amb. air"

Temp TAir_outwb OUT 16 -20 40 "Outlet wet bulb temp. from tower"

Temp TLiq2Out_r OUT 16 -20 100 "Outlet liquid 2 requested temperature"

Temp Td LOC 4 SMALL 50 "Wet bulb temperature difference"

Temp Tm LOC 13 -20 60 "Wet bulb mean temperature"

Temp T_Tank OUT 19 -30 100 "Temperature in Liquid 2 tank"

Temp T_TankIn LOC 19 -30 100 "Temperature entering Liquid 2 tank"

Temp TAir_wb LOC 10 -20 40 "Help variable: Ambient wet bulb

temperature"

HumRatio HumAir IN 0.006 0 BIG "Humidity ratio ambient air

(kg water/kg dry air)"

Control TowerCtrl IN 1 0 1 "Control signal; 0 = Tower Off 1 = Tower On"

Control Ctrl_In IN 0.5 0 1 "Fan control in signal, help variable"

Control Ctrl LOC 0.5 0 1 "Fan control, 0 = Min. massflow

1 = Max. massflow "

Generic Warm A_S 1 0 1 "Is liquid 2 heated? 0 = Yes, 1 = No"

Generic Cold A_S 1 0 1 “Is liquid 2 cooled beneath set point? 0 = Yes,

1 = No"

Generic TowerOn A_S 1 0 1 "Tower operating mode 0 = Off, 1 = On"

Generic G1 A_S 1 -BIG BIG "G-stop Warm"

Generic G2 A_S 1 -BIG BIG "G-stop TowerOn"

Generic G3 A_S 1 -BIG BIG "G-stop Cold"

Factor Eff LOC 0.56 SMALL 1 "Tower efficiency"

Factor UA_e LOC 925 SMALL BIG "Effective heat transfer coeff.-area product in

tower"

Factor UA LOC 350 SMALL BIG "Heat transfer coeff.-area product in tower"

Factor Cratio LOC 0.55 SMALL 1 "Relation Cmin/Cmax"

Factor CMin LOC 924 SMALL BIG "CMin = min ( CLiq1, CAir )"

Factor CMax LOC 1666 SMALL BIG "CMin = max ( CLiq1, CAir )"

Factor CLiq1 LOC 1666 SMALL BIG "Heat capacity flow of liquid 1"

Factor CLiq2 LOC 1884 SMALL BIG "Heat capacity flow of liquid 2"

Factor CAir LOC 924 SMALL BIG "Heat capacity flow of air "

Factor NTU LOC 1 SMALL BIG "Number of transfer units in tower"

Factor COP LOC 5 SMALL BIG "Overall coefficient of performance (efficiency)"

Factor n_speed LOC 0.5 SMALL BIG "Fan speed (rpm) at massflow rate of air in

tower"

Factor EtaFan LOC 0.42 SMALL 1 "Fan efficiency (in tower) as a function of MAir"

Factor k LOC 0.6 SMALL BIG "Help variable when calculating Eff for a cross

flow tower"

172

HeatFlux_k Qtower LOC 5 SMALL BIG "Heat transfer rate in tower (in kW)"

HeatCapM cpAir_e LOC 2657 SMALL BIG "Effective specific heat of air in tower "

ElPowerCons Pfan LOC 52 SMALL BIG "Supplied electric power to cooling tower fan"

ElPowerCons PLiq1 LOC 9 SMALL BIG "Supplied electric power to pump in liquid 1"

ElPowerCons PLiq2 LOC 3 SMALL BIG "Supplied electric power to pump in liquid 2"

ElPowerCons Pspray LOC 2.2 SMALL BIG "Supplied electric power to spray water pump"

ElPowerCons Ptot LOC 63 SMALL BIG "Total supplied electrical power to cooling

tower"

PARAMETERS

/* type name role def min max description */

Factor EtaFan_d S_P 0.6 0.01 1 "Fan efficiency (in tower) at design air flow rate"

Factor EtaHex C_P 0.67 0.01 1 "Heat exchanger efficiency (between circuit 1

and 2)"

Factor EtaLiq1 S_P 0.5 0.01 1 "Total pump efficiency in circuit with liquid 1"

Factor EtaLiq2 S_P 0.5 0.01 1 "Total pump efficiency in circuit with liquid 2"

Factor EtaSpray S_P 0.5 0.01 1 "Total pump efficiency in (open) circuit with

spray water"

Factor UA_ed C_P 1928 SMALL BIG "Design rate of effective heat transfer coeff.-area

product in tower"

Factor UA_d C_P 531 SMALL BIG "Design rate of heat transfer coeff.-area product in

tower"

Factor CAir_d C_P 2543 SMALL BIG "Design rate of heat capacity flow of air "

Factor CLiq1_d C_P 2563 SMALL BIG "Heat capacity flow of liquid 1 at design rate"

Factor CLiq2_d C_P 2912 SMALL BIG "Heat capacity flow of liquid 2 at design rate"

Factor MLiq1MAir_d S_P 1.0 0.01 10 "Massflow relation of Liquid 1 and Air

(MLiq1_d/MAir_d). Normally in range 0.5 to 2"

Factor Min_MAir S_P 0.1 SMALL 1 "Min. airflow when calculating electricity use in

fan(s)"

Factor MAir_0 S_P 0.02 SMALL 1 "Min. airflow in fraction of design rate mass flow

of air in tower"

Factor MLiq1_0 S_P 0.02 SMALL 1 "Min. liquid 1 flow in fraction of design rate

liquid 1 flow"

Factor n_speed_d S_P 1 SMALL BIG "Reference fan speed (rpm) at design rate mass

flow of air in tower"

Factor n1 S_P 0.5 0.01 1 "Exponent in power law relation between EtaFan

and MAir"

Factor n2 S_P 0.65 0.01 1 "Exponent in power law relation between UA

and MAir"

Factor n3 S_P 0.43 0.01 1 "Exponent in power law relation between UA

and MLiq1"

Pressure pAtm S_P 101325 SMALL BIG "Ambient air pressure"

Pressure pw_d C_P 1024 SMALL BIG "Design rate partial vapour pressure of ambient

air"

Pressure dpFan_d S_P 300 SMALL BIG "Pressure difference over fan at design rate

mass flow of air in tower"

Pressure dpLiq1 S_P 10000 SMALL BIG "Pressure difference in circuit 1 (liquid 1)"

Pressure dpLiq2 S_P 15000 SMALL BIG "Pressure difference in circuit 2 (liquid 2

passing heat exchanger)"

Pressure dpSpray S_P 15000 SMALL BIG "Pressure difference in spray water circuit

(open circuit)"

HeatCapM cpLiq1 S_P 3685 500 5000 "Liquid 1 specific heat (3685= 40% propylene

glycol at 15°C)"

Density rhoLiq1 S_P 1039 500 2000 "Liquid 1 density (1039= 40% propylene

glycol at 15°C)"

HeatCapM cpLiq2 S_P 4187 500 5000 "Liquid 2 specific heat"

Density rhoLiq2 S_P 999 500 2000 "Liquid 2 density"

HeatCapM cpAir S_P 1005 900 1012 "Ambient air specific heat"

HeatCapM cpAir_ed C_P 3656 SMALL BIG "Design effective specific heat of air in tower "

Appendix A

173

Density rhoAir S_P 1.19 0.5 2 "Ambient air density"

MassFlow MAir_d C_P 0.7 SMALL BIG "Design rate massflow of air in tower"

MassFlow MLiq1_d C_P 0.7 SMALL BIG "Design rate massflow of liquid 1"

MassFlow MLiq2_d C_P 0.7 SMALL BIG "Design rate massflow of liquid 2"

Temp TLiq2Out_set S_P 16 -20 100 "Liquid 2 outlet set temperature"

Temp TAir_indb_d S_P 25 -20 40 "Design inlet dry bulb air temperature"

Temp TAir_inwb_d C_P 19.5 -20 40 "Design inlet wet bulb air temperature"

Temp TAir_out_d C_P 22.5 -20 40 "Given value of TAir_out_wb "

Temp TAir_outwb_d S_P 22.5 -20 40 "Help: Design outlet air wet bulb temp."

Temp TAir_in_d C_P 19.5 -20 40 "Help: Design inlet air wet/dry bulb temp. to

tower"

Temp TLiq1Cold_d C_P 23.5 -20 100 "Design cold side liquid 1 temp."

Temp TLiq1Warm_d C_P 26.5 -20 100 "Design warm side liquid 1 temp."

Temp T_ApproachP S_P 4 0.1 30 "Design difference between TAir_in_d and

TLiq1Cold_d in primary circuit"

Temp TRange_P S_P 3 0.1 30 "Design difference between TLiq1Warm_d

and TLiq1Cold_d in primary circuit"

Temp TApproachS S_P 1.5 0.1 30 "Design difference between TLiq1Cold_d and

TLiq2Out in secondary circuit"

Temp TRange_S S_P 3 0.1 30 "Design difference between TLiq2Out and

TLiq2In in secondary circuit"

Temp dT_Ln C_P 4 0.1 BIG "Log-mean temperature difference between

liquid 1 and air"

HumRatio HumAir_d C_P 0.012 SMALL BIG "Design humidity ratio ambient air

(kg water/kg dry air)"

Factor RelHum_d S_P 60 SMALL 100 "Design relative humidity of ambient air (%)"

Enthalpy hAirIn_d C_P 55446 -BIG BIG "Design enthalpy inlet (=ambient) air"

Enthalpy hAirOut_d C_P 66501 -BIG BIG "Design enthalpy outlet air"

Mass M_Tank S_P 100 SMALL BIG "Mass of liquid in tank and material in system"

HeatFlux_k QTower_d S_P 7.69 SMALL BIG "Design cooling capacity in tower (in kW)"

HeatFlux_k QCooling_d S_P 7.69 SMALL BIG "Design cooling load (in kW)"

Generic Mode_WetDry S_P 0 0 1 "Control Mode; 0=Wet Mode

1=Dry Mode"

Generic Mode_ClOp S_P 0 0 1 "Control Mode; 0=Closed circuit tower

1=Open circuit tower"

Generic Mode_Fan S_P 0 0 1 "Control Mode; 0=Fan with variable speed

control

1=Single speed fan"

Generic Mode_Tower S_P 0 0 1 "Control Mode; 0=Counterflow tower

1=Crossflow tower, both streams

unmixed"

Generic CTRL_T_TankIn S_P 0 0 1 ""

174

PARAMETER_PROCESSING

/**** Check for impossible comb. of Mode_WetDry (wet/dry) and Mode_ClOp (Closed/Open) *****/

IF Mode_WetDry > 0.5 AND Mode_ClOp > 0.5

THEN CALL NMF_ERROR ("Combination of dry Mode and open circuit Mode impossible!")

END_IF ;

/******************** Psychrometics at design stage *********************/

/* Ambient air partial vapour pressure, pw (Pa) and humidity ratio, HumAir_d (kg water/kg dry air) */

pw_d := RelHum_d * Satpres(TAir_indb_d) / 100 ;

HumAir_d := Humrat(pAtm, pw_d) ;

/* calculate wet bulb temperature of entering air stream at design stage */

TAir_inwb_d := Wetbulb(TAir_indb_d, HumAir_d) ;

/* Design inlet air wet bulb/dry bulb temperature from tower */

TAir_in_d := IF Mode_WetDry < 0.5

THEN TAir_inwb_d /* Cooling Tower is in wet mode */

ELSE TAir_indb_d /* Cooling Tower is in dry mode */

END_IF ;

/* enthalpies of entering and leaving air streams at design stage (Cooling Tower in wet mode) */

hAirIn_d := Enthal(TAir_indb_d, HumAir_d) ;

hAirOut_d := 9362.5 + 1786.1 * TAir_outwb_d + 11.35 * TAir_outwb_d**2 + 0.98855* TAir_outwb_d**3 ;

/* hAirOut = f (TAir_outwb_d) (ref [1]) is a good approximation as long as

humidity ratio in leaving air is fairly close to saturation curve. This is

normally the case when cooling tower is in wet mode. */

/************ Calculation of UA_d and MLiq1_d at design rate***************/

/* Equations based on model in ref. [1] */

/* The design effective specific heat of air i cooling tower */

cpAir_ed := IF Mode_WetDry < 0.5 /* Cooling Tower in wet mode */

THEN (hAirOut_d - hAirIn_d) / (TAir_outwb_d - TAir_inwb_d)

ELSE cpAir /* Cooling Tower in dry mode */

END_IF ;

/* Design outlet liquid 1 (cold side) temperature from tower in primary circuit */

TLiq1Cold_d := TAir_in_d + T_ApproachP ;

-

/* Design inlet liquid 1 (warm side) temperature to tower in primary circuit */

TLiq1Warm_d := TLiq1Cold_d + TRange_P ;

/* The design mass flow of liquid 1 and 2 */

MLiq1_d := QTower_d * 1000 / (cpLiq1 * (TLiq1Warm_d - TLiq1Cold_d)) ;

MLiq2_d := MLiq1_d * QCooling_d * cpLiq1 * TRange_P / (QTower_d * cpLiq2 * TRange_S) ;

/* Design volume flow of air to massflow of air */

Mair_d := MLiq1_d / MLiq1MAir_d ;

/* Design heat capacity flow of air and liquid */

Appendix A

175

CAir_d := cpAir_ed * Mair_d ;

CLiq1_d := cpLiq1 * MLiq1_d ;

CLiq2_d := cpLiq2 * MLiq2_d ;

/* Calculation of leaving air temperature. When Cooling Tower is in wet mode

TAir_out_d = TAir_outwb_d which is a supplied parameter. The leaving wet bulb

temperature has to be calculated iteratively elsewhere. A routine for

calculating TAir_outwb_d should be added */

TAir_out_d := IF Mode_WetDry < 0.5

/* Cooling Tower is in wet mode: */

THEN TAir_outwb_d

/* Cooling Tower is in dry mode: */

ELSE QTower_d + (cpAir * Mair_d * TAir_indb_d) / (cpAir * Mair_d)

END_IF ;

/* Log-mean temperature difference between liquid and air in primary circuit*/

dT_Ln := ((TLiq1Warm_d - TAir_out_d ) - (TLiq1Cold_d - TAir_in_d)) /

LOG((TLiq1Warm_d - TAir_out_d) / (TLiq1Cold_d - TAir_in_d)) ;

/* The design effective heat transfer coefficient-area product in primary circuit */

UA_ed := QTower_d * 1000 / dT_Ln ;

/* The design heat transfer coefficient-area product in primary circuit */

UA_d := UA_ed * cpAir / cpAir_ed ;

/******* Calculation of efficiency of intermediate heat exchanger at design rate ********/

EtaHex := IF CLiq2_d < CLiq1_d

THEN TRange_S / (TRange_S + TApproachS)

ELSE (CLiq2_d / CLiq1_d) * TRange_S / (TRange_S + TApproachS)

END_IF ;

/*******************************************************************************************/

END_MODEL

175

APPENDIX B

Pilot plant at Kvarnberget, Gothenburg

In this appendix the pilot plant at Kvarnberget, Gothenburg in Sweden is described Heat exchanger (between freeze protected primary circuit and secondary circuit)

A plate heat exchanger for free cooling, product Alfa Laval type CB76-100E or similar Design data: Heat transfer capacity 16,5 kW Warm side: Design volume flow, water, 1,3 l/s Design temperature in/out 25,6/22,6 °C Design pressure drop 10 kPa Cold side: Design volume flow, propylene glycol solution 40%, 1,45 l/s Design temperature in/out 21,6/24,6°C Design pressure drop 10 kPa Evaporative cooler

Air cooled heat exchanger with water spray function, product Asarums Industries model XP90-5 460 rpm-15, vertical direction of air. Design data:

Rated cooling capacity 17 kW Air flow rated 8,61 m3/s Ambient air condition 23°C/70% RH Liquid flow 1,58 l/s prop. glycol 40 % Liquid temperature return 24,6°C Liquid temperature supply 21,6°C Pressure drop liquid 54 kPa Dri-Batic flow 1,35 l/s 3 Fans diam. 890 mm/ 450 rpm Motor spec. 3/230 V 0,34kW 1,64A supplied with variable speed drive. In the secondary circuit there are chilled beams which have a design ∆Tm of 8°C and a design specific cooling capacity of 50 W/m2 floor area. The circulation the primary and secondary circuits is maintained by pumps equipped with variable frequency drives (VFD). The control system strives to maintain a constant supply temperature in the secondary circuit. The constant supply temperature is maintained through sequential control where, in order and at increasing cooling demand, the pump in the primary circuit increase its speed, the fans in the free cooler increase their speed and finally the spray water is turned on. The control system also includes a function which disconnects the free cooling system and connects the conventional comfort cooling system, through a set of shut-off valves, if the indoor

APPENDIX B

176

temperature exceeds 25°C. This function was applied through demand from both the building owner Vasakronan and the tenant ÅF-konsult. The part of the building at which the pilot plant is installed is erected during the 1970’s with corresponding building standard. The office space which the free cooling system is serving has a floor area of 450 m2 with an outer wall facing south east. The people density is nominally 14 m2/person and at normal operation the density is about 16 – 20 m2/person. Design power density for lighting is 15 W/m2. There are no external sun shading devices.

177

APPENDIX C

Monitoring system at Kvarnberget, Gothenburg

In this appendix the monitoring system at Kvarnberget, Gothenburg in Sweden is described

The monitoring of the pilot plant with this measurement system was carried out during the

May 1 to the August 31, 2007. Specifications of the sensors used are listed at the end of this

appendix. The following points where monitored:

Outdoor

Solar radiation: Measurement of total and global diffuse solar radiation with pyranometers.

Sample interval is 30 seconds and samples are processed to one hour average figures. The

solar radiation is measured on the roof of the office building at Kvarnberget

Ambient air temperature: One sensor, placed on the roof. Sample interval 6 minutes and

samples are processed to one hour average figures.

Ambient air relative humidity: One sensor, placed on the roof. Sample interval 6 minutes

and samples are processed to one hour average figures.

Indoor

Volume flow secondary circuit: Flow rate is measured with a volume flow meter at the

office floor which is monitored. Pulses from the meter is sampled in 15 minutes interval and

processed to one hour average figures of flow rate.

Liquid temperature secondary circuit: Two sensors placed in the supply and return liquid

coolant serving the monitored office floor. Sample interval is 15 minutes and samples are

processed to one hour average figures.

Indoor air temperature: Three sensors distributed to one in each “room” in the open office

space. The sensors are mounted on an inner wall at 1,1 m above floor level. At one of the

sensors there are two additional sensors which measure the air temperature at 10 cm below the

false ceiling and 10 cm above floor level respectively to record the temperature gradient in the

office space. Sample interval is 15 minutes and samples are processed to one hour average

figures.

Indoor air relative humidity: One sensor placed in one of the “rooms” in the open office

space at 1,1 m above floor level. Sample interval is 15 minutes and samples are processed to

one hour average figures.

Supply air temperature: One sensor placed in the supply duct. Sample interval is 15 minutes

and samples are processed to one hour average figures.

Electric energy: Electric power is recorded in the distribution box serving the office space.

Electric energy is integrated during 15 minutes and samples are processed to one hour average

figures.

178

Free cooler

Volume flow primary circuit: Flow rate is measured with a volume flow meter in the

primary circuit. Pulses from the meter is sampled in 15 minutes interval and processed to one

hour average figures of flow rate. The primary circuit has a freeze protected liquid (ethylene

glycol at 40% concentration).

Liquid temperature primary circuit: Two sensors placed in the supply and return of the

freeze protected liquid close to the plate heat exchanger. Sample interval is 15 minutes and

samples are processed to one hour average figures.

Liquid temperature secondary circuit: Two sensors placed in the supply and return liquid

coolant (water) serving the monitored office floor close to the plate heat exchanger. Sample

interval is 15 minutes and samples are processed to one hour average figures.

Electricity: Electric power to the fans in the free cooler and circulating pumps in primary and

secondary circuits is measured. Pulses from the meter is sampled in 15 minutes interval and

processed to one hour average figures of electric energy.

Volume flow spray water: Flow rate is measured with a volume flow meter in the supply

pipe of fresh water. Pulses from the meter is sampled in 15 minutes interval and processed to

one hour average figures of flow rate.

List of sensors, office floor Sensor Channel Description Sensor

Analogue sensor

REFRES A00 Reference resistance 100 Ohm

GT01 A01 Room temperature, south west Pt100 #1

GT02 A02 Room temperature middle, ceiling Pt100 #4

GT03 A03 Room temperature middle, intermediate Pt100 #5

GT04 A04 Room temperature middle, floor Pt100 #6

GT05 A05 Room temperature north east Pt100 #2

GT06 A06 Temperature coolant, supply Pt100 #309046

GT07 A07 Temperature coolant, return Pt100 #309050

GT08 A08 Temperature supply air Pt100 kanal#2

GF09 A09 Rel. humidity, office Vaisala A

Digital sensor

EL00 D00 Electric power Ackr. electric meter

FL01 D01 Volume flow, coolant Krohne

Calculated data

QE00 E00 Cooling power, secondary circuit GT06, GT07, FL01

Appendix C

179

List of sensors, outdoor

Sensor Channel Description Sensor

Analogue sensor

REFRES A00 Reference resistance 100 Ohm

GTU01 A01 Outdoor temperature Pt100

GFU02 A02 Outdoor relative humidity Vaisala B

GST03 A03 Solar radiation, total Kipp&Zonen #850918

GSD04 A04 Solar radiation, diffuse Kipp&Zonen #850870

Digital sensor

None

Calculated data None

List of sensors, evaporative free cooler

Sensor Channel Description Sensor

Analogue sensor

REFRES A00 Ref. resistance 100 Ohm

GTK01 A01 Temp. primary supply Pt100 #309045

GTK02 A02 Temp. primary return Pt100 #309048

GTK03 A03 Temp. secondary supply Pt100 #309052

GTK04 A04 Temp. secondary return Pt100 #309053

Digital sensor

FLK01 D00 Volume flow, coolant Krohne

FLK02 D01 Volume flow, spray water Krohne

ELK01 D02 Electricity, fans

ELK02 D03 Electricity, pump prim. circ.

ELK03 D04 Electricity, pump sec. circ.

XK01 D05 Operational mode, cooling system

Calculated data

QEK00 E00 Cooling power, primary circuit A01, A02, D00

180

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