Performance Analysis of Hybrid Liquid Desiccant …Performance Analysis of Hybrid Liquid Desiccant...

Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System

I

QUEENSLAND UNIVERSITY OF TECHNOLOGY

Performance Analysis of Hybrid Liquid Desiccant

Solar Cooling System

By

Zhipeng ZHOU(Joe ZOE)

A THESIS SUBMITTED FOR THE DEGREE OF

MASTER OF ENGINEERING (RESEARCH)

School of Engineering System

Queensland University of Technology

2009


II

AUTHORSHIP

The work contained in this thesis has not been previously

submitted for a degree or diploma at any other higher education

institution. To the best of my knowledge and belief, the thesis

contains no material previously published or written by another

person except where due reference is made.

Signature of Author

Date


III

© Copyright 2009

By



IV

ACKNOWLEDGMENTS

I wish express my sincere gratitude to Prof. John Bell, my principal supervisor, who gave

me constant encouragement, guidance and friendship throughout this process.

Completing this thesis would not have been possible without the help and encouragement

from so many people. My sincerest gratitude goes to all of those who have been directly

and indirectly involved. I am also grateful for the helpful advice from Dr. Kame

Khouzam, who served as the associate supervisor. I feel honoured to have worked close

to so many brilliant graduate students, Nick Ward, Dong Choon (Daniel) Sin and Travis

Frew. All their help and support will be remembered a lifetime.

Finally, it is not possible to describe the thankfulness I feel towards, my parents and

brother. Without their abundant love and support, this work would not have been

completed.



V

ABSTRACT

This thesis investigates the coefficient of performance (COP) of a hybrid liquid desiccant

solar cooling system. This hybrid cooling system includes three sections: 1) conventional

air-conditioning section; 2) liquid desiccant dehumidification section and 3) air mixture

section. The air handling unit (AHU) with mixture variable air volume design is included

in the hybrid cooling system to control humidity. In the combined system, the air is first

dehumidified in the dehumidifier and then mixed with ambient air by AHU before

entering the evaporator. Experiments using lithium chloride as the liquid desiccant have

been carried out for the performance evaluation of the dehumidifier and regenerator.

Based on the air mixture (AHU) design, the electrical coefficient of performance (ECOP),

thermal coefficient of performance (TCOP) and whole system coefficient of performance

(COPsys) models used in the hybrid liquid desiccant solar cooing system were developed

to evaluate this system performance. These mathematical models can be used to describe

the coefficient of performance trend under different ambient conditions, while also

providing a convenient comparison with conventional air conditioning systems. These

models provide good explanations about the relationship between the performance

predictions of models and ambient air parameters. The simulation results have revealed

the coefficient of performance in hybrid liquid desiccant solar cooling systems

substantially depends on ambient air and dehumidifier parameters. Also, the liquid

desiccant experiments prove that the latent component of the total cooling load

requirements can be easily fulfilled by using the liquid desiccant dehumidifier. While

cooling requirements can be met, the liquid desiccant system is however still subject to

the hysteresis problems.


VI

TABLE OF CONTENTS

1 INTRODUCTION ................................................................................................. 1

2 The RESEARCH PROBLEM................................................................................ 6

2.1 Research Background .................................................................................... 6

2.2 Research Problem ........................................................................................ 11

2.3 Research Objectives and Scope ................................................................... 11

2.4 Contributions of This Research ................................................................... 13

3 LITERATURE REVIEW .................................................................................... 15

3.1 Dehumidification Cooling Concept ............................................................. 15

3.2 Solid Desiccant Cooling System and Solid Desiccants ............................... 16

3.3 Liquid Desiccant Cooling System and Liquid Desiccants ........................... 17

3.4 Hybrid System ............................................................................................. 21

3.5 Mathematical Models of the Dehumidifier/Regenerator System ................ 23

3.6 Different Control Strategies for Indoor Air Humidity ................................. 25

3.6.1 Bypass air (BA) control ........................................................................... 26

3.6.2 Variable Air Volume (VAV) control ....................................................... 27

3.7 Concluding Remarks .................................................................................... 27

4 PERFORMANCE STUDY OF HYBRID LIQUID DESICCANT SOLAR

COOLING SYSTEM ................................................................................................... 29

4.1 Introduction .................................................................................................. 29

4.2 Model Formulation ...................................................................................... 31

4.3 System Description ...................................................................................... 34

4.4 Standard Assumptions ................................................................................. 38

4.5 Air Mixture Modelling and Air Enthalpy Calculation ................................. 39

4.5.1 Air Mixture Modelling ............................................................................. 39

4.5.2 Air Mixture Parameters............................................................................ 40

4.5.3 Air Mixture Enthalpy Calculation ........................................................... 42

4.6 Conventional Air Conditioning Performance and Hybrid Cooling System

Electrical Coefficient of Performance ..................................................................... 44

4.6.1 COPcon and ECOP Definition .................................................................. 44

4.6.2 COPcon, ECOP Results and Discussion .................................................... 45

4.7 Thermal Coefficient of Performance ........................................................... 49

4.7.1 TCOP Definition ...................................................................................... 49

4.7.2 TCOP Results and Discussion ................................................................. 53

4.7.2.1 TQ Thermal Energy Remains Constant ........................................... 53

4.7.2.2 Relationship between TCOP and Point1, 2 states ............................ 54

4.8 System Coefficient of Performance ............................................................. 56

4.8.1 COPsys Definition ................................................................................... 56

4.8.2 COPsys and ECOP, TCOP Relationship ................................................. 57

4.8.3 COPsys and Cooling Load Relationship ................................................. 61

4.8.4 COPsys and Solar Energy Relationship ................................................... 62


5 REGENERATION AND DEHUMIDIFICATION TEST, RESULTS AND

DISCUSSION .............................................................................................................. 66

5.1 Introduction .................................................................................................. 66

5.2 Experimental Configuration......................................................................... 66

5.3 Experimental Components ........................................................................... 69

5.3.1 Selection of Desiccant.............................................................................. 69


VII

5.3.2 Regenerator/Dehumidifier Description .................................................... 70

5.3.3 Test Equipment Description .................................................................... 71

5.4 Regeneration Test ........................................................................................ 72

5.4.1 Introduction .............................................................................................. 72

5.4.2 Test Parameters ........................................................................................ 73

5.4.3 Regeneration Results and Discussion ...................................................... 74

5.4.3.1 Liquid Desiccant Concentration ...................................................... 74

5.4.3.2 Regeneration With the Hot Water Pump ......................................... 76

5.4.3.3 Regeneration Without the Hot Water Pump .................................... 78

5.4.3.4 Humidity Analysis in the Regeneration Test with the Hot Water Pump

79

5.4.3.5 Regeneration Test Results Discussion ............................................. 80

5.5 Dehumidification Test ................................................................................. 82

5.5.1 Introduction .............................................................................................. 82

5.5.2 Dehumidification Results and Discussions .............................................. 83

5.5.2.1 Liquid Desiccant Concentration ...................................................... 83

5.5.2.2 Humidity Results Analysis .............................................................. 84


6 AIR MIXTURE AND AMBIENT AIR DATA ANALYSIS .............................. 89

6.1 Introduction .................................................................................................. 89

6.2 Air Mixture Rate Calculation Analysis........................................................ 89

6.3 Weather Data Analysis ................................................................................ 92

6.3.1 Summer Weather Data Analysis .............................................................. 92

6.3.2 Whole Year Weather Data Analysis ........................................................ 95

6.4 Ambient Air Latent Load and Sensible Load Analysis ............................... 98

6.5 Concluding Remarks .................................................................................. 101

7 CONSLUSIONS AND RECOMMENDATIONS ............................................. 103

7.1 Conclusions ................................................................................................ 103

7.2 Recommendations for Future Research ..................................................... 105

8 REFERENCES .................................................................................................. 107

9 APPENDICES ................................................................................................... 110


VIII

LIST OF FIGURES

Figure 1-1 Common Vapour-compression Refrigeration System ........................... 1

Figure 1-2 Air-conditioning and Refrigerator on Psychrometric Chart .................. 2 Figure 2-1 Psychrometric Chart of Ideal Air Disposed Process .............................. 6 Figure 2-2 Psychrometric Chart of Conventional Air-conditioning (Including

Dehumidification Process) ............................................................................................. 7 Figure 2-3 Psychrometric Chart of Conventional Hybrid Cooling System............. 7

Figure 2-4 Psychrometric Chart of Hybrid Cooling and Mixture System............ 10 Figure 2-5 A Conceptual Hybrid Vapour-Liquid Desiccant Cooling System ....... 10 Figure 3-1 A Conceptual Solid Desiccant Cooling System................................... 16 Figure 3-2 A Conceptual Liquid Desiccant Cooling System ................................ 17

Figure 3-3 A Conceptual Hybrid Vapour-Liquid Desiccant Cooling System ....... 22 Figure 3-4 Psychrometric Chart Depicting Hybrid Cooling Process .................... 22 Figure 3-5 Schematic of Three Different Humidity Control Strategies ................ 26 Figure 4-1 COP Calculation Flow Chart ............................................................... 33

Figure 4-2 Different Experimental Test Points in the Hybrid Cooling System .... 36 Figure 4-3 Hybrid AC Air mixture and Cooling Situations in Psychrometric Chart36 Figure 4-4 Different Experimental Test Points in the Conventional AC System .. 37 Figure 4-5 Conventional AC Air Cooling Situations in Psychrometric Chart ...... 37

Figure 4-6 Conventional Air Conditioning Cooling Load and COP Relationship 46 Figure 4-7 Hybrid Air Conditioning Cooling Load and ECOP Relationship ....... 48

Figure 4-8 TCOP and Point 1, Point 2 Temperature Relationship ........................ 54 Figure 4-9 Hybrid Cooling System Cooling Load and TCOP Relationship ......... 55 Figure 4-10 TCOP and Point 1, Point 2 Absolute Humidity Relationship .............. 55

Figure 4-11 Hybrid Cooling System ECOP, TCOP and COPsys Relationship ...... 60 Figure 4-12 Hybrid Air Conditioning Cooling Load and COPsys Relationship ..... 62

Figure 4-13 Hybrid Cooling system COPsys with Solar Energy and without Solar

Energy 63

Figure 5-1 Schematic Diagram of the Regeneration and Dehumidification Systems . 67 Figure 5-2 Air and Desiccant Parameters in the Hybrid Cooling System ............. 68 Figure 5-3 Dehumidifying, Heating and Cooling Process on the Psychrometric Chart68 Figure 5-4 Solubility boundary of aqueous solutions of lithium chloride (Conde, 2004)

...................................................................................................................................... 70 Figure 5-5 Regeneration and Dehumidification Test System................................ 71 Figure 5-6 Schematic Diagram of Regeneration Test ........................................... 73 Figure 5-7 Schematic Diagram of Liquid Desiccant Solution Concentration in the

Regeneration Test ........................................................................................................ 75 Figure 5-8 Regeneration Test with Hot Water Pump ............................................ 76 Figure 5-9 Regeneration Experimental Results Trend Analysis in Psychrometic chart

77 Figure 5-10 Regeneration Test without Hot Water Pump ....................................... 78 Figure 5-11 Humidity Situation in Regeneration Test with Hot Water Pump ........ 79 Figure 5-12 Humidity Situation in Regeneration Test ........................................... 81 Figure 5-13 Schematic Diagram of Dehumidification Test .................................... 82

Figure 5-14 Schematic Diagram of Liquid Desiccant Solution Concentration in

Dehumidification Experiment ...................................................................................... 84 Figure 5-15 Humidity Results Analysis in Dehumidification Experiment ............. 85 Figure 5-16 Results Trend Analysis in Dehumidification Experiment ................... 86


IX

Figure 5-17 Dehumidification Experimental Results Trend Analysis in Psychrometic

chart.............................................................................................................................. 87 Figure 6-1 Mixture Air Process in Psychrometric Chart ....................................... 90 Figure 6-2 Schematic Diagram of Air Mixture Points .......................................... 90 Figure 6-3 Air Mixture Mass Rate and Air Humidity Analysis ............................ 92

Figure 6-4 Brisbane Temperatures and Humidity Analysis .................................. 95 Figure 6-5 Brisbane Whole Year Weather Statistics ............................................. 98


X

LIST OF TABLES

Table 3-1 Theoretical models for packed bed absorbers (Öberg, 1998) .................. 24

Table 4-1 DAIKIN Air Conditioning Specification(RMK140J Series Composition)46 Table 4-2 Input data on the performance of conventional air conditioning and hybrid

liquid desiccant air conditioning* ................................................................................ 47 Table 4-3 ECOP, TCOP and COPsys on the conventional air conditioning and hybrid

liquid desiccant air conditioning* ................................................................................ 58

Table 4-4 Hybrid System Parameters Value/Range ................................................. 61 Table 5-1 Test Equipments List ............................................................................... 72


XI

List of Abbreviations and nomenclature

ASHRAE

CaCl2

CPU

CFC

CIBSE

COP

CELD

DBTs

ECOP

HLDSAC

LiBr

LiCl

MSDS

TCOP

TEG

VAV

American Society of Heating, Refrigerating and Air

Conditioning Engineers

Calcium Chloride

Central Processing Unit

Chloro Fluoro Carbon

Chartered Institution of Building Services Engineers

Coefficient of Performance

Cost-Effective Liquid Desiccant

Dry Bulb Temperatures

Electrical Coefficient of Performance

Hybrid Liquid Desiccant Solar Air Conditioner

Lithium Bromide

Lithium Chloride

Material Safety Data Sheets

Thermal Coefficient of Performance

Triethylene Glycol

Variable Air Volume

c cooling

con conventional air conditioning

e the energy constant (kJ)

liquid desiccant solution concentration (%)

h enthalpy (kJ/kg)

L latent load (kJ)

m mass flux per unit time (kg/s)

air density of air (kg/m3)

TQ thermal energy supplying in dehumidification (kJ)

ratio of latent load to total load


XII

RH relative humidity (%)

airS average air flow rate (m3/s)

t temperature (°C)

T temperature (°C)

0T condenser temperature (°C)

RT evaporator temperature (°C)

humidity ratio (kg/kg, kg of water vapour per kg of dry air )

W compressor energy (kJ)

cW conventional compressor electricity cost


1

1 INTRODUCTION

With dramatic climatic changes in recent years, demand for research concerning

cooling systems with solar dehumidification is becoming significant. The application

of solar energy in cooling systems can reduce energy demand and decrease

Chlorofluorocarbon (CFC) usage (which depletes the ozone layer). Most traditional

air-conditioning / cooling systems are based on a closed mechanical system where a

fixed amount of refrigerant is continuously cycled through evaporation and

condensation processes (Prasitpianchai, 1999). These common systems are based on

the vapour-compression principle, which has system evaluation based upon the

coefficient of performance (COP) parameter. While there are other methodologies

that can be followed to design a cooling system (such as with the absorption air-

conditioner), the vapour-compression methods currently have the highest COP, and

lowest purchase cost. The fundamental operation of the vapour-compression system

is displayed in Figure 1-1, which has four components being the compressor,

condenser, expansion valve, and evaporator.

Figure 1-1 Common Vapour-compression Refrigeration System

A review concerning the basic operation of the common vapour-compression

refrigeration system shown in Figure 1-1 can be given by considering the entry of the

circulating refrigerant into the compressor. In the compressor the refrigerant is

compressed to a higher pressure and changes from a saturated vapour state, to a

superheated vapour thermodynamic state. The superheated vapour has partially

travelled through the condenser and superheat has been removed by being cooled


2

with water or air. During the remaining time in the condenser, the hot vapour

changes from a gas state, to a liquid state at constant pressure. The saturated liquid

then passes through the expansion value and experiences a sharp decline in pressure.

The result is that part of the liquid immediately transforms into vapour. This state

transformation occurs at constant enthalpy and is often referred to as ―adiabatic

flash‖, meaning it is isenthalpic. Adiabatic flashing lowers the temperature of the

liquid and vapour refrigerant so that it is colder than the temperature in the area

required to be cooled. By routing the cold refrigerant mixture through coils or tubes

in the evaporator, a fan circulates warm air that totally vaporises the refrigerant. At

the same time, the circulated air is cooled, which lowers the temperature of the area

required to be cooled. The resulting saturated vapour then returns to the compressor

inlet and completes the thermodynamic cycle. The resulting performance of air-

conditioning and refrigeration can be analysed with the Psychrometric chart in Figure

1-2. Discussion of the Psychrometric chart occurs throughout this thesis, with

particular emphasis in Chapters 2, 3 and 4.

Figure 1-2 Air-conditioning and Refrigerator on Psychrometric Chart

With the objective of this research concerned with the efficient electrical energy

design of cooling systems, relevant environment factors must be analysed. The

fundamental design of any cooling systems is based on the amount of heat energy

required to be removed from an indoor environment, with equipment that will

maintain the specified temperature when the worst case outdoor temperature is being


3

experienced. However, most industry situations accept the design outdoor

temperature, which is included most outdoor temperature situations. This is normally

called ambient air design temperature. The amount of heat to be removed by the

cooling system is referred to as the cooling load, which is divided into the following

load types:

Latent cooling load

Sensible cooling load

The latent cooling load refers to the wet bulb temperature of the indoor area, while

sensible cooling load refers to the dry bulb temperature. Factors that affect the latent

cooling load involve moisture that is introduced into the area through people,

equipment, or air that has infiltrated into the indoor area through cracks in the

surrounding walls, ceiling, etc. The sensible cooling load is influenced by many

factors such as sunlight, glass windows, doors, lights or roofs. The ability to induce

and sustain a state of dryness, otherwise known as dehumidification, provides system

control over the latent load. Hygroscopic substances that absorb water provide this

function and are known as desiccants. Traditional air conditioning systems usually

simultaneously cool and dehumidify in an energy intensive process (Gaffar, 2002).

Separating latent and sensible cooling in an air-conditioning system will offer

significant potential for energy savings and improve the vapour compression

coefficient of performance.

In the search for improvement of cooling system efficiency, this research has chosen

to focus on the design of a hybrid cooling system. A hybrid cooling system is

composed of a conventional air-conditioning cooling section (already reviewed), and

a desiccant dehumidification section. The purpose of the hybrid design is to dispose

of latent and sensible cooling load separately. In dealing with the dehumidification

section, there are two types of desiccant that can be used. These desiccants are

referred to as either a liquid desiccant, or solid desiccant. There are many advantages

in using liquid desiccants instead of solid desiccants. (Ertas, Anderson and Kiris,

1992). Such advantages include maintaining adequate ventilation within the enclosed

area, while also maintaining comfortable and healthy humidity. In the proposed

hybrid cooling system research, a liquid desiccant was chosen to be used in the


4

dehumidification section. While there are advantages in using a liquid desiccant,

there are however various problems associated with the hybrid cooling system. These

problems include control difficulties within the liquid desiccant spraying section, and

the COP of the whole hybrid system under different environmental conditions (Ertas,

Anderson and Kiris, 1992).

In this research program, the Hybrid Liquid Desiccant Solar Air Conditioner

(HLDSAC) is studied theoretically and analysed experimentally. With the theoretical

review being provided in Chapter 3, it will be seen that the HLDSAC is a

combination of the conventional air-conditioning system, with an Air Handling Unit

(AHU). The AHU is a device used to condition and circulate air as part of a heating,

ventilation, and air-conditioning (HVAC) systems. Experimental results will be

discussed in Chapter 4 and concern the liquid desiccant section. An effective air

mixture design will also be proposed to solve the humidity control problems

commonly associated with hybrid cooling systems. This design will allow some

ambient air into the system through AHU terminals to mix with the disposed dry air

to satisfy indoor air requirements. This thesis describes and compares system

performance (COP and energy cost) under different outdoor and indoor air situations.

The study also determines the optimal conditions for system design. The features of

solid desiccants, liquid desiccants and the mixture of different desiccants are

described in the literature review section of Chapter 3. A mathematical model is also

developed to describe the system performance for dehumidification and air mixture

processes between the ambient air and the air after dehumidifier. Energy analysis of

both the traditional air-conditioning system and the HLDSAC system will be

performed in Chapter 4.

It should be noted that a significant factor the affects system performance is the ratio

of the ambient air latent load, to the sensible load. The liquid desiccant

dehumidification section can be applied under different climatic conditions.

Therefore, this research also incorporates the design of an experiment to assess the

actual weather data and evaluates the COP of the system under the different climatic

conditions. The operation of the entire systems, in combination with the collection of

weather data is simulated out under the summer climatic conditions of Brisbane,


5

Australia. All weather data used in the proposed research was obtained from the

Australian Bureau of Meteorology (BoM).

This thesis is divided into seven chapters. Chapter 1 as previously discussed is an

introduction to cooling systems and provides an overview of the research program.

Chapter 2 explores the research problem and introduces relevant AHU design to

solve highlighted problems. Chapter 3 provides a literature review of different

cooling system configurations, with detailed discussion of desiccant properties and

system performance. This chapter reviews relevant work that establishes the

background for this research. Chapter 4 describes the performance analysis model,

where a detailed description of the model is provided. Various methods are used to

determine the COP of the proposed hybrid cooling system, including analysis of

system cooling and dehumidification. Chapter 5 presents the experimental study for

the liquid desiccant dehumidification system. Chapter 6 includes analysis of the air

mixing system and discusses the relationship between ambient air data and system

performance. Ambient air latent load and sensible load analysis are also presented in

this chapter. Finally, Chapter 7 contains conclusions and recommendations. The last

section provides references and appendices. The appendices contain weather data and

different test point parameters.


6

2 The RESEARCH PROBLEM

2.1 Research Background

The significant benefits offered by hybrid cooling systems are lower electrical power

consumption by the compressor, and a higher coefficient of performance (COP) of

the system. COP increases in hybrid systems because they remove latent heat from

liquid desiccant. The other reason is because a hybrid system is not required to

reheat the coil to maintain the space temperature as in the conventional air

conditioning. The ideal state points of the operating air situation are shown in Figure

2-1, which is a psychrometric chart showing a graph of the physical properties of

moist air under constant environment pressure.

D

A

A'

A''

Aim Point

Figure 2-1 Psychrometric Chart of Ideal Air Disposed Process


7

D

A

A'

A''

Reheating

D'

B

B'

B''

Aim Point

Figure 2-2 Psychrometric Chart of Conventional Air-conditioning (Including

Dehumidification Process)

Points A, A‘ and A‘‘ represent different ambient air situations with different

temperature and humidity. Point D shows the aim point of the air situation that the

air-conditioning system is expected to achieve. Figure 2-2 illustrates how ambient

air points A, A‘ and A‘‘ can arrive at point D directly on the psychrometric chart

with the conventional system. Most of the conventional air conditioning (Figure 2-2 )

and hybrid cooling (Figure 2-3) systems can only change ambient air points A, A‘

and A‘‘ to point D via different paths, indicating greater operational cost.

A''

A

D

A'

Over

Dehumidification C'

C

C''Aim Point

Figure 2-3 Psychrometric Chart of Conventional Hybrid Cooling System


8

When the moisture level in the air is high, such as with air at the seaside in summer

time, the conventional air-conditioning must work as a mechanical dehumidifier to

remove moisture by passing air over a surface the has been cooled below the air‘s

dew point. This cold surface may be the exterior of a chilled-water coil or a direct-

expansion refrigerant coil. To prevent overcooling the space (and avoid the need to

add heat energy from another source), a mechanical dehumidifier also usually has

means to reheat the air, normally using recovered and recycled energy (e.g.,

recovering heat from hot refrigerant vapour in the refrigeration

circuit).(ASHRAE,2008) The conventional air-conditioning system with a typical

dehumidification process is shown in Figure 2-2 , which cools air to the dew point

while achieving dehumidification (A→B→D‘→D). The system therefore need

reheat D‘ to D to achieve aim point. The conventional hybrid cooling system does

not need reheat and can dehumidify directly. It however may have some over

dehumidification problems as shown in Figure 2-3 (A→C→D). The literature review

will provide further discussion of this hybrid system problem. Some assumptions of

constant outdoor air temperature and humidity are still needed during simulations

(Dai et al., 2001) in conventional hybrid air conditioning, because it is difficult to

change the dehumidification ability by liquid desiccant dehumidifier design (Gaffar,

2002, Lazzarin, Gasparella and Longo, 1999). Lazzarin, Gasparella and Longo (1999)

present some experimental data about LiCl water solution in the dehumidification

experiment. This data indicates that the dehumidification ability does not increase

significantly even for higher desiccant flow rate ratios. Liquid desiccant running

performance in Nelson Fumo‘s experiments also presents similar results (Fumo and

Goswami, 2002). From the literature review, it can be seen that the traditional liquid

desiccant dehumidifier is not flexible enough to meet different air requirements.

Traditional air-conditioning employs humidifiers to overcome the humidity control

situation. According to the mechanism used for evaporation of water vapour from

water humidifiers, this can be classified as steam with heating element humidifiers,

atomizing humidifiers, or wetted element humidifiers (Wang, 2000). These methods

all need to add extra water or water vapour into the system. Addition of water or

water vapour into the system is a complex heat and mass transfer process (Tashtoush,


9

Molhim and Al-Rousan, 2005). It is difficult to accurately calculate the properties of

the resulting mixture. System equipment requirements also become complex.

The psychrometric chart of a hybrid air conditioning system incorporating the air

mixture system is presented in Figure 2-4, while the system function blocks are

shown in Figure 2-5. This design can solve the problem of adjusting humidity in the

liquid desiccant system and provide more accurate humidity control of the output air.

The whole system is composed of three main sections, where the first section is a

conventional air conditioning section. The second section is a liquid desiccant

dehumidifier and regenerator section. The dehumidification section is used to remove

moisture (latent heat) from the process air, and the conventional air conditioning

section is used to cool air (sensible heat). The third section AHU involves the

Variable Air Volume (VAV), where relative humidity sensor controllers are used

between the first and the second sections to control the mixing of ambient air into the

system. The AHU design aims to balance the air humidity after dehumidifier to

satisfy different indoor requirements. It uses VAV and several sensors to detect the

air after dehumidifier humidity/flow rate and control the fan to let ambient air into

the cooling system to mix with the dry air. The air mixture flow is continually

modified by conventional air conditioning to remove sensible heat before supplying

the indoor environment. This process guarantees that supply air can satisfy both the

air humidity and temperature according to indoor requirement.


10

C

A

D

E

Aim Point

Mixing Air Process

`

Figure 2-4 Psychrometric Chart of Hybrid Cooling and Mixture System

Dehumidification Unit

Conventional Air Conditioning

Unit

Expansion

ValveCompressor

Air Handling Unit Terminal

Controller

Condenser

Evaporator

Dehumidifier

Heat

Exchanger

Regenerator

Solar Heat

Resource

Ambient

Air

Ambient

Air

Ambient

Air

Indoor

Space

Exhaust

Co

nc

en

tra

te

De

sic

ca

nt

Dilute Desiccant

Ambient

Air Exhaust

Disposed

Air

1 2

3 4

5 6

8 7

9 10

11

12

13

1416

17

15

18

1

Figure 2-5 A Conceptual Hybrid Vapour-Liquid Desiccant Cooling System


11

2.2 Research Problem

The principal research problem addressed in this thesis is to conduct the system

performance analysis (COP and energy cost) of hybrid liquid desiccant air-

conditioning system with an AHU controller.

Analysis of system performance with the AHU controller is provided in the

following three sections. 1) The conventional air-conditioning system component, as

measured with the ECOP (Electrical Coefficient of Performance). 2) The liquid

desiccant component, as measured with the TCOP (Thermal Coefficient of

Performance). 3) The complete system as analysed with the System Coefficient of

Performance COPsys (based on mixture air design).

2.3 Research Objectives and Scope

The following research objectives were identified to provide the necessary

information to address the research problem:

To develop a total system performance analysis model for the mixed air and

hybrid cooling process. This model should be able to forecast the whole

system performance and energy analysis.

To conduct performance analysis of the conventional air-conditioning system,

and setup the ECOP model.

To analyse the liquid desiccant section with the TCOP model.

To develop an ambient air mixture system (AHU) to provide humidity control,

thereby addressing problems associated with existing hybrid air conditioning

technology.


12

To develop a mathematical model of the air mixture system to determine

optimal operational parameters to achieve the highest efficiency for specific

ambient conditions.

To investigate the optimal range concerning parameters of the liquid

desiccant dehumidification experimentally, then compare this with

calculation results of the theoretical model.

The objective of this performance analysis involves separate analysis of the

conventional cooling section, dehumidification section, and air mixing subsystem.

The performance analysis can describe the whole system operation, including ECOP,

TCOP and COPsys. Dehumidification and air mixture optimization analysis is also

included in this section. The complete mathematical model solution based on the air

mixture design will give the optimum ambient air and the air after dehumidifier flow

rate ratio relationship. These results can be used for AHU programming. Furthermore,

every factor that influences system performance can be easily analysed from the

mathematical expression. Hybrid liquid desiccant cooling systems are able to control

humidity and temperature independently. As a result, the evaporator temperature in

the conventional air conditioning section can be increased by up to 15 degrees, and

hence the COP of the cooling will be significantly improved (Ma et al., 2006).

System performance analysis is based on mixing with ambient air. As a result, the

research objective also includes analysing the air mixing design. This design uses

the AHU system to control air flow rate. Variable air volume air-conditioning

systems, which are deemed more economical than other alternative systems have

been adopted in buildings to maintain the varying cooling and heating demands (Qin

and Wang, 2005). The conventional VAV system is used to adjust the supply of air

volume to the indoor requirement to achieve economical cooling. The mixed air

component in a hybrid system will use VAV technology to control the ambient air

volume to mix the air after dehumidifier, and satisfy indoor air requirements. The

new mixed air AHU design includes three components: 1) CPU (Central Processing

Unit) micro-controller, 2) control air flow ratio section, and 3) sensor section. There

are several sensors in the system to monitor the air after dehumidifier and ambient air

data. All the air after dehumidifier data, such as humidity and temperature will be


13

recorded with sensors and transferred into the CPU. The CPU is programmed

according to indoor air requirements to calculate relevant data. The micro-controller

will then give orders to the controlling section. The controlling section uses a fan to

adjust inlet air into the mixture section. A new mixing air analysis based on VAV

hardware and hybrid design is developed to detect VAV terminal flow air mixture.

The other objective of this research is to determine the operational parameters of a

packed bed liquid desiccant dehumidifier and to develop an analytical solution for

the air mixture design used in the hybrid liquid desiccant cooling system. The air

mixture model is used to describe the relationship between the volume of the air after

dehumidifier states and the volume of new air states. Ambient air data analysis is

also included here to evaluate the air mixture process situation and efficiency. The

ambient air data includes latent load and sensible load, and ambient air analysis will

combine both dehumidification model results and air mixture section. Ambient air

data is based on meteorological data for the summer season in Brisbane. This

analysis will give optimal parameters for ambient air and liquid desiccant

dehumidification parameters. Ambient air latent load and sensible load rate analysis

is based on mixture gas theories used in air conditioning design.

2.4 Contributions of This Research

The principle objective of this research is to investigate the performance of a solar

hybrid liquid desiccant cooling system. In addition, this research will explore ways to

improve humidity control by mixing ambient air with the process air to solve the

hysteresis and ―over dehumidification‖ problems associated with current hybrid

cooling systems. This research includes experimental analysis concerning the

performance of the current hybrid cooling system designs. An inclusion of

theoretical analysis involving the performance of the proposed hybrid cooling system

embodying an ambient air mixture component for improved humidity control is also

provided. The important contributions of this research program are:

An experimental investigation of a liquid desiccant dehumidification and

regeneration unit that verifies the hysteresis and ―over dehumidification‖


14

problems associated with current hybrid cooling system technology. This

problem leads to poor humidity control. This is shown to be due to the

complex fluid flow and heat and mass transfer processes involved in the

liquid desiccant absorption cycle.

A solution to the hysteresis and ―over dehumidification‖ problems is

proposed, which involves incorporating an AHU ambient air mixture system

between the liquid desiccant dehumidifier and conventional air conditioning

sections of the hybrid cooling system. In many situations, mixing ambient air

with the over dehumidified process air enables the hybrid system to more

rapidly and more accurately obtain the desired humidity.

Theoretical modelling of the ambient air mixing process in the AHU system

was used to determine the ambient air mass flow rate required to be mixed

with the process air in order to obtain the desired humidity under various

conditions. This theoretical work will contribute to the design of control

systems for hybrid liquid desiccant cooling systems embodying an ambient

air mixture component.

Theoretical analysis concerning the performance of the hybrid system

(incorporating the proposed VAV ambient air mixing component) is used to

determine the expected performance of the system under different ambient

conditions. This work enables assessment involving the performance of the

proposed design compared to other existing or future cooling systems. The

performance parameters ECOP, TCOP and COPsys are evaluated.

The range of ambient environmental conditions over which the proposed hybrid

cooling system will operate is identified. A drawback concerning the use of ambient

air to improve humidity control involves the system being unable to increase the

humidity of the process air, if the ambient humidity is lower than the desired level.

However, it is shown that this problem will only affect the proposed system over a

period of two days through-out the year in Brisbane, Australia (based on observed

environmental data from 2005).


15

3 LITERATURE REVIEW

3.1 Dehumidification Cooling Concept

Dehumidification cooling systems are essentially open absorption cycles, utilising

water as the refrigerant in direct contact with air. The desiccant (sorbent) can be

either solid or liquid and is used to facilitate the exchange of sensible and latent heat

of the conditioned air stream (Grossman, 2001, 53-62). Using desiccant as the

sorbent in the open cycles has several advantages relative to the closed absorption

cycle. These advantages include: 1) operation at ambient pressure, and do not require

a vacuum or elevated pressure; 2) heat and mass transfer between the air and the

desiccant takes place in direct contact and 3) cooling and dehumidification of the

conditioned air may be provided in different quantities. While the desiccant offers

the previously mentioned benefits, it also has several disadvantages, which include: 1)

a low COP due to inefficient regeneration; 2) the need to pump large air volumes; 3)

the need to replace the desiccant after some period of operation because the desiccant

is contaminated by dirt and dust contained in the air (Grossman, 2001, 53-62). The

critical step in the dehumidification cooling system is the reduction of water in the

air used by different desiccants. There are two basic types of desiccants which can be

used in the cooling system: solid desiccants (e.g., silica gel and solid lithium

chloride), or liquid desiccants (e.g., lithium chloride solution and glycols). The

driving force for the absorption process is the difference in vapour pressure between

the air and the desiccant (Prasitpianchai, 1999, 29-32).


16

3.2 Solid Desiccant Cooling System and Solid Desiccants

The desiccant bed is typically configured as a rotary wheel in a solid desiccant

cooling system. A conceptual schematic of a solid desiccant system is shown in

Figure 3-1.

Figure 3-1 A Conceptual Solid Desiccant Cooling System

The air to be dehumidified is passed through one side of the wheel, while a hot air

stream passes through the other side for simultaneous desiccant regeneration and

dehumidification (Pennington, 1955). After the air is dehumidified, evaporative

cooling is used to lower the air temperature as it enters the conditioned space.

Concentrated desiccant is brought continuously into contact with the air during the

dehumidification process. An extended contact surface is commonly utilised to

enhance the heat and mass transfer between the air and the solid desiccant. Water is

absorbed from the air into the desiccant, removing the latent load. The desiccant

must be regenerated to allow for repeated use (Öberg, 1998). For this reason, the

desiccant is heated to release water and is then brought into contact with the moisture

scavenging air stream during the regeneration process. Before the concentrated

desiccant returns to the dehumidification, it is cooled to minimize the heat addition

of the air to be conditioned, and to lower the desiccant‘s vapour pressure. The use of

return air rather than outdoor air for pre-cooling the dehumidified air steam in the in

the rotary heat exchanger is an alternative arrangement.


17

The advantage of using solid desiccants is that it has a higher drying capability than

liquid desiccants. Unfortunately, it requires a higher regeneration temperature (more

than 70°C). A high-pressure drop in the air stream also requires high energy system

operation. A liquid desiccant in comparison requires a lower regeneration

temperature (50-60°C) but it also has a lower degree of dehumidification.

There are several solid desiccants used in the cooling systems, such as silica gel,

molecular sieve, zeolite and activated carbon. These desiccants are chosen for their

capacity to remove moisture from air, and can be applied to air-conditioning systems

(Gaffar, 2002, 22-23). The details of the adsorption principle and different

characteristics will be presented below.

3.3 Liquid Desiccant Cooling System and Liquid Desiccants

A conceptual liquid desiccant cooling system is showed in Figure 3-2. Concentrated

desiccant is brought into contact with the air in the dehumidifier. An extended

contact surface in the dehumidifier is commonly utilized to enhance the heat and

mass transfer between the air and the desiccant. Some packed bed material is used in

this project to achieve this aim.

Figure 3-2 A Conceptual Liquid Desiccant Cooling System


18

Liquid desiccant and solid desiccant cooling systems operate using a similar

principle. However, the use of liquid desiccants offers several design and

performance advantages over solid desiccants. When solar energy is used for

regeneration, liquid desiccants have special advantages in the regeneration process.

They advantages are:

The liquid desiccant systems have a lower air pressure drop and the need for a

lower regeneration temperature compared with solid desiccant. Low

regeneration temperatures mean the system can use low-grade heat resources,

such as solar energy or exhaust heat.

The ability to pump the liquid desiccant makes it possible to connect several

small desiccant dehumidifiers to a larger regeneration unit. This would be

especially beneficial in large commercial buildings, such as central air-

conditioning system.

Prasitpianchai (1999) notes that using a liquid desiccant can also enable more

efficient heat transfer since highly efficient liquid-liquid heat exchangers may be

employed.

A liquid desiccant system does not require simultaneous air dehumidification

and desiccant regeneration, as it is possible to store the dilute liquid until

regeneration heat is available. Concentrated desiccant may be stored at room

temperature for use during the times when no source of regeneration heat is

available. (Prasitpianchai, 1999) Liquid desiccants can effectively realize energy

storage at room temperature.

Finally, an important benefit of a liquid desiccant system relative to indoor

health issues is its ability to eliminate microbial contamination, bacteria, viruses,

and moulds. Liquid desiccant dehumidifiers have been widely used in food

processing and hospitals (Prasitpianchai, 1999).


19

Low vapour pressure characterizes hygroscopic liquids which can be used as

desiccants. The driving force for mass transfer is the difference between the vapour

pressure in the air and in the desiccant. In addition to low vapour pressure, desiccants

should have low viscosity and good heat transfer characteristics. The best properties

of liquid desiccant are: 1) non-corrosive, 2) odourless, 3) non-toxic, 4) non-

flammable, 5) stable, 6) readily available, and 7) inexpensive. Furthermore, the

surface tension of a liquid desiccant is important since it directly influences the static

hold up and surface wetting in the desiccant-air contact equipment (Fumo and

Goswami, 2002, 351-361).

Several liquid desiccants are commercially available: triethylene glycol (TEG),

ethylene glycol, and brines such as calcium chloride, lithium chloride, lithium

bromide, and calcium bromide which are used singly or in combination. At this point

in time, commonly used desiccants in the desiccant cooling system are aqueous

solutions of lithium chloride, calcium chloride, mixtures of these solutions, and TEG.

Lithium Chloride

Lithium Chloride is the most stable liquid desiccant and has large dehydration

concentration (30% to 45%), but its cost is relatively high. It is expected that it will

reduce the relative humidity to as low as 15% (Fumo and Goswami, 2002). Lithium

chloride is a good candidate material since it has good desiccant characteristics and

does not vaporize in air at ambient conditions. A disadvantage with LiCl is that it is

corrosive and another problem is that carry over in LiCl liquid desiccant system.

Consequently, a conventional liquid-desiccant system must use a droplet filter or

demister to prevent carryover of desiccant out of the conditioner and regenerator. In

well-maintained systems, the droplet filter/demister will essentially eliminate

desiccant carryover (Lowenstein, Slayzak and Kozubal, 2006).

Calcium Chloride

Prasitpianchai showed (1999) that for the same mass transfer potential, calcium

chloride is the least expensive among the four desiccants: TEG, LiCl, CaCl2, and

LiBr. Another advantage of calcium chloride is its relatively low viscosity. This is


20

important for reducing the pumping power required. Furthermore, calcium chloride

is also readily available. The disadvantage of calcium chloride is its corrosion

potential due to electrolytic halide ions. An uninhibited solution of calcium chloride

has surface corrosion rates in the 2-20mils/year range for aluminium, steel, and

copper. A satisfactory rate would be less than 2mils/year. The corrosion can be

avoided by using non-metallic materials in parts of the system that will be in direct

contact with the desiccant.

Trimethylene Glycol

At the University of Florida, Öberg and Goswami conducted a study of a hybrid solar

liquid desiccant cooling system using triethylene glycol (TEG) as the desiccant

(Öberg, 1998). Their experimental work concluded that glycol works well as a

desiccant. However, pure triethylene glycol does have a small vapour pressure that

causes some of the glycol to evaporate into the air. Although triethylene glycol is

non-toxic, any evaporation into the air supply stream makes it unacceptable for use

in the air conditioning of an occupied building.

Desiccant Mixtures

Two or more liquid salt desiccants can be mixed together to achieve the most ideal

properties. Improved characteristics can be expected as well as a considerable

reduction in cost by combining different liquid desiccants. Some available desiccants

have good properties but are also expensive, such as: LiCl. Other desiccants, such as

CaCl2, have relatively poor properties compared to LiCl, but are very inexpensive.

Therefore, to achieve good properties and low cost, one solution is to mix different

desiccants together and to test these combinations. Ertas, Anderson and Kiris(1992)

investigated a Cost-Effective Liquid Desiccant (CELD, comprised of 50% each of

LiCl and CaCl2 salt content), by mixing lithium chloride (99.3%) and the calcium

chloride (90%).


21

3.4 Hybrid System

There are two kinds of load of air conditioning: latent (dehumidification) loads and

sensible (temperature control) loads. For the purpose of dehumidification

conventional vapour compression air conditioning systems may operate at a lower

temperature than what is required to meet the latent cooling load. Therefore, this

system uses more energy to reheat the overcooled air to achieve suitable temperature.

In the hybrid system, the desiccant dehumidifier handles the latent cooling load, and

a conventional vapour compression or absorption refrigeration system handles the

sensible cooling load (Öberg, 1998). The hybrid system also offers some design

flexibility in solar-powered and conventional vapour systems.

Peterson and Howell (1991) patented a hybrid liquid desiccant vapour compression

air conditioning system in 1991. This system was divided into two sections: standard

vapour-compression equipment and aqueous solutions of liquid desiccant system.

Because of the use of the circulating liquid desiccant and an adiabatic humidifier,

this hybrid system is a more efficient than traditional systems.

In the hybrid cooling system, the latent load is catered for by the dehumidifier and

only sensible cooling is carried out by the conventional air conditioning. The

possible combinations of hybrid cooling systems can be classified as follows (Gaffar,

2002):

1 Hybrid vapour-absorption/solid desiccant air conditioning systems

2 Hybrid vapour-absorption/liquid desiccant air conditioning systems

3 Hybrid vapour-compression/solid desiccant air conditioning systems

4 Hybrid vapour-compression/liquid desiccant air conditioning systems

A conceptual hybrid vapour and liquid desiccant cooling system is presented in

Figure 3-3. The cooling process is shown on the psychrometric chart Figure

3-4(Gaffar, 2002). In most hybrid desiccant air-conditioning systems, moisture is

removed from the air by bringing it into contact with the desiccant and the sensible

cooling section is circulated by vapour compression cooling systems, vapour

absorption cooling systems, or evaporative cooling systems (Mago and Goswami,

2001).


22

Figure 3-3 A Conceptual Hybrid Vapour-Liquid Desiccant Cooling System

Figure 3-4 Psychrometric Chart Depicting Hybrid Cooling Process

Figure 3-4 includes reheating process. Reheat systems are strongly discouraged in

the cooling system. Reheating is only limited to laboratory, health care, or similar

applications where temperature and relative humidity must be controlled accurately.

Dehumidification

2 1

5’ 5” 3 4

1-2-3-4 Conventional Cooling 1-5-4 Hybrid Desiccant Cooling

Reheat

Sensible Cooling

Sensible Cooling

Exhaust Heat Source

Air

Dehumidifier

Ambient Air

Condenser

Evaporator

Compressor

Expansion Valve

Conventional Air Conditioning Unit

Indoor Space

Dilute Desiccant

Ambient Air

Cooling Water

Heat Exchanger

Concentrate Desiccant

Regenerator


23

On the other hand, the exact target dry bulb temperature can be achieved by reheat

coils.

An alternative hybrid system is proposed by Costello (1976). This system can be

considered as a modified absorption system into which two vapour compressors have

been inserted (Costello, 1976). In this system, the vapour compression parts permit a

reduction in the cost of the heat and mass exchangers in the absorption parts of the

system. Unfortunately, this hybrid system still belongs to a closed absorption cycle,

and some parts requiring vacuum are need in the system. The use of open liquid

desiccant systems is more flexible in comparison to a closed absorption systems.

This research will select an open liquid desiccant system, and a conventional air-

conditioning section which uses vapour compressors to deal with sensible heat.

3.5 Mathematical Models of the Dehumidifier/Regenerator System

Stevens (1989) developed a model based on fitted algebraic equations for the

dehumidification section. Khan and Ball (1992) and Khan (1994) developed seasonal

performance simulations for the dehumidification calculation. Gandhidasan (2004)

also developed a simplified model for the preliminary design of an air

dehumidification process. These models are all focused on the packed bed

component and improving dehumidification efficiency.

In order to analyse the thermal performance of the dehumidification/regeneration

processes, a packed tower and models of the heat and mass transfer characteristics of

that packing are required. There are several types of mathematical models existing

for the liquid desiccant regenerator. One type is the empirical model. These models

are easy to formulate using experimental data, but they are limited to the equipment

and range of conditions for which the data is taken. Another type is the finite-

difference model which requires fewer assumptions, and involves more calculations.

Most manufacturers will provide a computationally simple model for their products,

but this method relies on some factors which depend on the mass flow rates and the


24

size of the heat mass exchanger. The range of theoretical models describing packed

bed dehumidification is summarized in Table 3-1.

Table 3-1 Theoretical models for packed bed absorbers (Öberg, 1998)

Name Model

Verifying

Experiments Additional Comments

Factor and Grossman

(Factor and Grossman1980,

541-550)

finite difference Yes Slug flow, temperature and concentration

gradient in flow direction only, adiabatic

process, negligible heat and mass transfer

resistances in the liquid phase, the surface

area for heat and mass transfer is the same.

Gandhidasan et al.

(Gandhidasan et al 1987,

89-93)

finite difference No In addition to the assumptions by Factor and

Grossman, the resistance to mass transfer in

the liquid phase was considered.

Khan and Ball (Khan and Ball

1992, 525-533)

based on algebraic

correlations

No The correlations were obtained from data

obtained using a finite difference model.

Sadasivam and Balakrishnan

(Sadasivam and Balakrishnan

1992, 572-577)

Effectiveness-NTU Yes Negligible change in the liquid flow rate

throughout the tower, unit Lewis number, and

linear saturated air enthalpy versus

temperature curve.

Stevens et al. (Stevens,

Braun and Klein ,1989).

Effectiveness-NTU Yes Same as those by Sadasivam and

Balakrishnan


25

3.6 Different Control Strategies for Indoor Air Humidity

Air-conditioning demands efficient control of both temperature and humidity.

Conventional air-conditioning systems cool the air below its dew point to reduce the

moisture content (Subramanyam et al., 2004, 2679-2688). Sometimes this process is

followed by reheating the dehumidified air in order to get the desired humidity level

in the conditioned space. For human comfort, the relative humidity must be within a

specified range. Applications like libraries, museums, and computer rooms, require

not only low temperatures but also low humidity (Subramanyam et al., 2004, 2679-

2688). ASHRAE(ASHRAE, 1992) and ISO (ISO7300) standards also define a

comfort zone for the human being based on overall heat balance. The acceptable

thermal environment of indoor spaces designed for human occupancy is dependent

upon operating temperature and relative humidity.

To achieve comfortable thermal air conditions, there are various techniques for air

dehumidification. Traditionally, latent loads and sensible loads are treated in a

coupled way (Zhang, 2006, 1228-1242). Humidity control, though with certain

limitations, can also be achieved by air bypass control, variable speed fans and

capacity control of the compressor (Shirey, 1993, 694-703). There are three

conventional methods to control indoor air humidity. One method that can be used is

the variable air volume control technology (Chua et al., 2006). In the following

research only liquid desiccant is used to dehumidify the air after dehumidifier.

Therefore the capacity control of compressor technology is not further discussed in

this thesis.


26

3.6.1 Bypass air (BA) control

The schematic of bypass air control methods is shown in Figure 3-5. The bypass air

Indoor

Return Air

Exhaust Air

Recirculate

Air

Ambient Air

Bypass

Air

Control

Coolant

Cooling

Coil

Control

Air

DampersSupply Air

Figure 3-5 Schematic of Three Different Humidity Control Strategies

should include a cooling coil into the system to control humidity. The mixture of

ambient air and re-circulated air is disposed of by the cooling coil for

dehumidification. Air flow passes by the coolant (such as, chilled water) and is

dehumidified by the cooling coil. The dampers in the bypass section can control air

flow rate into the cooling coil or the air that passes through the bypass channel. The

air after dehumidifier and the bypassing air mix together to satisfy indoor

requirements. Air flow volume depends on the dampers section control. A bypass

system is generally restricted to small installations where a simple method of

temperature control as well as a modest initial cost is desired, and energy

conservation is less important (ASHRAE, 2001b).


27

3.6.2 Variable Air Volume (VAV) control

Variable air volume air-handling systems are commonly used for conditioning and

delivering the air to occupied zones. VAV boxes are an integral part of such systems

and are the final piece of equipment that air passes through prior to reaching the

occupants (Schein and House, 2003). Variable air volume systems can be applied to

interior or perimeter zones, with common or separate fans, with common or separate

air temperature control, and with or without auxiliary heating devices (ASHRAE,

2001b). Humidity control is a potential problem with VAV systems. If humidity is

critical, as in certain laboratories, process work, etc., systems may have to be limited

to constant volume airflow (ASHRAE, 2001b). Based on the simulation results from

the coil model, it has been observed that VAV control yields the best coil

dehumidification performance (Chua et al., 2006). Another advantage for VAV

control is that it requires less fan power and is less energy intensive (Mumma and

Bolin, 1997, 463-470). Because VAV boxes can be adjusted airflow volume

according to requirement, they are used as the supplied air section in this research.

3.7 Concluding Remarks

The current desiccant cooling and dehumidification concepts were discussed in this

chapter. In comparing desiccant and conventional systems, conventional air

conditioning were shown to require a large amount of electrical energy to dispose of

the latent cooling load in humid climates. Using a desiccant to dispose of cooling

load can save electrical energy. There are several desiccant cooling systems reported

in literature, such as the liquid desiccant and solid desiccant cooling systems. Liquid

desiccant cooling has good ability to regenerate desiccant at relatively low

temperatures. It provides opportunities for a cooling system using solar energy to

recycle liquid desiccant. Therefore, the liquid desiccant system has been selected for

the following study. LiCl is chosen as the candidate desiccant because of its superior

characteristics. The proposed hybrid cooling system comprises of desiccant

dehumidification and conventional cooling into one system. This design can separate

the latent cooling load and sensible load disposal.


28

An air conditioning cooling system should have the ability to provide efficient

control of indoor humidity and temperature, while at the same time minimizing the

electrical energy requirements for air conditioning. Therefore, solar energy

regenerating liquid desiccant combined with conventional air conditioning offers a

great potential, since the sensible and latent cooling loads can be individually

handled in an effective manner with minimum energy impact. Regardless of the

system type, additional studies are warranted in order to further advance the energy

savings and environmental benefits which may be achieved with liquid desiccant

cooling technologies (Öberg, 1998).

Although hybrid desiccant cooling systems can offer certain benefits over other

cooling techniques, conventional liquid desiccant technology do not have adequate

flexibilities for the control of humidity. Nevertheless, widespread utilization of other

humidity control technologies can be used in the hybrid system. Several different

humidity controlling technologies are mentioned in this chapter. As discussed in

chapter 2, the mixed air design to solve conventional liquid desiccant system

problems will be incorporated into conventional hybrid cooling system.

When a mixed air design is incorporated into a hybrid liquid cooling system, the

ambient air supply has a large impact on the overall system performance. New

hybrid system performance requires analysis of ECOP, TCOP and COPsys to

properly evaluate the new system. Because humidity absorption and desorption occur

in a randomly packed tower and include heat and mass transfer between the gas and

liquid phases, variables such as air temperature, humidity, desiccant concentration

and flow rates all affect the system performance. Thus, the experimental study of the

liquid desiccant system should also be undertaken to compare the theoretical

modeling analysis with experimental results.


29

4 PERFORMANCE STUDY OF HYBRID LIQUID DESICCANT SOLAR COOLING SYSTEM

This chapter presents a theoretical study concerning the performance analysis of a

hybrid cooling system that is based on the AHU mixed air humidity control design.

Theoretical results and predictions of this system are presented, with theoretical

results being compared to experimental data. Further analysis provides a

juxtaposition of a conventional air conditioning system, with the hybrid cooling

system performance as discussed in this chapter. Performance analysis is provided

using the ECOP, TCOP and COPsys parameters, which will be reviewed in the

hybrid liquid desiccant cooling system section. Correlations between the

conventional air conditioning system and the hybrid liquid desiccant cooling system

are presented. These correlations provide elucidation of performance estimates and

provide analysis of energy trends concerning operation of the hybrid liquid desiccant

cooling systems.

4.1 Introduction

The research program discussed in this chapter is divided into three sections. The

first section involves reviewing reasons for choosing the COP model for the

conventional air conditioning system, and the ECOP model for hybrid liquid

desiccant cooling system. A theoretical definition for the conventional cooling

system COP is developed to describe electrical energy cost and performance. In this

research, the conventional cooling system performance model is based on the

DAIKIN air conditioning specification (Daikin, 2004). This decision is validated by

the Australian Gas Light Company (AGL) Energy Shop, which recommends the

Daikin cooling system for energy efficiency reason (AGL, 2008a). It should be noted

that other evaporative units, such as provided by LG can be used to reduce air

humidity (AGL, 2008b), they are however more expensive in operation.

Cooling load and thermal energy input are used to identify the performance of the

conventional cooling section in a hybrid cooling system. The hybrid liquid desiccant

cooling system, which includes a conventional cooling section, uses ECOP to


30

describe the hybrid system performance in this chapter. The performance relationship

between the conventional air conditioning COP and hybrid liquid desiccant cooling

system ECOP are included in this chapter.

The second research section concerns TCOP analysis of the hybrid liquid desiccant

cooling system. This is provided because thermal energy input is not included in the

conventional air conditioning; subsequently TCOP is not calculated for the

conventional air conditioning system. In the hybrid liquid desiccant cooling system,

solar energy or electricity heat source is used as thermal energy in the hybrid liquid

desiccant cooling performance study. When solar energy is used as input thermal

energy, the regeneration heat source comes from the sun and is gained freely, where

0TQ . As a result of this energy characteristic, TCOP discussion is not included in

the hybrid liquid desiccant cooling system. When electrical heat is used as the

thermal energy input in the hybrid system, 0TQ , TCOP should therefore be

analysed. Cooling load in TCOP calculations is similar to the procedures used in

calculating ECOP. Some sensible and latent heat theory is therefore used to calculate

cooling load. The heat gains occurring in a room can be considered in two ways: as

sensible gains and latent gains (Jones, 2001). TCOP calculation is dependent on

latent gains and regeneration heat input. TCOP and cooling load correlations are

shown in this section.

The third section involves a combined system analysis. The system COP model is

based on energy balance to calculate the total energy consumption for the system.

Conventional air conditioning COP is compared with Hybrid cooling system COP.

Because the hybrid cooling system includes a conventional cooling and liquid

desiccant dehumidification sections, the system COP is compounded by conventional

cooling and dehumidification energy. This research program reveals that the chosen

variables such as cooling load, TCOP and ECOP, have the greatest significance in

affecting the system performance.

The COP theoretical analysis, some relative experimental data, such as ambient air

temperature and humidity are used. This arises because the real process in the hybrid

liquid desiccant system has different trends, and the dehumidification process is also


31

complex. Therefore, it is necessary to compare experimental data, with experimental

results to support the AHU design used in the hybrid system. Two experiments are

performed to provide comparison with the theoretical model. The experiments are

the regeneration test, and the dehumidification test. This comparisons and analysis is

presented in Chapter 5.

4.2 Model Formulation

The conventional coefficient of performance, or COP (sometimes CP), of a

refrigerator is the ratio of input heat to the evaporator, to the supplied work from the

compressor. The COP is therefore mathematically represented by Equation 4-1:

Equation 4-1 Q

COPW

From Equation 4-1, Q is the useful cooling supplied by the evaporator, and W is the

work consumed by the compressor. In this research, the conventional air

conditioning and hybrid liquid desiccant cooling system are included in performance

analysis. The performance of the conventional air conditioning is defined as COPcon,

while the hybrid liquid desiccant cooling system can be evaluated by means of the

electrical COP, thermal COP, and system COP. Corresponding definitions are shown

in Equation 4-2 (points1,2 in Figure 4-5) through to Equation 4-5 (Equation 4-2, 4-3

and 4-5 points in Figure 4-3) respectively, according to these refrigeration

coefficients.

Conventional air conditioning system:

Equation 4-2 1 2 ( )con acon

con con

Q m h hCOP

W W

Hybrid liquid desiccant cooling system:


32

Equation 4-3 ECOP = 3 4 ( )c a

c c

Q m h h

W W

Equation 4-4 TCOP 3 4 ( )c a

T T

Q m h h

Q Q

Equation 4-5 sysCOP = c

T c other

Q

Q W W

In these equations, the Q subscript ― con‖ denotes the conventional air conditioning

system, ― c ‖ the hybrid cooling system and ―T‖ thermal cost of whole system. The

COP is defined as useful heat moved or obtained, divided by the energy required to

drive the process for the cooling system. For a refrigeration cycle, the useful heat is

the refrigeration effect (Haines and Wilson, 1998, 462-463). In the classical sense,

the useful heat is moved by the evaporator absorbing heat from the cooling coil

during the process of evaporation (Ahmed, 1996, 76-77). Thus, the heat per unit

moved by the evaporator ( cQ ), can be defined by the enthalpy difference ( 3 4h h )

between the evaporator inlet and outlet. The other reason cQ should be defined by

the enthalpy difference between the evaporator inlet and outlet rather than the

enthalpy difference between ambient air and indoor air ( 1 4h h ) in the hybrid system,

is because the air after dehumidifier enthalpy has changed after the liquid desiccant

dehumidification and air mixing process. The air enthalpy difference between the

inlet and outlet of the evaporator is the real energy change by the conventional air-

conditioning section in the hybrid system. Daikin air conditioning COP data can be

used here. Ahmed (1996) evaluated the system COP in the hybrid system in a similar

manner. Therefore, in the following study, cQ per unit is defined by enthalpy

difference ( 3 4h h ) associated with the evaporator in the conventional air-

conditioning section. Total cooling load cQ is defined as 3 4 ( )am h h . In general, the

COP in this research is only used to evaluate the energy situation in this hybrid

system, and it is not used to compare with conventional air conditioning efficiency.


33

1 1 4 4, , , ,

Dehumidification Data

T T

2 2, ,from Experiment T

3 3, ,CalculationT

, CalcualtionotherW ECOP

1 2 3 4, , , ,Calculationh h h h

2 2

3 3 4 4

Compare , with

, and ,

T

T T

, CalculationsysTCOP COP

Figure 4-1 COP Calculation Flow Chart


34

The hybrid system integrates a conventional cooling system and air mixture AHU.

The hybrid desiccant system offers the advantages of independent control of

temperature and humidity and therefore reduces the energy cost and equipment size

(Elsayed etc, 2006). However, calculation of the coefficient of performance becomes

complex compared to a simple cooling system. The total hybrid system calculation

flowchart is displayed in Figure 4-1. The test points that provide system values for

this flowchart to be followed are shown later in Figure 4-2.

In the COP study, various input data such as the ambient air temperature and

humidity need to be confirmed before calculation. Some parameters, such as 2 2,T ,

depend on the experimental situations. 3 3,T can be calculated by using the air

mixture section theory in the air conditioning. ECOP and otherW are determined by

3 3,T . ECOP is used to describe the conventional air conditioning section

performance within the hybrid liquid desiccant cooling system, while COPcon is used

to describe only the conventional air conditioning system performance. The enthalpy

of different points in the mixing psychrometric chart can also be calculated. The final

step concerns the calculation of TCOP and COPsys.

4.3 System Description

The outside airflow that enters a building or zone by an air-handling unit can be

described by the outside air fraction oaX (Equation 4-6), which is the ratio of the

volumetric flow rate of outside air brought in by the air handler to the total supply

airflow rate.

Equation 4-6 oa oaoa

sa oa ca

Q QX

Q Q Q

When expressed as a percentage, the outside air fraction is called the percent outside

air. The design of outside airflow rate for a building‘s ventilation system is found


35

through evaluating the requirements of ASHRAE Standard 62. The supply airflow

rate is that required to meet the thermal load. The outside air faction and percent

outside air then describe the degree of recirculation, where a low value indicates a

high rate of recirculation, and a high value shows little recirculation. Conventional

all-air, air-handling systems for commercial and institutional buildings have

approximately 10 to 40% outside air (ASHRAE, 2001a).

100% outside air means no recirculation of return air, which is discharged directly to

the outside as relief air. An air-handling unit that provides 100% outside air is

typically called a makeup air unit (MAU) (ASHRAE, 2000a). In order to compare

performance calculation results, a conventional air conditioning system and hybrid

liquid desiccant cooling system are represented as a MAU model. There is no

recirculation of return air included within the system.

A comparison of system performance between the hybrid and conventional systems

is provided in Figure 4-2 and Figure 4-4. The corresponding state-points at these test

points in shown on the psychrometric charts of Figure 4-3 and Figure 4-5,

respectively. In this analysis, the conventional system and hybrid system are all

selected to have an indoor situation of (24°C, 50%), and outdoor situation of (33°C,

45%). The after evaporator air situation is also assumed to be 12°C 90%. However,

because the AHU ambient air mixing and dehumidification unit are used within the

hybrid cooling system to compare the conventional air conditioning system, the

relative air disposed paths from point 1 to 5 on the psychrometric charts show a

significant difference between the two systems.


36

Regenerator

Dehumidifier

AHU

Evaporator

Ambient

Air

Waste Air

Out

..2

.

Liquid

Desiccant

.. Indoor

Air.3 4Ambient

Air .1

Spill Air and

Leakage Air

Ambient

Air

33°C 45%14g/kg

33°C 45%14g/kg

33°C 45%14g/kg

27°C 50%11g/kg

24°C 50%9.5g/kg

.33°C 45%14g/kg

35°C 15%5g/kg

34.2°C 24%8g/kg 12°C 90%

8g/kg 5

.6

Figure 4-2 Different Experimental Test Points in the Hybrid Cooling System

1

3

2..

4

.

33°C 45%14g/kg

35°C 15%5g/kg

34.2°C 24%8g/kg

12°C 90%8g/kg

5

6

24°C 50%9.5g/kg

27°C 50%11g/kg

.

h1= 70kJ/kg

h6= 55.5kJ/kgh3= 54.5kJ/kg

h5= 48kJ/kg

h2= 48.5kJ/kg

h4= 31.5kJ/kg

.

.

Figure 4-3 Hybrid AC Air mixture and Cooling Situations in Psychrometric

Chart


37

Condenser

Evaporator

Ambient

Air

..1

Indoor

Air

2

Spill Air and

Leakage Air

Ambient

AirExhaust

33°C 45%14g/kg

12°C 90%8g/kg

.24°C 50%9.5g/kg

.27°C 50%11g/kg

5

6

Figure 4-4 Different Experimental Test Points in the Conventional AC System

2

.

33°C 45%14g/kg

12°C 90%8g/kg

5

6

24°C 50%9.5g/kg

.

h1= 70kJ/kg

h6= 55.5kJ/kg

h5= 48kJ/kg

h2= 31.5kJ/kg

.27°C 50%11g/kg

1.

Figure 4-5 Conventional AC Air Cooling Situations in Psychrometric Chart


38

4.4 Standard Assumptions

There are some standard assumptions and constraints used in most formulations

including the experimental chapter presented in this thesis. These assumptions are:

1. All the inlet air state properties of different entrances are considered at a

steady state for each period in this research.

2. All air streams are considered to be mixed properly in the mixture.

3. Pressure drop effects upon air velocity in the direction of flow are negligible.

(Abbud, 1999)

4. The liquid desiccant solution is a homogeneous liquid sorbent.

5. There are no mixing or liquid desiccant carry over or leakage problems

between different disposal procedures.

6. The air after the dehumidifier and liquid desiccant in the

dehumidifier/regenerator can be described as ideal air and ideal liquid

situations(Öberg, 1998).

7. Heat and mass transfer only occur in the liquid desiccant and the air after

dehumidifier contact boundary(Öberg, 1998).

8. The friction between liquid desiccant and the air after dehumidifier is

negligible.

9. The liquid desiccant sorption hysteresis is neglected and the heat of

adsorption is a single-valued function of the air stream (Gaffar, 2002).

10. There is no flux coupling when both mass and heat transfer occurs.

11. Steady state performance of the liquid desiccant system.

12. The properties and transport parameters of the air after dehumidifier and

desiccant material are constant (Grossman, 2001).

13. The dehumidifier/regenerator is adiabatic.

14. Whole system thermodynamic equilibrium between the air after dehumidifier

and the liquid desiccant at all times (Öberg, 1998).

15. Heat lost during heat transfer between the liquid desiccant and the contact air

is neglected (Öberg, 1998).


39

4.5 Air Mixture Modelling and Air Enthalpy Calculation

The variable air volume (VAV) air-conditioning system has been deemed more

economical than alternative systems and has been widely adopted in buildings to

maintain cooling and heating demands (Qin and Wang, 2005, 1035-1048).

Conventional VAV systems are used to adjust supply air volume to indoor

requirements to achieve economical cooling. In this study, the air mixture design in

the AHU section uses a VAV system to control the new air volume to the air after

dehumidifier mix. The new air mixture will supply the indoor area when it satisfies

indoor air requirements. In the mathematical modeling, the air mixture theory for air

conditioning is used. The definition of enthalpy is also described in this section.

Temperature, humidity rate, flow rate and enthalpy are four factors that influence

system performance, and can be easily analysed from the mathematical expression.

The equations in this section are identified according to research hypotheses and

system design.

4.5.1 Air Mixture Modelling

AHU controlling data is based on the air mixture model analysis. According to the

model, the air mixture states can be calculated. Eight variables are involved in the

control of the AHU terminal program: (1) supply air temperature 4t , (2) supply air

humidity 4 , (3) required ambient air flow rate 1m , (4) required ambient air

temperature 1t , (5) required ambient air humidity 1 , (6) the air after dehumidifier

flow rate 2m , (7) the air after dehumidifier temperature 2t , and (8) the air after

dehumidifier humidity 2 . In the following section, these parameters are discussed.

Figure 4-3 shows the situation when the system air stream is disposed from point 1 to

point 4 on the psychrometric chart. In this model, point 1 is the ambient air state,

point 2 is the air after dehumidifier state, point 3 is the air mixing state and point 4 is

the inlet air state. Dry air at state 2 mixes with moist air at state 1, forming a mixture

air at state 3. The principle of the conservation of mass allows two mass balance


40

equations to be written: 1 2 3m m m for the dry air and 1 1 2 2 3 3m m m for the

associated water vapour, ( 1 1,m is the air mixing mass flux and humidity ratio).

Hence

Equation 4-7 1 3 1 3 2 2( ) ( )m m

Therefore

Equation 4-8 1 3 2

3 2 1

m

m

Similarly, making use of the principle of the conservation of energy,

Equation 4-9 1 3 2

3 2 1

h h m

h h m

The three state points must lie on a straight line in a mass energy co-ordinate system

(Jones, 2001). When two airstreams mix adiabatically, the mixture state lies on the

straight line that joins the constituent state points. The position of the mixture state

point is such that the line is divided inversely as the ratio of the masses of dry air in

the constituent airstreams.

4.5.2 Air Mixture Parameters

The process of mixing air occurs in a short time period. When two airstreams mix

adiabatically, the mixture state lies on the straight line joining the state points of the

constituents according to Figure 4-3. The requirements of supply air are based on the

assumption of having the absolute humidity at 0.008kg/kg, and the temperature as 12

C. This air state has a relative humidity of 90%, at 12 C. Because this is a hybrid

cooling system, conventional air conditioning is used to control and achieve the

required temperature. This research only focuses on dehumidification and achieves


41

suitable humidity. Therefore, the purpose of the air mixture is to accurately control

supply air moisture content.

Because the relative humidity is defined as the ratio of the current vapour pressure of

water to the equilibrium vapour pressure (or saturation vapour pressure), air pressure

and air temperature will influence the relative humidity. Therefore, in the following

analysis, we change all the relative humidity experimental data into absolute

humidity under relative air temperature situations. In the air-conditioning system,

supply air should have a higher air pressure than the ambient air to keep ambient air

out of the indoor space. The ―positive pressure‖ needed in the supply air system is

normal, which is about 5-10 Pa. This is only a small pressure difference and the

influence of the relative humidity can be neglected. The air after dehumidifier enters

the air mixture chamber at state 2 (as shown in Figure 4-2). Because the fan system is

stable, the air after dehumidifier flow rate is used according to the following

experimental results found in this research:

. 0.269air inS m3/s

Therefore

. 269air inS L/s

While the density of air will change as the temperature changes, this change is less

than 2% from 21 °C to 27 °C. Here, the selected air humidity (28 °C and 100 kPa RH

70%) density is:

air = 1.168 kg/m3

Here: .air inS is inlet average air flow rate (m

3/s)

air is density of air (kg/m3)

and, the air after dehumidifier mass flow 2m can be written as:

2 .air in airm S


42

Applying all the previously discussed parameters, the following equation can be

derived:

2m = 0.314 kg/s

According to Equation 4-8, we can write:

3 21 2

1 3

m m

After the ambient air mix (state 1), the air mixture becomes as in state 3 and the

humidity of state 3 is 0.008kg/kg as assumed by:

3 = 0.008kg/kg

Therefore, the mixed ambient air mass flow can be calculated as:

Equation 4-10 21

1

0.0080.314

0.008m

Chapter 5 shows that the result of this calculation is in accordance with Equation

4-10. All of the humidity data is based on experimental results. Air mixture

parameters can be used in the ECOP, TCOP and COPsys calculation and analysis for

the hybrid system. The range of humidity is between the air after dehumidifier and

air mixing state. Since this is a hybrid cooling system, the temperature of the indoor

air is controlled by a conventional air conditioning unit. The temperature range can

therefore be controlled between 0°C and the ambient air temperature.

4.5.3 Air Mixture Enthalpy Calculation

In thermodynamics, the enthalpy or heat content (denoted as h or Δ h ) is a quotient

or description of the thermodynamic potential of a system. This can be used to

calculate the system energy obtainable from a closed thermodynamic system under


43

constant pressure. In this study, enthalpy is used to calculate and evaluate the

ambient air and the air after dehumidifier energy situation for the hybrid cooling

system.

The enthalpy, h , used in Psychrometry is the specific enthalpy of moist air,

expressed in kJ/kg dry air, defined by Equation 4-11 (Jones, 2001):

Equation 4-11 a gh h h

Where ah is the enthalpy of dry air and gh is the enthalpy of water vapour. Both are

expressed in kJ/kg, and is the moisture content in kg/kg.

An approximation equation for the enthalpy of dry air over the temperature range

0°C to 60°C is (Jones, 2001):

Equation 4-12 1.007 0.026ah t

However, for lower temperatures (down to -10°C) the approximation equation is

(Jones, 2001):

Equation 4-13 1.005ah t

Because the air temperature is always above 0°C in this study, the equation for

enthalpy of dry air is selected as Equation 4-12.

Values of gh for the enthalpy of vapour over water have been taken from NEL steam

tables (1964). These values have been slightly increased to account for the influence

of barometric pressure and modified to fit the zero datum.

For the purpose of the approximation calculation, without recourse to the CIBSE

psychrometric tables, we may assume that in the temperature range 0°C to 60°C, the

vapour is generated from liquid water at 0°C and that the specific heat of superheated


44

steam is constant (Jones, 2001). The following equation can then be used for the

enthalpy of water vapour:

Equation 4-14 2501 1.84gh t

Equations 4-15 to 4-164 can now be combined, as typified by Equation 4-17, to give

an approximation for the enthalpy of the humid air at a barometric pressure of

101.325kPa:

Equation 4-17 (1.007 0.026) (2501 1.84 )h t t

4.6 Conventional Air Conditioning Performance and Hybrid

Cooling System Electrical Coefficient of Performance

Liquid desiccant cooling has the ability to provide an efficient control of indoor air

humidity and temperature, while at the same time reducing the electrical energy

requirements for air conditioning (Öberg, 1998, 29-32). However, the hybrid liquid

desiccant cooling system still needs some electricity to run the conventional air

conditioning to dispose of the sensible cooling load. Therefore, ECOP is a more

important performance parameter since it shows the consumption of high quality

electrical energy for cooling. It is also necessary to compare hybrid system ECOP

with conventional air conditioning system COPcon to understand the difference

between the two systems.

4.6.1 COPcon and ECOP Definition

The Coefficient of Performance for conventional air conditioning system is defined

as the ratio of the cooling load, to electrical energy input to the system. The COPcon

of this system can therefore be defined by Equation 4-18:

Equation 4-18 1 2 ( )con acon

con con

Q m h hCOP

W W


45

Here conQ per unit is the rate of heat removal from the cooling section in the

conventional air conditioning. This equals the point 1, and point 2 (show in Figure

4-4) enthalpy difference ( 1 2h h ) between the evaporator. conW is the total electricity

cost used in conventional air conditioning system.

In similarity to COP calculation, Electrical Coefficient of Performance (ECOP) for

the hybrid liquid desiccant cooling system is defined as the ratio of the cooling load

to the electrical energy input to the system. The ECOP can therefore be calculated

according to Equation 4-19

Equation 4-19 3 4 ( )c a

c c

Q m h hECOP

W W

Here cQ per unit is the rate of heat removal from the sensible cooling section in the

hybrid cooling system and equals the point 3, and point 4 (shown in Figure 4-2 and

Figure 4-3) enthalpy difference ( 3 4h h ) of the evaporator. Point 3 and 4 are the

states of the air at the exit of the AHU and supply air to the indoor environment. cW

is the electricity cost rate of the conventional air conditioning section compressor,

which is used in the cooling process.

4.6.2 COPcon, ECOP Results and Discussion

To determine the comparison between the electrical performance of hybrid cooling

system and conventional air conditioning system, the same vapour compression

cooling unit is used for the two systems. This research selected the Daikin RMK140J

series model as vapour compression cooling unit. RMK140J specification can be

found in the Table 4-1. Total cooling capacity range is 7.5-18.9kW and power supply

is 220-230-240 V, 50Hz.


46

Table 4-1 DAIKIN Air Conditioning Specification(RMK140J Series Composition)

Capacity Range 5HP

Outdoor Unit (Combination Model Name) Max. 7 units (1 unit only is impossible)

No. of Indoor Units to be Connected Max. 3 units

Total Cooling Capacity Index of Indoor Units to be Connected

7.5-18.9kW

BP Unit (Combination Model Name) BPMK928A42 BPMK928A43

No. of Indoor Units to be Connected Max. 2 units Max. 3 units

Total Capacity Index 7.5-18.9kW 7.5-18.9kW

According to Equation 4-20, 1 2 ( )con acon

con con

Q m h hCOP

W W

, the conventional air

conditioning COP change depends on the cooling load conQ , and system energy input

conW . Details of the cooling load can be found in Table 4-2. Results from the

experimental study, the theoretical modeling of the conventional air conditioning

cooling load, and COP relationship process are depicted graphically in Figure 4-6.

Figure 4-6 Conventional Air Conditioning Cooling Load and COP Relationship

12.31

12.65

12.26

9.60

12.0614.20

15.98

5.0

4.2

4.0 4.8

3.1

3.8

2.5

0.0

1.0

2.0

3.0

4.0

5.0

6.08.00

9.00

10.00

11.00

12.00

13.00

14.00

15.00

16.00

17.00

18.00

23.3°C 24.8°C 25.9°C 27.8°C 29.0°C 30.0°C 31.6°C

CO

Pc

on

Co

olin

g L

oa

d (k

W)

Outdoor Temp (°C)

Conventional Air Conditioning Cooling Load and COPcon Relationship

Cooling load COPcon

Cooling Load

Qc (kW)

COP


47

Table 4-2 Input data on the performance of conventional air conditioning and hybrid liquid desiccant air conditioning*

T1 RH1 AH1 m1 QC (W/kW) QT (W/kW) WC (W/kW) Wother (W/kW)

Outdoor Air Temp. (°CDB)

Outdoor Air Relative

Humidity (%)

Outdoor Air Absolute

Humidity (g/kg)

Supply Air Mass Flow Rate (kg/s)

Cooling Load (W/kW) Thermal Energy

Input (W/kW) Cooling Electrical

Input (W/kW) Total Other Electrical

Input (W/kW)

CAC HAC CAC HAC CAC HAC CAC HAC

23.3 92 16.6 0.0286 12307.5 7549.0 0.0 4000.0 2471.4 1125.0 0.0 100.0

24.8 83 16.4 0.0282 12650.7 7626.9 0.0 4000.0 3048.4 1201.1 0.0 100.0

25.9 71 15.0 0.0408 12255.2 8087.9 0.0 4000.0 3102.6 1446.8 0.0 100.0

27.8 45 10.5 0.0698 9604.0 8079.8 0.0 4000.0 2000.8 1513.1 0.0 100.0

29.0 54 13.6 0.0390 12061.6 7571.6 0.0 4000.0 3903.4 1314.5 0.0 100.0

30.0 60 16.1 0.0275 14202.1 7319.8 0.0 4000.0 3777.1 1338.2 0.0 100.0

31.6 56 16.5 0.0469 15975.8 8315.1 0.0 4000.0 6467.9 1718.0 0.0 100.0

CAC=Conventional Vapour Compression Air Conditioning; HAC=Hybrid Liquid Desiccant Air Conditioning; QC=Total Cooling Load on the system (W/kW); QT=Total Thermal Energy Input (W/kW); WC=Total Cooling Electrical Input (W/kW); Wother=Total Other Electrical Input (W/kW) *Outdoor air data is from Australia Commonwealth Bureau of Meteorology


48

This figures show that when outdoor temperature increases, the cooling load will

increase from 12.31kW to 15.98kW. During the same time period, the whole system

COP decreases from 5.0 to 2.5. This indicates that when the outdoor temperature is

high, cooling load will also increase. The conventional system COP will therefore be

reduced.

In similarity to COP, the hybrid liquid desiccant cooling system ECOP can be

described graphically and is shown in Figure 4-7. The hybrid system ECOP is a

function of cooling load (kW), where ECOP results predicted in each figure

accompany the changing cooling load (kW). The figure shows a similar this trend.

For example, when the low outdoor temperature is 23.3ºC, Figure 4-7 shows the

cooling load (kW) approximates to 7.5kW. The ECOP rises from 4.8 to 6.7, but the

total trend of ECOP is decreasing as the cooling load has increased. Therefore, when

the cooling load increases to 8.08kW, ECOP should be decreased to a low level.

From Figure 4-7, ECOP can be found that minimum decreasing to 5.3.

Figure 4-7 Hybrid Air Conditioning Cooling Load and ECOP Relationship

Hybrid Air Conditioning Cooling Load and ECOP Relationship

8.32

7.327.63

7.55

7.57

8.08

8.09

4.8

6.7

6.4

5.65.8

5.35.5

6.00

6.50

7.00

7.50

8.00

8.50

9.00

23.3°C 24.8°C 25.9°C 27.8°C 29.0°C 30.0°C 31.6°COutdoor Temp (°C)

Co

olin

g L

oad

(kW

)

0.0

1.0

2.0

3.0

4.0

5.0

6.0

7.0

8.0E

CO

P

Cooling load ECOP

Cooling Load

Qc (kW)

ECOP


49

The experimental study of the conventional air conditioning COPcon, and hybrid

liquid desiccant cooling system ECOP showed that the following variables

significantly influence the system performance: the whole system cooling load (kW)

and outdoor temperature. The COPcon and ECOP generally increase with decreasing

cooling load (kW). Figure 4-6 and Figure 4-7 both show this trend and COPcon range

changes from 5.0 to 2.5, while ECOP changes from 6.7 to 4.8. At the same time, the

conventional air conditioning system cooling load also changed from 12.31kW to

15.98 kW. The cooling load of the hybrid system changed from 7.55kW to 8.32kW.

Conventional system cooling load is much higher than the hybrid system because the

conventional system cooling load includes not only sensible load, and but also latent

load. On the contrary due to liquid desiccant dehumidifier reasons, the latent load in

the hybrid system has been removed before the air after dehumidifier passes the

evaporator. This means the evaporator disposes only the sensible cooling load. This

is the reason why the hybrid system cooling load is much less than the conventional

system. The other parameter affecting the system performance is outdoor

temperature. From Figure 4-6 and Figure 4-7, outdoor temperature has increased

from 23.3ºC to 31.6ºC. At the same time, the cooling load of the conventional air

conditioning system, and the hybrid liquid desiccant cooling system respectively

increased from 12.31kW to 15.98 kW, and 7.55kW to 8.32kW. The hybrid cooling

system ECOP decreased from 6.7 to 4.8 with the outdoor temperature increasing in

Figure 4-7. The conventional air conditioning cooling system COPcon has also shown

a decreasing COPcon with the outdoor temperature increasing. When the outdoor

temperature increases, sensible cooling load will also be increased. With this

environment, the total cooling will be raised when the latent load maintains stability

and system performance will therefore be decreasing.

4.7 Thermal Coefficient of Performance

4.7.1 TCOP Definition

Thermal coefficient of performance can be introduced into hybrid cooling systems to

evaluate their thermal performance. Thermal energy input is not included in

conventional air conditioning system; therefore, TCOP is only used to evaluate the


50

hybrid liquid desiccant cooling system. With respect to the system studied here, the

regeneration heat source is free solar energy, therefore TQ = 0 can be assumed in the

system. TCOP does not need be analyzed when the system uses free solar energy.

When the system uses electricity to regenerate liquid desiccant, 0TQ , and TCOP

becomes one parameter required to evaluate the thermal energy usage of the hybrid

cooling system. Thus, in this study, the thermal performance of the system is only

used to evaluate the hybrid cooling system and when the system uses electricity to

work as a thermal resource.

One TCOP model analysed in this thesis is a simplified model that is based on the

mass and energy conservation equations. This TCOP is shown in Equation 4-21:

Equation 4-21 TCOP 3 4 ( )c a

T T

Q m h h

Q Q

In order to compare ECOP, TCOP and COPsys, cQ should be the same definition

with ECOP. Here cQ is the rate of heat removal from the sensible cooling in the

evaporator. cQ is the sensible cooling load, and cQ per unit equals the point 3 and

point 4 enthalpy difference ( 3 4h h ) in the evaporator. 3h and 4h represent the

enthalpy of the air mixture, and the enthalpy of the air at the inlet situation. TQ is the

energy rate of the total thermal used during the cooling process.

Due to solar energy being freely available, the thermal energy in TCOP calculation is

only from electricity. Therefore the energy input in this experimental study of the

dehumidification section is 4.0 kW, and cooling load changes over a range from

7.55kW to 8.32 kW.

According to the air mixture section (as discussed in Section 4.5), the disposed and

ambient air enthalpy can be described by the following equation:

Equation 4-17..... (1.007 0.026) (2501 1.84 )h t t


51

Therefore, the enthalpy of point 1 and 2 are:

Equation 4-22…..…. 1 1 1 1(1.007 0.026) (2501 1.84 )h t t

Equation 4-23……... 2 2 2 2(1.007 0.026) (2501 1.84 )h t t

Neglecting the of heat loss in the mixture process, and using the principle of

conservation of energy, the enthalpy of different points in Figure 4-3 can be

described as:

1 3 2

3 2 1

h h m

h h m

Therefore, it can be written as

Equation 4-24…………… 1 1 2 23

1 2

m h m hh

m m

Here 1m and 2m are ambient air mass rate, and the air after dehumidifier mass flow

rate supplied to the air mixture section.

Results from the air mixture section are:

2m = 0.314 kg/s

Jones (2001) described air mixing states and air mixing humidity relationships.

According to mixing air theory, 1 2 3m m m for the dry air and 1 1 2 2 3 3m m m

for the associated water vapour, here 1 1,m is the air mixing mass flux and humidity

ratio, and 2 2,m is the air after dehumidifier mass flux and humidity. Therefore, 1m

can be written:


52

Equation 4-25 3 21

1 3

0.314m

Therefore, Equation 4-25 is changed to:

Equation 4-26 1 23 1 2

1 2 1 2

( ) ( )m m

h h hm m m m

Equation 4-27 3 1 22 1

1 2

1 1

1 1

h h hm m

m m

The state of points 1 and 2 can be calculated by the following equations:

Equation 4-28 1

2

m

m=

3 2

1 3 3 2 2

1 3 1

0.1340.008

0.134 0.008

Equation 4-29 2

1

m

m= 1 3 1

3 2 3 2 2

1 3

0.0080.134

0.0080.134( )

so that

Equation 4-30 2 13 1 2

1 2 1 2

0.008 0.008( ) ( )h h h

Applying Equation 4-22 and Equation 4-23 to Equation 4-30, and recognizing 1t , 1

and 2t , 2 as the temperature and humidity point 1 and 2, Equation 4-30 can be given

as:


53

Equation 4-31

2 13 1 1

1 2

1 22 2

1 2

0.0081.007 0.026 (2501 1.84 )

1000

0.008 1.007 0.026 (2501 1.84 )

1000

h t t

t t

4.7.2 TCOP Results and Discussion

Experimental data and theoretical modeling calculation results are also used for

TCOP analysis. The TCOP, cooling load ( 3 4h h ), and point 1 and point 2 humidity

relationship results are presented in Figure 4-8, Figure 4-9 and Figure 4-10.

4.7.2.1 TQ Thermal Energy Remains Constant

This research analyses the relationship between the TCOP, and cooling load when

the system use electricity to work as a stable thermal energy input. According to

Equation 4-4, TCOP 3 4 ( )c a

T T

Q m h h

Q Q

, system TCOP should be in direct

proportion to sensible cooling load cQ , where heating energy TQ is kept stable. This

research study has kept the heating energy constant at 4.0 kW in the TCOP analysis.

Figure 4-8 shows and predicts the trend of the TCOP and cooling load cQ

relationship. TCOP increases from 1.9 to 2.1 with the cooling load changing. From

the calculation results, the TCOP is almost in direct proportion to the cooling load

changing.

On the other hand, because cooling load equals enthalpy difference between point 3

and point 4 ( 3 4cQ h h ), and 4h is a constant at the inlet air state

( 44 4 41.007 0.026 (2501 1.84 )

1000h t t

, eg. 4 12t ºC, 4 0.008kg/kg), TCOP

is also in direct proportion to 3h . Also Equation 4-31 shows that 3h is determined by

1t , 2t and 1 , 2 . Therefore, as discussed, we can draw the conclusion that the liquid


54

desiccant hybrid cooling system (with air mixture AHU design) cooling load would

be governed by the ambient air situation ( 1t , 1 ), and the air after dehumidifier

situation ( 2t , 2 ) after the dehumidifier. Generally, the TCOP increases with sensible

cooling load increasing when the input energy maintains stability.

Figure 4-8 TCOP and Point 1, Point 2 Temperature Relationship

4.7.2.2 Relationship between TCOP and Point1, 2 states

According to the TCOP definition, TCOP 3 4 ( )c a

T T

Q m h h

Q Q

, and Equation 4-31,

TCOP is in direct proportion to 3h , and 3h depends upon the state of points 1 and 2

( 1t , 2t and 1 , 2 ), when the point 4 parameters maintains stability.

2 13 1 1

1 2

1 22 2

1 2

0.0081.007 0.026 (2501 1.84 )

1000

0.008 1.007 0.026 (2501 1.84 )

1000

h t t

t t

TCOP and State 1, Stae 2 Temperature Relationship

31.6

29.0

30.0

27.8

25.924.8

23.3

34.634.7

35.5

33.8 33.6 33.3

35.0

2.1

1.9

1.9

2.0

1.9

2.0

1.8

15.0

20.0

25.0

30.0

35.0

40.0

1 2 3 4 5 6 7

Test Point

Air

Tem

pera

ture

(°C

DB

)

1.7

1.8

1.8

1.9

1.9

2.0

2.0

2.1

2.1

TC

OP

Outdoor Air (State 1) Temp °CDB State 1 Temp °CDB TCOP


55

Figure 4-9 and Figure 4-10 show the TCOP, and points 1, 2 ( 1t , 2t and 1 , 2 )

relationship when TQ is kept stable. Here, the point 1 state ( 1t , 1 ) is the ambient air

parameters. The point 2 state is the air after dehumidifier parameters.

Figure 4-9 Hybrid Cooling System Cooling Load and TCOP Relationship

Figure 4-10 TCOP and Point 1, Point 2 Absolute Humidity Relationship

Hybrid Air Conditioning Cooling Load and TCOP Relationship

8.32

7.557.63 8.09 8.08

7.32

7.57

2.1

1.8

2.0

1.9

2.0

1.91.9

6.00

6.50

7.00

7.50

8.00

8.50

9.00

1 2 3 4 5 6 7

Test Point

Co

olin

g L

oad

(kW

)

1.7

1.8

1.8

1.9

1.9

2.0

2.0

2.1

2.11 2 3 4 5 6 7

TC

OP

Cooling Load TCOP

TCOP and State 1, Stae 2 Absolute Humidity Relationship

16.5

16.616.4

15.0

10.5

16.1

13.6

6.26.26.26.26.0

6.26.1

2.1

1.8

2.0

1.9

2.0

1.9

1.9

0.0

2.0

4.0

6.0

8.0

10.0

12.0

14.0

16.0

18.0

1 2 3 4 5 6 7

Ab

so

lute

Hu

mid

ity (

g/k

g)

1.7

1.8

1.8

1.9

1.9

2.0

2.0

2.1

2.1

TC

OP

State1 Absolute Humidity (g/kg) State2 Absolute Humidity (g/kg) TCOP


56

4.8 System Coefficient of Performance

Because the conventional air conditioning system does not include any other energy

input for the compressor, the system COP for the conventional air conditioning

system is the COPcon mentioned in Section 4.6 The hybrid liquid desiccant cooling

system is defined as the sensible cooling load, divided by the system‘s total power

consumption. The sensible cooling load per unit is the enthalpy difference 3 4h h

between the point 3 and point 4. The total energy cost includes: i) the conventional

cooling section energy cost, ii) thermal dehumidifier energy cost, and iii) other

energy costs. The conventional cooling system energy can be determined from the

Daikin engineering cooling system data. In this study, due to the thermal energy

originating from free solar energy, the system only calculates the thermal energy

when it used as electricity for heating. Other energy consumption includes the

electricity use of liquid pumps and air fans. All these energy consumption are

included in the COP calculations.

4.8.1 COPsys Definition

The coefficient of performance for the hybrid liquid desiccant cooling system,

including the air mixture design, is defined as:

Equation 4-32 sysCOP = c

T c other

Q

Q W W

3 4 ( )a

T c other

m h h

Q W W

Here cQ is the rate of heat removal from the sensible cooling part, and cQ per unit

equals the point 3 and point 4 enthalpy difference ( 3 4h h ) of the system. TQ is the

rate of the total thermal energy input during the dehumidification process. In this

study, the energy cost is derived only from electricity energy. cW is the cost rate of

the conventional compressor electricity used in the cooling process. otherW is the cost

rate of the other electricity energy in the hybrid system, including liquid pumps and

air fan parts.


57

Combining ECOP definition equation c

c

QECOP

W and Equation 4-32 gives the

change in the following equation:

Equation 4-33 csys

cT other

QCOP

QQ W

ECOP

So that,

Equation 4-34 1

1sysotherT

c c

COPWQ

Q ECOP Q

Then,

Equation 4-35 1

1 1sysother

c

COPW

TCOP ECOP Q

As otherW can be kept constant during the experiment and cQ equals cooling load data,

other

c

W

Qcan be calculated as a variable in Equation 4-35. This study shows the

relationship between the hybrid cooling system COPsys, and the conventional air

conditioning system COP.

4.8.2 COPsys and ECOP, TCOP Relationship

Many factors affect system performance during the dehumidification and sensible

cooling process. For instance, the heat transfers have been used as air dehumidifiers

in the liquid desiccant heat exchange section. The heat transfer efficiency can be

changed under different flow rate situations and can affect system performance. In

this research, these small factors are considered as constants during the COPsys

calculation (assumptions as shown in Section 4.4). The COPsys depends on TCOP,

ECOP, and other

c

W

Q which changes according to Equation 4-35. All the test data is

shown in Table 4-3.


58

Table 4-3 ECOP, TCOP and COPsys on the conventional air conditioning and hybrid liquid desiccant air conditioning*

T1 RH1 m1 QC (W/kW) QT (W/kW) ECOP TCOP COPsys

Outdoor Air Temp.

(°CDB)

Outdoor Air

Relative Humidity

(%)

Outdoor Mass

Flow Rate (kg/s)

Cooling Load (W/kW)

Thermal Energy Input (W/kW)

ECOP=Qc/Wc TCOP=Qc/QT COPsys=Qc/(QT+Wc+Wother)

CAC HAC CAC HAC

CAC HAC CAC HAC CAC HAC

Equal COP

Use Solar

Energy

Not Use Solar

Energy

Use Solar

Energy

Not Use Solar

Energy

Equal ECOP

Use Solar

Energy

Not Use Solar

Energy

23.3 92 0.0286 12307.5 7549.0 0.0 4000.0 5.0 6.7 6.7 - - 1.9 5.0 6.2 1.5

24.8 83 0.0282 12650.7 7626.9 0.0 4000.0 4.2 6.4 6.4 - - 1.9 4.2 5.9 1.5

25.9 71 0.0408 12255.2 8087.9 0.0 4000.0 4.0 5.6 5.6 - - 2.0 4.0 5.2 1.5

27.8 45 0.0698 9604.0 8079.8 0.0 4000.0 4.8 5.3 5.3 - - 2.0 4.8 5.0 1.5

29.0 54 0.0390 12061.6 7571.6 0.0 4000.0 3.1 5.8 5.8 - - 1.9 3.1 5.4 1.4

30.0 60 0.0275 14202.1 7319.8 0.0 4000.0 3.8 5.5 5.5 - - 1.8 3.8 5.1 1.4

31.6 56 0.0469 15975.8 8315.1 0.0 4000.0 2.5 4.8 4.8 - - 2.1 2.5 4.6 1.4

CAC=Conventional Vapour Compression Air Conditioning; HAC=Hybrid Liquid Desiccant Air Conditioning; QC=Total Cooling Load on the system (W/kW); QT=Total Thermal Energy Input (W/kW); ECOP=Qc/Wc; TCOP=Qc/QT; COPsys=Qc/(QT+Wc+Wother); Use Solar Energy: HAC use solar energy to work as thermal energy input *Outdoor air data is from Australia Commonwealth Bureau of Meteorology


59

In order to simplify the analysis, the other

c

W

Qcan be calculated as a variable during the

COPsys calculation. Therefore, Equation 4-35 can be changed to:

Equation 4-36 1

1 1sysCOP

eTCOP ECOP

Here e = other

c

W

Q, according to Table 4-1, otherW is considered as pump energy

consumption and other energy, totalling approximately 0.1kW. cQ changes from

7.55kW to 8.32 kW, therefore e changes from 0.0132 to 0.0120 according to otherW

and cQ . At the same time, the TCOP and ECOP are changing from 1.9 to 2.1 and 6.7

to 4.8 in this section. Therefore, the 1

TCOP equals 0.526 to 0.476, and

1

ECOP

changes from 0.149 to 0.208. The result of 1

TCOP+

1

ECOPis ranged from 0.675 to

0.684, and e is much smaller than 1

TCOP+

1

ECOP, only accounting for 1.95% to

1.75% of total. So, 1

TCOP+

1

ECOP influences COPsys significantly in relation to

the influences of e to COPsys.

Figure 4-11 shows that the changes of COPsys depend on ECOP and TCOP

variations. Due to solar energy being a free energy input, when the hybrid system

uses solar energy to regenerate liquid desiccant section, the whole system COPsys

becomes greater by changing value from 4.6 to 6.2. On the other hand, if the hybrid

system does not use solar energy and electricity is selected to work as inlet heating

energy, the whole system COPsys decreases to 1.4 to 1.5. This results because

electricity cost should be calculated as the energy input in liquid desiccant section.

In Figure 4-11, the hybrid cooling system COPsys (using solar energy) and ECOP

approximate parallel lines in the figure. At the bottom of Figure 4-11, a similar

characteristic can be viewed with the hybrid cooling system COPsys (using no solar


60

energy) and TCOP approximating a parallel relationship. The performance difference

between COPsys (using solar energy) and ECOP changes in values from 0.2 to 0.5.

At the same time, COPsys (not using solar energy) and the TCOP difference varies

from 0.4 to 0.7. This means that when the system use solar energy, thermal energy

input can be calculated as free Therefore the COPsys changes are mostly dependant

on ECOP changes. Otherwise, thermal energy input using electricity and TCOP

should be calculated into the system performance. COPsys changes are mostly

depended on the TCOP changes.

Figure 4-11 Hybrid Cooling System ECOP, TCOP and COPsys Relationship

As can be viewed in Figure 4-11, ECOP decreases from 6.7 to 4.8 (decrease of 28%).

At the same time, COPsys (using solar energy) decreases from 6.2 to 4.6 (decrease of

25%). TCOP shows some small fluctuations during the experiment from 1.9 to 2.0

and falls again to 1.8. After this fall, it rises up to 2.1 at the end. Total TCOP changes

approximate 10%. COPsys (not using solar energy) maintains stability with values

ranging from 1.4 to 1.5 (about 5%). Overall, system performance variation depends

on thermal performance, and electrical performance changes in the hybrid cooling

system.

Hybrid System ECOP, TCOP and COPsys Relationship

4.6

5.15.4

5.05.2

5.96.2

2.11.9 1.8

2.02.01.91.9

1.41.4 1.41.51.51.51.5

4.8

6.7 6.4

5.65.8

5.35.5

0.0

1.0

2.0

3.0

4.0

5.0

6.0

7.0

1 2 3 4 5 6 7

Test Points

EC

OP

an

d T

CO

P

0.0

1.0

2.0

3.0

4.0

5.0

6.0

7.01 2 3 4 5 6 7

CO

Psys

COPsys(use solar energy) TCOP COPsys(not use solar energy) ECOP

Hybrid cooling

system COPsys

without solar energy

Hybrid cooling

system COPsys with

solar energy


61

4.8.3 COPsys and Cooling Load Relationship

According to Equation 4-32, the COPsys definition is csys

T c other

QCOP

Q W W

.

When the system heating resource is selected as solar energy, it belongs to a free

resource with 0TQ . At the same time, otherW is a constant when the system fan and

other equipment (Fan, Pump energy as shown in Table 4-4) are in operation.

Table 4-4 Hybrid System Parameters Value/Range

Parameters Value/Range

Total load (kW) 40-60kW

Temperature of outdoor air (ºC) 20-40 ºC

Relative humidity of outdoor air (%) 20-85%

Temperature of indoor air (ºC) 25ºC

Relative humidity of outdoor air (%) 40%

Temperature of the air after dehumidifier (ºC) 20-65ºC

Relative humidity of the air after dehumidifier (%) 15-85%

Coefficient of air heat exchanger 0.5

Coefficient of liquid desiccant heat exchanger 0.5

Coefficient of hot water heat exchanger 0.5

Conventional refrigeration power of chillers(kW) 15kW

Power of liquid desiccant Pump 50W

Air Fan 50W

Therefore, the system performance can be defined as csys

c other

QCOP

W W

, where

otherW is a constant. COPsys variation is dependent on cQ and cW variation. Figure

4-12 presents a hybrid cooling system COPsys, and cooling load relationship. As

cooling load 3 4h h increases from 7.75 kW to 8.32 kW, COPsys decreases from 6.2

to 4.6. This means that under the same electrical energy for the fan and pumps, the

hybrid whole system performance will be reduced when the system disposes more

cooling load.


62

Figure 4-12 Hybrid Air Conditioning Cooling Load and COPsys Relationship

4.8.4 COPsys and Solar Energy Relationship

Because of the COPsys definition, csys

T c other

QCOP

Q W W

, when the system does

not use solar energy to regenerate liquid desiccant, 0TQ , COPsys becomes higher

when using solar energy. Figure 4-13 shows the trend that when the hybrid cooling

system uses solar energy, the hybrid system COPsys is much higher than COPsys

without solar energy usage. The difference between COPsys using solar energy and

without solar energy is 4.7 to 3.2. When the hybrid cooling system does not use solar

energy, the liquid desiccant section has to use electricity ( 0TQ ) to heat the liquid

desiccant for regeneration of liquid desiccant. This causes more energy to be used

compared to the hybrid cooling system using solar energy ( 0TQ ).

Hybrid Air Conditioning Cooling Load and COPsys Relationship

8.32

7.57

7.32

8.088.09

7.637.55

4.6

6.2

5.9

5.2

5.4

5.0

5.1

6.00

6.50

7.00

7.50

8.00

8.50

9.00

1 2 3 4 5 6 7

Co

olin

g L

oad

(kW

)

0.0

1.0

2.0

3.0

4.0

5.0

6.0

7.0

CO

Psys

Cooling Load (kW) Hybrid System COPsys (with solar energy)


63

Figure 4-13 Hybrid Cooling system COPsys with Solar Energy and without Solar

Energy


Performance measurement of the conventional air conditioning system, and the

hybrid liquid desiccant cooling system were reviewed in this chapter. The

conventional system was measured with COPcon , while the hybrid liquid desiccant

cooling system was measued with ECOP, TCOP and COPsys models. A detailed

study of these refrigeration models permitted the relationship and comparison

between the different systems. The performance of the system models provides

predictions based on fundamental equations, minimizing the assumptions and the use

of empirical correlations. AHU air mixture and enthalpy calculation used in the

performance were also presented.

Although the finite difference model describing the performance of the regeneration

has been reported in some literature, simpler algebraic equations correlating the

Hybrid Cooling System COPsys with Solar Energy and without Solar

Energy

4.6

1.4 1.41.5

1.51.51.5 1.4

6.2

5.95.2

5.4

5.0

5.1

0.0

1.0

2.0

3.0

4.0

5.0

6.0

7.0

1 2 3 4 5 6 7Test Point

CO

P

Hybrid cooling system COPsys (not use solar energy)

Hybrid cooling system COPsys (use solar energy)

Hybrid cooling

system COPsys

without solar energy

Hybrid cooling

system COPsys with

solar energy


64

performance to design variables would be more convenient in desiccant cooling

system performance simulations. COPcon, ECOP, TCOP and COPsys correlations can

be drawn from the experimental data. Furthermore, since experimental conditions are

complex in desiccant regeneration and dehumidification processes, system

performance calculations are included but still rely on certain assumptions.

Based on the equations for the system performance simulations, we can draw the

following conclusions about the correlations: i) The COPcon varies in direct

proportion to cooling load with the conventional air conditioning system. ECOP also

approximates variation in direct proportion to the cooling load in the hybrid cooling

system. ii) In the TCOP section, TCOP varies is direct proportional with an

increasing cooling load. When the thermal energy consumption TQ is assumed to be

constant, TCOP variation were dependant on the humidity and temperature of the

ambient air. iii) When the hybrid cooling system uses solar energy, COPsys is

dependent on ECOP variation. When the system does not use solar energy, COPsys

variation was shown to change in direct proportion with TCOP variation. ECOP and

TCOP changes can affect the whole hybrid system COP under different situations.

iv). According to the mixture air models, the point 2 air parameters (enthalpy,

temperature and humidity) can be calculated. The point 2 air parameters are used in

the COP calculation. All the different data is summarized in the APPENDIX in this

thesis.

According to correlation analysis between the different COP‘s, the variables found to

have the greatest impact on the performance of the hybrid cooling system with air

mixture design were: i) the cooling load, ii) the ambient air temperature and iii) the

humidity. The air after dehumidifier temperature and humidity can be determined

from the total cooling load. The ambient air from the inlet fan and the air after

dehumidifier from the dehumidifier did have a significant effect on the cooling load,

and therefore affected the system performance. In general, the COPsys level from the

present study depends on the cooling load, ECOP and TCOP situation.

Also, since the correlation considers the change of ambient air, the ambient air

humidity and ambient air temperature are available in the literature. For these reasons,


65

two performance correlations were developed by analyzing the experimental data

from the present study. One correlation was for the ambient air parameters, while the

other correlation concerned the ECOP, TCOP and COPsys parameters. These

correlations give the performance within the experimental data. In addition, the

correlations gave excellent predictions of the influence of design variables, both for

air dehumidification and desiccant regeneration.

The overall COP of the hybrid liquid desiccant cooling systems predicted by

calculation ranges from 6.2 to 4.6 when system uses solar energy, or 1.5 to 1.4

without use solar energy. A total average approximating a 69% to 78% difference in

performance was found when the hybrid system using solar energy to dispose latent

load was compared to the hybrid cooling system without solar energy. Calculations

have shown a much better savings in coupling a desiccant dehumidifier with solar

energy using. The conventional cooling system, using a vapor compression system to

deal with sensible and latent heat load both had a performance COPcon variation

ranging from 5.0 to 2.5.


66

5 REGENERATION AND DEHUMIDIFICATION TEST, RESULTS AND DISCUSSION

5.1 Introduction

The hybrid system is divided into three main sections: dehumidification,

conventional air conditioning and the air mixture terminal (VAV). For the

dehumidification and air mixture terminal section, analysis can be performed based

on air conditioning theory. Experimental studies in this thesis include the

dehumidifier/regenerator test, with analysis and discussion of test results being

included in this chapter. The experimental system uses LiCl liquid desiccant as a

dehumidification medium according to the design requirement and properties of LiCl

are presented in this chapter. The main experimental equipment concerns the

regenerator/dehumidifier, with testing including the regeneration and

dehumidification sections. The following subsections explain how these

experimental components were selected and arranged, to obtain test data for the

hybrid liquid desiccant system.

5.2 Experimental Configuration

The dehumidifier/regenerator is the apparatus where a strong or weak solution is

sprayed over a packed bed column and water is absorbed, or evaporated from the

desiccant solution respectively. The dehumidifier/regenerator uses solar energy

working as a heat source to recycle the desiccant solution, which is shown in Figure

5-1. The liquid desiccant concentration has been changed during the cycled between

the dehumidifier and regenerator equipment. Based on system design theory, the

experiment is divided into two parts: dehumidification and regeneration. The

dehumidification and regeneration processes are similar but the air after dehumidifier

is totally different for each process. The air after the dehumidifier and desiccant flow

direction are shown in Figure 5-1.


67

Dehumidifier

Ambient Air

Inlet

Disposed

Air Outlet

Regenerator

Outlet Air Liquid

Desiccant

Circuit

Heater

Hot water

Hot

Resource

Liquid

Desiccant

Pump

Heater

Desiccant

Tank

Ambient

Air Inlet

Figure 5-1 Schematic Diagram of the Regeneration and Dehumidification Systems

Due to experimental conditions reason, the whole experiment has been divided into

two steps. In the first step, the dehumidification test uses high concentration liquid

desiccant to do the dehumidification process. The entire desiccant is only recycled by

a pump between the dehumidifier and the desiccant tank. In the second step, the

regenerator uses low concentration liquid desiccant after the dehumidification

process. The heating component uses a normal electrical heater to simulate the use of

solar energy to heat water. Hot water is used to heat ambient air to achieve desiccant

regeneration. The dehumidifier and regenerator experiments are based on the

schematic diagram shown in Figure 5-2.

The title of experimental parameters used in Figure 5-2 is provided below:

Air flow flux ( m ),

Liquid desiccant solution concentration ( )

Temperature (T )

Humidity ( )

Enthalpy ( h )


68

HOT AIR

Dehumidification

CONTROL VOLUME

Desiccant Inlet

Regeneration

CONTROL VOLUME

Desiccant Outlet

Environmental Air Dry Air

Desiccant Inlet

Humid Air

+ a a a a a a am d h dh T dT

HIGH CONCENTRATION

PHASE

d d dm h T

d d dm h T

LOW CONCENTRATION

PHASE

HIGH CONCENTRATION

PHASE

HIGH HUMIDITY PHASELOW HUMIDITY PHASE

LOW HUMIDITY PHASEHIGH HUMIDITY PHASE

a a a am h T + a a a a a a am d h dh T dT

' ' ' 'a a a am h T' ' ' ' ' ' 'a a a a a a am d h dh T dT

+

Environmental Air

Indoor

'' a a a am h T

i i i im h T

Figure 5-2 Air and Desiccant Parameters in the Hybrid Cooling System

Cooling andDehumidifying

Heating andDehumidifying

A

Figure 5-3 Dehumidifying, Heating and Cooling Process on the Psychrometric Chart

Figure 5-3 provides a graphical illustration of the air conditioning dehumidification

process, together with heating and cooling. In actual practice, both the dry-bulb

temperature and moisture content of the air generally change simultaneously. When


69

this occurs, the resulting air sample moves from point A (Figure 5-3) to another

enthalpy state. The exact angle and direction depend upon the proportions of sensible

and latent heat added or removed. Sensible heat causes a change in the air‘s dry-bulb

temperature with no change in moisture content. Latent heat causes a change in the

air‘s moisture content with no change in dry-bulb temperature. The angle of the

different air states are discussed in this chapter during dehumidification and

regeneration testing.

5.3 Experimental Components

5.3.1 Selection of Desiccant

According to the literature review, desiccants are broadly classified as solid and

liquid, and have the property of extracting and retaining moisture from air brought

into contact with them. By using either type, the moisture in the air is removed and

the resulting dry air can be used for air conditioning or drying purposes. Since the

required regeneration temperatures are low, solar energy can be successfully used for

regeneration of liquid desiccants (Ertas, Anderson and Kiris, 1992). In this research,

the experimental section was chosen to apply liquid desiccants for dehumidification.

In these experiments, the system is designed using a single pump to move the liquid

desiccant; therefore the experiment needs high dehumidification ability and a high

solubility desiccant to work. According to the required properties, lithium chloride is

the most stable liquid desiccant and has a large dehydration concentration (30% to

45%) (Ertas, Anderson and Kiris, 1992). Thus, lithium chloride has been selected as

the liquid desiccant for experimentation.

Figure 5-4 gives the solubility boundaries of lithium chloride salt solutions that are

defined by several lines at different temperatures and in different concentrations. The

solubility line defines the conditions at which crystals start to form. For higher

concentrations, the solubility boundary defines the conditions at which salt hydrates

or anhydrous salt crystallize from the solution. This is the crystallization line (Conde,

2004). In the following experiment, the concentration is kept below 50% (by weight)

to ensure no crystallization occurs.


70

Figure 5-4 Solubility boundary of aqueous solutions of lithium chloride (Conde,

2004)

5.3.2 Regenerator/Dehumidifier Description

The regenerator/dehumidifier test equipment is shown in Figure 5-5. This equipment

consists of a packed bed, an air fan, a cycle pump, and a tank containing liquid

desiccant. The packed material is the type generally used for air contact equipment.

The system needs packed material because the packed material offers a large surface

contact area for the heat and mass transfer between the liquid desiccant and air.

Plastic Pall Ring has been used in this research. Plastic Pall Ring has an advance on

the Raschig Ring, not only does it have a similar cylindrical dimensions, but it also

has two rows of punched out holes. It also has fingers or webs turned into the centre

of the cylinder, which significantly increases the performance of the packing in terms

of throughput, efficiency and pressure drop. As Figure 5-5 shows, there is a net

frame and sponger filter used within the inlet to remove dust from the air. One


71

desiccant pump is used to cycle desiccant between the container and sprayer section.

Hot water is obtained from a solar panel or electrical heater and water is pumped into

the heat exchanger to heat the air. The heat comes from the hot water transfer into

hot air to regenerate the liquid desiccant. The desiccant tank is used as a container for

the liquid desiccant. A 40L liquid desiccant tank has been used in this experiment.

Figure 5-5 Regeneration and Dehumidification Test System

5.3.3 Test Equipment Description

There are several different types of test equipment used in the experiment as shown

in Table 5-1. Under the operating conditions, data is taken under steady-state

conditions. Measurements for the regeneration/dehumidification process include inlet

and outlet DBTs (Dry Bulb Temperatures) using Multipoint Recording

Thermometers ‗K‘type thermocouples (±0.1°C), relative humidity using RH sensors

(±0.5% RH) and air flow rates using Anemometer Pocket Air speed Temperature

Meter (±0.02 m/s). The liquid desiccant flow rate measurements are obtained by

using a Liquid Flow Meter ( 0.005kg/s) and liquid desiccant solution concentration

was measured by a Waterproof Hand-held Conductivity/TDS-PH/mv-Temperature

Air Outlet

Air Inlet

and Fan

Heat

Exchanger

Desiccant

Tank

Cycle

Pump

Air Flow

Direction

Packed

Bed

Materials


72

Meter at the inlet and outlet of the regenerator, dehumidifier and container. The

power consumed by the fan is determined using a Speed Controller (0-3 Amp.). All

calibration of the measuring equipments had been done before testing. Changing the

resistance of the electric power can vary the speed of the fan. The variable resistance

drive indicates the set frequency and the corresponding speed of the fan. Similarly,

another variable resistance drive is used to vary the speed of the liquid desiccant to

achieve varying liquid desiccant flow rates. Tests were conducted to determine the

influence of three design parameters, namely airflow rate, air humidity and air

temperature on the key performance parameters.

Table 5-1 Test Equipments List

Equipments Range/Error

Waterproof Hand-held Conductivity/TDS-PH/mv-Temperature

Meter ±0.01 s

Multipoint Recording Thermometers/Thermocouple K Model ±0.1 ºC

Flow Meter ±0.005kg/s

Manometer 0-2 bar

Scale ±0.1g

Measuring Cylinder (200,100, 50 ml), ±0.5ml

Digital Thermometer (-50 ºC -150 ºC)±0.5ºC

Speed Controller 0-5A

5.4 Regeneration Test

5.4.1 Introduction

The experimental setup for studying the performance evaluation of the regenerator is

shown in Figure 5-6. The regenerator was fabricated with packed bed materials

inside. The aim of the regeneration process is to transfer the ‗diluted desiccant‘ into

‗concentrated desiccant‘. LiCl solution was used as the absorbent solution and it was

distributed at the top end of the regenerator through three metal sprayers to spray on

to the packed bed material. The absorbent solution flows over the absorber as a thin


73

film. Liquid desiccant is sprayed on the packed bed structure, and hot air flows in the

opposite direction to contact the desiccant and achieve regeneration. A plastic liquid

desiccant pump (0-5L/s) was used to circulate the desiccant solution from the

solution tank to the sprayers. A plastic flowmeter ( 0.005kg/s) was attached to the

delivery side of the pump in order to record the flow rate of the desiccant over the

packed bed materials. This setup also was used for the dehumidification test.

Regenerator

Inlet Air

Outlet Air

Liquid

Desiccant

Inlet

Heater

Hot water

Hot water

Liquid

Desiccant

Outlet

Heater

Figure 5-6 Schematic Diagram of Regeneration Test

5.4.2 Test Parameters

The regenerator design is a half open system and the desiccant contact system is

adiabatic. The water vapour pressure of the desiccant solution exceeds the water

vapour pressure of the atmospheric air, and mass transfer takes place from the

desiccant to the atmospheric air. The solution leaving the regenerator becomes more

concentrated as a result of water evaporating from it in the regenerator (Ahmed,

1996). Therefore, testing the concentration of the desiccant solution and the air after

dehumidifier humidity difference are two primary parameters in the experiment. The

liquid desiccant concentration changes in regeneration R and the air after

dehumidifier humidity changes in regeneration R , which are defined as:


74

Equation 5-1 R out in

Equation 5-2 R out in

In these equations, in and out are the concentrations of inlet and outlet desiccant

solution and in andout are the air humidity of the inlet and outlet regenerator.

Other parameters are required for finding the regeneration heat and mass transfer

effectiveness. These parameters that are necessary for the regeneration test include:

Air flow rate L/s

Liquid desiccant solution concentration %(by weight)

Inlet and outlet air humidity kg/kg

Inlet and outlet air temperature C

Desiccant flow rate L/minute

The liquid desiccant solution concentration, time, inlet and outlet desiccant

parameters were recorded during the test. The desiccant flow system is continuous;

hence the desiccant concentration will increase and all changes were recorded during

the regeneration. Numerical analysis of regeneration is provided in the following

sections.

5.4.3 Regeneration Results and Discussion

5.4.3.1 Liquid Desiccant Concentration

Dehumidification or regeneration is determined by the difference in water vapour

pressure between the desiccant solution and the air after dehumidifier. Because the

concentrated liquid desiccant solution has a water vapour pressure lower than the

water vapour pressure of atmospheric air at ambient temperature, the moisture from

the ambient air is absorbed into the liquid desiccant. This becomes a

dehumidification process if there is no hot water pumped into the system. Figure 5-7

shows the existence of a single critical point between the regeneration and

dehumidification processes within the regeneration experiment. In Figure 5-7,

out in is defined. In Figure 5-7 (a), the critical point is shown when R =0.


75

This means that the outlet desiccant concentration equals inlet desiccant

concentration ( out in ). The concentration of the liquid desiccant does not change

when it arrives at the critical point. Before the critical point R >0, and after the

critical point R <0. There are three data points shown in Figure 5-7 (b), for the

regeneration experiment conducted during 14-15 June, 2006. The initial desiccant

conductivity value was 36.8 mS, equivalent to a concentration of 40.2% in the 14

June regeneration test. Air flow rate is 40L/s and average desiccant flow rate is

0.3L/min. After that, the experiment stopped the hot water pump for several hours.

On 15 June at 8:30am, the desiccant conductivity was at the maximum point of

40.2mS, 42.2%. The maximum test point desiccant conductivity was 40.2ms, and

concentration was 42.2%. On 15 June at 11:20am the desiccant measured 31.2 mS,

indicating a dilution of 31.8%. This is lower than the initial data because ambient air

moisture diluted the liquid desiccant when the air water pump stopped.

Figure 5-7 Schematic Diagram of Liquid Desiccant Solution Concentration in the

Regeneration Test


76

5.4.3.2 Regeneration With the Hot Water Pump

Figure 5-8 depicts the temperature and humidity situations in the regeneration test

with the hot water pump open. Inlet humidity and inlet temperature are quite stable

from the beginning to the end of this test. Inlet temperature remains approximately

17-18 C, and inlet absolute humidity stays at 7g/kg. The temperature of outlet air

from the regenerator decreases steadily from 44.8 C to 38.4 C, and outlet air

humidity decreases quickly from 35.4g/kg to 17.7g/kg. Outlet temperature and

humidity are all higher than at the inlet, meaning that moisture evaporates from the

liquid desiccant into the ambient air that is being heated. The hot water heats the

liquid desiccant and inlet air. This causes the liquid desiccant and outlet air

temperature to increase. The moisture evaporating from the desiccant process is

referred to as regeneration. The outlet air humidity decreases quickly in comparison

to the temperature decrease. This arises because the desiccant concentration increases

and less water is left in the desiccant. The evaporation speed becomes slower with

increasing concentration.

Figure 5-8 Regeneration Test with Hot Water Pump


77

......

h2-h1>044.8°C 57.3% 35.4g/kg -38.4°C 41.1% 17.7g/kg

2

1

18°C 56% 7.2g/kg - 17°C 59% 7.1g/kg

+++++ +

Figure 5-9 Regeneration Experimental Results Trend Analysis in Psychrometic

chart

Figure 5-9 is psychrometric chart showing experimental results for several

regeneration test points. Point 1 is the inlet air situation, while the outlet air after

regeneration is highlighted as points upon the lines identified as line 2. Because the

desiccant in the regenerator has been heated by solar energy to a higher temperature,

the output air sample also has a higher temperature compared to the air state at point

1. The moisture has been heated out during the regeneration process and the

desiccant has been concentrated.

As shown in Figure 5-9, the enthalpy of the air has changed from a state point of

approximately 36 kJ/kg (point 1), to values between 85 – 120 kJ/kg (line 2). This

represents a very sharp increase in enthalpy during the process. The enthalpy

difference (i.e. - ) after this regeneration process is also greater than zero. This

means that the latent heat in the desiccant has been moved as moisture into the air

sample. If the system stops heating the desiccant, the disposed process will become a

dehumidification process again. The following section discusses regeneration

without the hot water pump running. The critical point will therefore be used during

such a process.


78

5.4.3.3 Regeneration Without the Hot Water Pump

The regeneration test without the hot water pump is shown in Figure 5-10. With

similar desiccant solution concentration, humidity change also has a relative critical

point ( 0R out in ) between regeneration and dehumidification. Comparing

inlet air situations, the outlet air temperature and humidity decrease quickly, but

before the critical point (about absolute humidity 7.8g/kg), outlet air humidity is

greater than inlet air humidity R >0. This process therefore still belongs to

regeneration. After the critical point, outlet absolute humidity becomes lower than

the inlet absolute humidity ( R <0). The system changes back into a

dehumidification process. The moisture is absorbed into the desiccant from ambient

air. According to Figure 5-10, the regeneration process continues with the hot water

pump being stopped. Before the critical point, the system still belongs to

regeneration, and inlet humidity is lower than outlet humidity. After the critical point,

the system becomes a dehumidifier, and inlet humidity is greater than outlet humidity.

The system changes from regeneration into dehumidification with a delay of about 4

hours (16:40 to 20:50) after the hot water pump is stopped. As can be see from this

figure, the packed bed liquid desiccant regeneration system can be postponed quite a

long time, even with no heat supplement.

Figure 5-10 Regeneration Test without Hot Water Pump


79

5.4.3.4 Humidity Analysis in the Regeneration Test with the Hot Water Pump

Figure 5-11 shows that the inlet air humidity, outlet humidity, and the humidity

difference between inlet and outlet ( humidity) varying tendencies in the

regeneration test with the hot water pump. From 15:50 to 16:40, the inlet air

humidity was stable around 7.0 to 7.3 g/kg. However, the humidity of the outlet

rapidly decreased from 35.4 to 17.7g/kg. The humidity also displayed a similar

trend to decrease quickly from 28.2g/kg to 10.6g/kg. From the humidity shown, it

can be seen that R (moisture evaporate from liquid desiccant) decreases quickly.

This means that the evaporation moisture speed decreases from beginning to the end

of the regeneration test. More moisture can be evaporated from liquid desiccant at

the beginning, and the evaporation speed decreases with increasing liquid desiccant

solution concentration.

Figure 5-11 Humidity Situation in Regeneration Test with Hot Water Pump


80

5.4.3.5 Regeneration Test Results Discussion

Figure 5-12 illustrates all humidity points in the regeneration system with, and

without the hot water pump. It is obvious from this graph that there is a decreasing

outlet humidity trend from the test beginning to the end. There is a critical point

between regeneration and dehumidification. The experimental process basically

converts from regeneration to dehumidification after the water pump is stopped after

approximately 3.5 hours. Moisture is evaporated from the desiccant into air. Hence

the liquid desiccant concentration is at a maximum when the test arrives at a critical

point. After the critical point, the experimental process converts from regeneration

into dehumidification, and outlet air humidity becomes lower than inlet air humidity.

The liquid desiccant is diluted by incoming moisture, and desiccant concentration

decreases after the critical point.

In summary, it is clear that regeneration continues for 3.5 hours after the hot water

pump is stopped. This phenomenon exhibited by the system is hysteresis and the lag

in response is quite substantial. The other phenomenon shown in the system is a

steady decrease in the outlet air humidity and the changing air humidity ( R ) data.

Even when the inlet air humidity is kept stable, the air humidity difference R still

keeps decreasing from the beginning to the end. The decreasing speed becomes faster

when the hot water pump is running.


81

Humidity Situations in Regeneration Test

7.8

-6

-3

0

3

6

9

12

15

18

21

24

27

30

33

36

39

15:50 16:10 16:30 16:50 17:10 17:30 17:50 18:10 18:30 18:50 19:10 19:30 19:50 20:10 20:30 20:50 21:10 21:30 21:50

Time (10 mins)

Ab

so

lute

Hu

mid

ity

(g

/kg

)

Inlet Air Absolute Humidity Outlet Air Absolute Humidity ΔHumidity

Dehumidification

and Regeneration

Critical Point

DehumidificationRegeneration

Without Hot Water PumpWith Hot Water Pump

Figure 5-12 Humidity Situation in Regeneration Test


82

5.5 Dehumidification Test

5.5.1 Introduction

The experimental setup for studying the performance evaluation of a dehumidifier is

shown in Figure 5-13. The dehumidification process is similar to the regeneration

process. The aim of the dehumidification process is to change the ‗moist air‘ into ‗dry

air‘. Similar to the regeneration test, testing the changing humidity of the air after

dehumidifier is the primary objective of this experiment. LiCl solution was used as

the absorbent solution to dehumidify and it was distributed at the top end of the

dehumidifier through three metal sprayers set to spray on the packed bed material.

Liquid desiccant is sprayed on the packed bed structure, and ambient air runs in the

opposite direction to contact desiccant to achieve dehumidification.

Dehumidifier

Inlet Air

Disposed

Air

Liquid

Desiccant

Inlet

Liquid

Desiccant

Outlet

Figure 5-13 Schematic Diagram of Dehumidification Test

The key dehumidification process parameters are similar to the regeneration process

in the following analysis. The humidification process can be accomplished in

equipment such as a finned-tube surface with column spray tower or packed tower

(Ahmed, 1996). In this dehumidification test, raschig rings are used as the packed

material for the simulation study. Because it is a dehumidification test, the testing

concentration of desiccant solution and the air after dehumidifier humidity difference

are still two primary parameters of this experiment. The other important parameter for

systems perform analysis is temperature difference DT between inlet and outlet air.

The outlet temperature affects cooling load and performance in the conventional


83

cooling section. The changing liquid desiccant concentration D , changing the air

after dehumidifier humidity D and temperature difference DT , which are defined

as:

Equation 5-3 D out in

Equation 5-4 D out in

Equation 5-5 D out inT T T

In these equations, in and out represent the concentrations of the inlet and outlet

desiccant solutions, and in and out are the air humidity at the inlet and outlet of the

dehumidifier. In the following experiment, five parameters will be tested and

collected during the dehumidification test:

Air flow rate L/s

Liquid desiccant solution concentration %(by weight)

Desiccant flow rate L/min

Inlet and outlet air RH%

Inlet and outlet air temperature C

Similar to the regeneration test, all the parameters that are mentioned above are

recorded from the beginning to the end of test. Numerical analysis concerning

dehumidification is provided in the following section.

5.5.2 Dehumidification Results and Discussions

5.5.2.1 Liquid Desiccant Concentration

The liquid desiccant is circulated between the desiccant tank and dehumidifier.

Therefore, the desiccant concentration decreases from the beginning to the end of the


84

dehumidification test. The water vapour pressure of the strong liquid desiccant

solution is always lower than the water vapour pressure of the atmospheric air under

an ambient air temperature situation. Thus, the moisture is absorbed from ambient air

into desiccant. Figure 5-14 shows that there is no critical point such as in the

regeneration test. The outlet solution concentration is always lower than inlet

concentration. Therefore, the difference of concentration D ( out in ) is always

maintained below zero. Several data points are shown in the Figure 5-14 for the

dehumidification experiment performed on 29 April, 2006. The initial desiccant

conductivity value was 47.6 mS, equivalent to a concentration of 36.06% (by weight)

at 9:05am. During the dehumidification experiment, the conductivity value decreased

slightly from 35.88% to 35.50%. At the end of the experiment, the liquid desiccant

conductivity value dropped to 45.1 mS, equivalent to a concentration of 34.49% at

11:25am. In the 2.5 hour dehumidification experiment, the liquid desiccant

concentration decreased approximately 2%. The ambient air moisture diluted liquid

desiccant during the experiment and this caused the desiccant concentration to

decrease.

T

ξ

0

Dehumidification

.

36.06%

34.49%

..35.50%

35.88%.

Figure 5-14 Schematic Diagram of Liquid Desiccant Solution Concentration in

Dehumidification Experiment

5.5.2.2 Humidity Results Analysis

Figure 5-15 depicts the inlet humidity, outlet humidity and humidity difference ( D )

in the dehumidification test. Inlet air humidity changed from 8g/kg to 10g/kg. The

humidity change depends on the ambient air conditions. Outlet humidity began


85

around 4g/kg, and finished approximately 6g/kg. In the whole process, the outlet

humidity data is quite stable, mostly around about 5-6g/kg, compared to the variation

of ambient air humidity. The humidity difference ( D inlet outlet ) declined

quickly from the outset at 12:55pm. The decreasing trend continued for 2 hours until

2:45pm. After this point, the humidity difference became stable until the end of the

experiment.

Figure 5-15 Humidity Results Analysis in Dehumidification Experiment

It must be noted that the air humidity after dehumidifier is quite stable even when the

ambient air humidity changes quickly. The use of liquid desiccant and a packed bed

structure is one of the main reasons that the outlet humidity has a low dependence on

inlet humidity. The graph also shows the humidity difference decreased with

experimental time. This is because the liquid desiccant concentration decreased. The

inlet air humidity changed depending on the ambient air humidity situations. The inlet

air is the air state 1 in the hybrid system before the air mixing process. Similarly,

Outlet air from the dehumidifier is at air state 2 in the hybrid system.

All the air temperatures in dehumidification are shown in Figure 5-16. There is one

unstable state period at the beginning of the experiment. This period continues for

approximately 30 minutes. Outlet air temperature increased from 24 C to 35 C. The


86

air temperature difference ( DT ) increased quickly from 3 C to 13 C in the unstable

period. Outlet air temperature being the dehumidified air temperature, increased

quickly because of the liquid desiccants high temperature. The desiccant temperature

is increased because of the chemical reaction during the desiccant dilution with a

moisture chemical process. During the stable state period, the outlet air temperature

has similar humidity, remaining constant from 32 C to 34 C. The air temperature

difference ( DT ) shows that there have been some fluctuations in the stable state from

13:35pm to 16:15pm. The temperature change is from 8 C to 13 C. Figure 5-16 also

represents the outlet air temperature (the air after dehumidifier temperature)

stabilising when the experiment becomes stable; even the ambient air changes quickly.

Hence, it causes air temperature ( DT ) fluctuation.

Figure 5-16 Results Trend Analysis in Dehumidification Experiment

A comparison between the inlet and outlet air temperature and humidity trends from

the experiment is provided in Figure 5-16. This chart shows the whole trend of air

temperature and humidity from the beginning to the end of the experiment. In general,

the inlet air humidity is higher than outlet air humidity, and outlet air temperature is

higher than inlet air temperature. Comparing inlet air humidity and temperature, the


87

outlet air temperature and humidity are more stable. Thirty minutes after the

beginning, the outlet air humidity and temperature only changed over a very small

range until the conclusion of the experiment. One reason for the outlet air (the air after

dehumidifier) parameter‘s stability is that the liquid desiccant concentration did not

influence the humidity and temperature significantly when the experiment reached the

stable state.

.......

h1-h2<0h1=h2

h1-h2>0

24°C 58% 10.8g/kg - 27°C 43% 10g/kg

2

1

20°C 22% 3.2g/kg - 32°C 24% 7g/kg

+++ ++ ++ +

Time

Figure 5-17 Dehumidification Experimental Results Trend Analysis in Psychrometic

chart

Dehumidification test points are shown in the psychrometric chart of Figure 5-17.

This chart can be used to predict output air states of a dehumidification experiment.

This is highlighted with line 2, representing seven (7) output states. Each state point

highlights the enthalpy state of the air sample in hourly increments. The inlet air

samples are highlight at point 1 within the psychrometric chart of Figure 5-17. Line 2

represents the air samples after dehumidification. Outlet air temperature has increased

with time because of the desiccant dehumidification chemical process. From point 1

to line 2, the absolute humidity has also dropped after the dehumidification process.

The enthalpy points from point 1 to line 2 are also divided into three sections being,

1) , 2) and 3) . From Figure 5-17 is can be viewed

that the enthalpy state ( ) at point 1 has maintained a lower value of approximately

45 kJ/kg to 50kJ/kg. Conversely, the enthalpy states represented by line 2 have also

increased. This results because the air samples have continuously experienced a


88

greater period of time of passing the hot liquid desiccant. This therefore means that

the air conditioning needs more energy to dispose sensible heat, even though latent

heat is decreasing.

The results of the dehumidification experiment are shown in Figure 5-17. Analysis of

this psychrometric chart has shown that the experiment tends to support the theory

provided in Chapter 4 for the hybrid system displayed in Figure 4-3. Figure 5-17 has

therefore established a method where the outputs enthalpy states can be predicated

based on knowledge of process slope being known.


This chapter presents the results of a detailed study of the humidity and temperature

of inlet and outlet air in a packed bed dehumidifier/regenerator. The results from this

study provide valuable insight into the characteristics of packed bed dehumidifiers

and desiccant regenerators. A comparison between the dehumidification and

regeneration processes in this study shows that the liquid desiccant system is not

flexible. In the regeneration test, the system displays hysteresis and a significant

regeneration lag response (4 hours). In the dehumidification test, the outlet (disposed)

air is quite stable, and did not change much with change in the inlet (ambient) air

situation. On the other hand, there is one critical point in the regeneration test, and

regeneration can change into a dehumidification state if the system stops supplying

heat. The dehumidification test exists in an unstable state during the first 30 minutes.

The inlet air variation does not significantly influence the dehumidification and

regeneration performance. Therefore, the system should be appended to the AHU

system to increase system flexibility. Thus, a detailed study of the air mixture and

ambient air weather data should be carried out, to evaluate their effect on system

performance. In the following chapter, air mixture and weather data will be analysed

using the fundamental equations, empirical correlations and assumptions outlined in

Chapter 3.


89

6 AIR MIXTURE AND AMBIENT AIR DATA ANALYSIS

6.1 Introduction

The experimental data and analysis provided in Chapter 5 showed that the packed bed

liquid desiccant dehumidification/regeneration system has a lag response. According

to air mixture design, accurate humidity control can be achieved by mixing some

ambient air into the system. The packed bed liquid desiccant system has a flexibility

problem, and this problem can be solved by air mixture design. At the same time, the

ambient air humidity and temperature were subject to rapid variation. In this chapter,

the air mixture rate calculation and ambient air data analysis are presented in detail. In

this investigation, the entire ambient air data is from the Australian Bureau of

Meteorology (Brisbane Weather Data). An evaluation of the experimental data from

this investigation, combined with the weather data permits derivation of performance

correlations. These correlations will simplify performance estimates and the design of

liquid desiccant systems. Analysis of the ambient air latent load and sensible load is

also presented in this chapter.

6.2 Air Mixture Rate Calculation Analysis

According to the theory outlined in Chapter 4, the process of mixing air is short. In

the following air mixture study, we suppose that two air streams mix adiabatically to

satisfy supply air humidity requirements. Discussion of the air mixing process is

based on Figure 6-1. In this example, the assumed requirement of air is that is has an

absolute humidity of 0.007kg/kg, and temperature of 27 C. This air state equals

relative humidity of 30%, at 27 C. It provides a comfortable environment. During

discussion of the mixing process in this chapter, it should be noted that the same

parameters used in Chapter 4 (such as air flow rate) are used in this section. The data

used is also acquired from the experimental results.


90

Figure 6-1 Mixture Air Process in Psychrometric Chart

According to air conditioning theory and experimental results, several equations in the

air mixing process can be used in the following study. The air mixture process is

schematically illustrated in Figure 6-2.

Figure 6-2 Schematic Diagram of Air Mixture Points

Because the fan system is stable, the air flow rate after dehumidifier used in chapter 4

can be applied here. This is shown in Equation 6-1:

Equation 6-1

. 0.269air inS m3/s


91

The air density is also the same as selected in Chapter 4, which is shown in Equation

6-2

Equation 6-2

air = 1.168 kg/m3

Applying the parameters given above, the following result is shown in Equation 6-3.

Equation 6-3

Bm = 0.314 kg/s

According to equation (4.7) in Chapter 4, we can write:

Equation 6-4 C B

A B

A C

m m

The assumed humidity of point C (in Figure 6-1) is 0.007kg/kg, which is the assumed

requirement and is shown in Equation 6-5.

Equation 6-5

C = 0.007kg/kg

Therefore, the mass flow of the ambient air mix can be calculated as:

Equation 6-6 0.007

0.3140.007

BA

A

m

Figure 6-3 indicates the results of calculation according to Equation 6-6. All

humidity data is based on experimental results. Figure 6-3 shows the theoretical

results of adding the mixed air stream. Ambient air and the air humidity after

dehumidifier both change during the experiment, so the air mixture mass line also

changes transiently according to ambient and the air after dehumidifier mass.

However, the air mixture mass line exhibits a downward trend with increasing time.

This occurs because the air after dehumidifier humidity increases with time and over

the whole experimental period, whilst the ambient air humidity remains within a

narrow band. This air mixture mass can accurately maintain the indoor requirement


92

(absolute humidity 0.007kg/kg) throughout the experiment. It should be noted that the

units in Figure 6-3 use absolute humidity instead of relative humidity. This is because

the indoor requirement is for constant absolute humidity. In this way, the air mixture

design can satisfy different air requirements. It should also be noted that the air

mixture design allows for strict control of indoor air humidity. The range of humidity

is between the air after dehumidifier and air mixing states. Since this is a hybrid

cooling system, the temperature of the indoor air is controlled by a conventional air

conditioning unit. Therefore, the temperature range can be controlled between 0 C

and ambient air temperature.

Figure 6-3 Air Mixture Mass Rate and Air Humidity Analysis

6.3 Weather Data Analysis

6.3.1 Summer Weather Data Analysis

Weather information is a major factor on the thermal performance prediction of liquid

desiccant cooling systems in simulations. The other reason the weather data should be

determined for analysis is that air mixture (AHU system) is used in the cooling system.

All averaged, daily summer weather data is included in Figure 6-4. It should be noted

that in Figure 6-4, the data includes weather readings at two time schedules in the day,

0.0 3.0 6.0 9.0 12.0 15.0 18.0 21.0 24.0 27.0 30.0 33.0 36.0 39.0

0.0

1.0

2.0

3.0

4.0

5.0

6.0

7.0

8.0

9.0

10.0

11.0

Mix

Air

Mass

(g/s

)

Ab

so

lute

Hu

mid

ity

(g/k

g)

Time (s)

Mixing Air Mass Rate and Air Humidity Relationship

Point 1 Air Absolute Humidity (g/kg) Point 2 Air Absolute Humidity (g/kg)

Point 3 Air Absolute Humidity (g/kg) Point 1 Mixing Air Mass (g/s)


93

being 9am (Figure 6-4 (a) and (c)) and 3pm (Figure 6-4 (b) and (d)). On average,

December (Figure 6-4 (a) and (b)) and January (Figure 6-4 (c) and (d)) have the

highest temperature and humidity ratio values in the whole year for Brisbane.

Detailed weather data for Brisbane is included in Appendix 1. According to indoor air

humidity and assumed temperature requirements, there are two lines that can be

drawn in Figure 6-4. The lines concern an ambient air temperature of 27 C, and an

absolute humidity of 7g/kg. Several points are below these two lines, which mean

these points don‘t need to be disposed by cooling devices or a dehumidifier. These

points are already satisfied, or they are less than the indoor requirements. Therefore,

the system can stop the cooling or dehumidification process when ambient air satisfies

the requirements. On the other hand, according to the weather data, points which are

below the 27 C line can not satisfy below the 7g/kg line at the same time. In a similar

situation, the points that can satisfy below the 7g/kg line, are not below the 27 C line

at the same time. Thus, dehumidification or cooling processes still need to dispose

these points. The COPsys depends on ECOP changing because there is no

dehumidifier working, or the COPsys depends on TCOP because there is no cooling

device working. This changing relationship between COPsys and ECOP or TCOP is

presented in Chapter 4 COPsys modeling analysis section.

(a) Brisbane Temperature and Humidity Analysis (Dec 2005 9am)

0.52.54.56.58.510.512.514.516.518.520.5

3.06.09.0

12.015.018.021.024.027.030.033.036.0

Ab

so

lute

Hu

mid

ity (

g/k

g)

Air

Te

mp

(D

eg

ree

)

Time

Temperature and Absolute Humidity Analysis (Brisbane December 2005 9am)

9am Temperature (°C) 9am Absolute Humidity (g/kg)

Absolute Humidity (7g/kg) Line

Air Temp (27 Degree)


94

(b) Brisbane Temperature and Humidity Analysis (Dec 2005 3pm)

(c) Brisbane Temperature and Humidity Analysis (Jan 2006 9am)

0.5

2.5

4.5

6.5

8.5

10.5

12.5

14.5

16.5

18.5

20.5

3.06.09.0

12.015.018.021.024.027.030.033.036.039.0

Ab

so

lute

Hu

mid

ity (

g/k

g)

Air

Tem

p (

Deg

ree)

Time

Temperature and Absolute Humidity Analysis (Brisbane December 2005 3pm)

3pm Temperature (°C) 3pm Absolute Humidity (g/kg)


Air Temp (27 Degree) Line

4.55.56.57.58.59.510.511.512.513.514.515.516.517.518.5

3.0

6.0

9.0

12.0

15.0

18.0

21.0

24.0

27.0

30.0

33.0

Ab

so

lute

Hu

mid

ity (

g/k

g)

Air

Tem

p (

Deg

ree)

Time

Temperature and Absolute Humidity Analysis (Brisbane January 2006 9am)

9am Temperature (°C) 9am Absolute Humidity (g/kg)




95

(d) Brisbane Temperature and Humidity Analysis (Dec 2006 3am)

Figure 6-4 Brisbane Temperatures and Humidity Analysis

6.3.2 Whole Year Weather Data Analysis

Figure 6-5(a), (b) (c) and (d) depict the changing daily and weekly ambient air

temperature and humidity experienced in Brisbane throughout a one year time period.

These graphs are displayed by the weather tool Ecotect 520, and all weather data is

recognized by the software database (Weather data .WEA file Australia, Brisbane

UQ1 19 May 02). These figures show the ambient air humidity changing quickly from

an extremely high level (around 40 C, 85%) to a low level (around 10 C, 10%) over

the whole year. Obviously, the humidity can change quickly even on different days in

the same week. However, the average temperature and humidity still vary according

to the different seasons. For example, the ambient air temperature in winter is lower

than summer, and average humidity in winter is lower than summer. These long term

weather data trends also support the hypothesis that ambient air humidity and

temperature on some days already satisfy indoor requirements. Therefore, this whole

year COPsys analysis using Chapter 4 COPsys analysis can still apply.

4.5

6.5

8.5

10.5

12.5

14.5

16.5

18.5

3.06.09.0

12.015.018.021.024.027.030.033.036.0

Ab

so

lute

Hu

mid

ity (

g/k

g)

Air

Tem

p (

Deg

ree)

Time

Temperature and Absolute Humidity Analysis (Brisbane January 2006 3am)

3pm Temperature (°C) 3am Absolute Humidity (g/kg)




96

(a) Brisbane Whole Year Relative Humidity (Daily Trend)

(b) Brisbane Whole Year Relative Humidity (Weekly Trend)


97

(c) Brisbane Whole Year Dry Bulb Temperature (Daily Trend)

(d) Brisbane Whole Year Dry Bulb Temperature (Weekly Trend)


98

(e) Brisbane Whole Year Wet Bulb Temperature (Daily Trend)

Figure 6-5 Brisbane Whole Year Weather Statistics

(WeaTool V1.10)

6.4 Ambient Air Latent Load and Sensible Load Analysis

The entire load of the air conditioning consists of latent load and sensible load. The

latent load is caused by the dehumidification process, and sensible load is caused by

the cooling process. The latent load is treated by the liquid desiccant system, while the

sensible load is overcome partially by conventional air conditioning in the hybrid

cooling system. Therefore, the latent load and sensible load in the air after

dehumidifier section should influence system performance. Here, sensible cooling

load, latent cooling load and total cooling load are defined as:

Sensible cooling load:

Equation 6-7 S = ( )p o iC T T m

Latent cooling load:

Equation 6-8 L = ( )fg o ih m

Total Load:


99

Equation 6-9 totalQ = out inL S L S S

Here the outS is the sensible load from processed outdoor air and inS is the remaining

sensible load which has been changed by the fan, pump and other devices.

The Sensible Heat Ratio (SHR) expresses the ratio between the sensible heat load and

the total heat load, which is shown in Equation 6-10:

Equation 6-10

sensible heat

total

SSHR

Q

Here

S sensible heat load (kW)

totalQ total heat load (kW)

Equation 6-10 can be modified to:

Equation 6-11 0 0( ) /( )p i iSHR c t t h h

Where

pc = specific heat capacity of air (kJ/kg.ºC)

0t = outlet temperature (ºC)

it = inlet temperature(ºC)

0h = outlet enthaply (kJ/kg)

ih =inlet enthaply (kJ/kg)

Similarly, the Room Sensible Heat Ratio (RSHR) express the ratio between the

sensible heat load and the total heat load in the room. It can be expressed as:

Equation 6-12

room sensible heat

total heat

SRSHR

Q

Here

room sensible heatQ sensible heat load in the room (kW)


100

total heatQ total heat load in the room (kW)

Equation 6-12 can be modified to:

Equation 6-13 ( ) /( )p r i r iRSHR c t t h h

Where

pc= specific heat capacity of air (kJ/kg.ºC)

rt = room temperature (ºC)

it = inlet temperature(ºC)

rh= outlet enthaply (kJ/kg)

ih= inlet enthaply (kJ/kg)

At the same time, the ratio of latent load to total load is defined as Equation 6-14 (Ma

et al. 2006)

Equation 6-14 total load

L

Q

So that,

Equation 6-15 SHR total

S

Q+

total

L

Q

Then,

Equation 6-16 SHR 1S L

S L

According to Equation 6-16, (the ratio of latent load) can be calculated by 1- SHR .

If sensible heat ratio is bigger than the ratio of latent load, the ambient air ratio of

latent load to total load is very small. The quantity of heat required to regenerate the

liquid desiccant is also accordantly small. It means that the ambient air humidity is in


101

the low level. Therefore, according to TCOP definition, c

T

QTCOP

Q , TQ is small,

while TCOP will increase while assuming cQ is stable. If the ambient air humidity

already satisfies indoor requirements as indicated in the previous section, the system

does not need a heat supplement. The COPsys will totally depend on ECOP change

only, and ECOP only depends on the cooling quantity required to deal with sensible

load. Under the same sensible load situation, the whole system COPsys change

depends on TCOP change. The quantity of water absorbed into concentrated desiccant

in the dehumidifier is theoretically equal to that latent load removed from ambient air.

If ambient air latent load is small, the air mixture section is also very small to satisfy

indoor air requirements. Air mixture part has only a small effect on the whole system.

On the other hand, if the ratio of latent load to total load in the ambient air, , is large

enough, the heat quantity required to regenerate liquid desiccant is also large. Under

this situation, where c

T

QTCOP

Q , and TQ is large, the TCOP will become low

because more energy will be needed to deal with latent load. This is the situation if

cQ does not have a large change, while the cooling requirement and air mixture part

decide the sensible cooling load. Therefore ECOP changing simply depends on the air

mixing amount and indoor cooling requirements. According to COPsys definition,

csys

T c other

QCOP

Q Q W

, COPsys will decrease when is as large as a threshold

value, if the cooling requirement does not change significantly.


One air mixture analysis example used in a hybrid liquid desiccant cooling system is

presented in this chapter. The mixing air result shows that air mixture (AHU) design

can solve the conventional liquid desiccant system problem. The humidity can be

accurately achieved by mixing ambient air into the system. However, the inlet

disposing air ratio decreases as the ambient air humidity increase, when the system

dehumidifier supply is stable. On the other hand in this example, the ambient air data


102

was only recorded over several hours. The ratio of latent load to total load, , was

defined in this chapter. Room Sensible Heat Ratio (RSHR) and relationship with are

also mentioned here. The effect of on the ECOP, TCOP and COPsys is presented in

this study. According to the long term meteorologic data, there is ambient data

already satisfying indoor requirements. Therefore, the weather data can be separated

into two different groups: one that satisfies the requirements and one that does not

satisfy the requirements. When the ambient air already satisfies the requirements, the

sensible cooling load or latent cooling load can be saved. Supply air can be used from

ambient air directly or mixture partly according to supply humidity and temperature

requirement. Based on this, some applications requiring air mix preconditioning may

result in large annual electrical energy saving and improved indoor humidity control.

Thus, the analysis includes daily averaged summer weather data and whole year

weekly humidity and temperature trends. Ambient air latent load is caused by ambient

air humidity, and the latent load can affect system performance significantly. Some

quantitative analysis can be undertaken in the future research.


103

7 CONSLUSIONS AND RECOMMENDATIONS

It is clear that the effects of the design variables on the performance of the hybrid

liquid desiccant solar cooling system are of great interest. The coefficients of

performance, including COPcon, ECOP, TCOP and COPsys of hybrid liquid

desiccant solar cooling system have been proposed and analysed in this thesis. This

solar cooling system used some air mixture (AHU) design to adjust system humidity.

Based on this design, some mathematical models for coefficient of performance have

been devised. The air mixture design appears to be attractive for air adjustment and

humidity control. The relationship between COPcon, ECOP, TCOP and COPsys has

been developed and used to derive mathematical models for each component of the

system. The system coefficient simulations using these mathematical models have

been performed to identify the optimum system configuration under different weather

conditions. Some experimental data for air humidity and temperature in the hybrid

liquid desiccant solar cooling system are also included in this study.

In the following sections, the major conclusions that can be made regarding this work

will be presented. Some recommendations will also be made on possible extensions of

this work for further research and development.

7.1 Conclusions

The experimental data of dehumidification and regeneration in this study

shows that the liquid desiccant system is not flexible. The packed bed

structure dehumidifier/regenerator displays a lag response when the desiccant

pump stops running.

It is possible to design and construct an air mixture hybrid liquid desiccant

system to solve the over dehumidification problem. This design is based on air

mixture concepts in air-conditioning.

Because the conventional air conditioning section is not responsible for the

latent load, performance of conventional air conditioning is dependent on


104

cooling load. The ambient air temperature has a significant effect on the

electrical coefficient of performance of system; in the other words, the

electrical coefficient of performance decreases with the increasing outdoor

temperature.

The present study revealed that when the thermal energy remains stable, the

thermal coefficient of performance of the system depends on ambient air

humidity. In the same time, the total cooling load strongly affects the thermal

coefficient of performance.

According to correlations analysis between different COP, COPsys variation

depends on thermal performance and electrical performance changes. Several

variables have been found to have the greatest impact on the performance of a

hybrid cooling system (COPsys) with air mixture design. They are: 1) the

ambient air temperature and humidity, 2) the air after dehumidifier

temperature and humidity.

The overall COPsys of the hybrid cooling system displays big difference when

the system uses solar energy or without uses solar energy. It is much higher

COP when the hybrid cooling system using solar energy.

The air mixture design for the hybrid liquid desiccant system shows a good

result for balancing humidity of the air after dehumidification. The air mixture

model calculation of the system provides some predictions based on air

mixture theory.

Based on the long term meteorologic data, some ambient air already satisfies

indoor requirements. Some applications for air mixing preconditioning can

result in large annual electrical energy savings.


105

7.2 Recommendations for Future Research

The present investigation gives valuable insight into the design of hybrid liquid

desiccant cooling systems. The important air mixture design for liquid desiccant

system has been proposed, and the derived performance correlations are valuable for

future system design. The following suggestions need consideration for further

research and development.

Based on the promising seasonal results from the performance simulation of

air mixture for the proposed desiccant cooling system, daily performance

simulations of the air mixture system for applications should be carried out in

the future. The Typical year weather data will be used to do analysis in the

whole hybrid system.

With respect to the air mixture design variables previously mentioned, an

optimization based on cost, energy efficiency, and the quality of the air

conditioning process would be valuable.

Furthermore, as some auxiliary controlling sections for the terminal controller

will be required for the air mixture process, relative software programming

and hardware design details should be explored in future research.

Experimental results in full scale demonstrations of liquid desiccant cooling

show that it will be necessary to address issues such as desiccant carry-over

into the air after dehumidifier, liquid desiccant and equipment design, the

compactness of the system, and possible filter degradation for the desiccant

carry-over.

With further research it should be possible to develop cost-competitive and

energy efficient hybrid liquid solar cooling system, using some air mixture

techniques. After all, solar cooling has the advantage of using the largest

amount of solar energy to regenerate liquid desiccant to save electricity energy.


106

The ratio of latent load to total load, , can be defined in this study. Some

further quantitative analysis of the effect of for the ECOP, TCOP and

COPsys could be undertaken in a following liquid desiccant cooling system

study.

The desiccant carryout energy and relative mass transfer affect the hybrid

cooling system efficiency. Therefore the energy carryout calculations in the

dehumidification process can be done in the future study.


107

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110

9 APPENDICES

Appendix 1

Daily Weather Observations for Brisbane, Queensland for December 2005/January 2006


111

Table 1 Daily Weather Observations for Brisbane, Queensland for December 2005 All data come from Australia Commonwealth Bureau of Meteorology

Date 9am Temperature (°C) 9am relative humidity (%) 9am MSL pressure (hPa) 3pm Temperature (°C) 3pm relative humidity (%) 3pm MSL pressure (hPa) 2005-12-1 25.9 68.0 1013.1 24.8 78.0 1012.0

2005-12-2 24.1 81.0 1012.1 23.6 83.0 1008.8

2005-12-3 26.0 60.0 1007.6 30.2 33.0 1004.8

2005-12-4 26.4 44.0 1009.3 32.9 24.0 1005.6

2005-12-5 28.6 60.0 1010.6 31.0 64.0 1007.9

2005-12-6 30.0 61.0 1011.1 30.6 63.0 1008.0

2005-12-7 29.7 57.0 1012.0 30.4 62.0 1008.4

2005-12-8 28.9 64.0 1010.8 30.8 62.0 1005.4

2005-12-9 30.3 58.0 1007.9 30.6 44.0 1006.1

2005-12-10 27.2 58.0 1010.6 32.7 40.0 1006.3

2005-12-11 27.3 60.0 1012.7 28.3 52.0 1010.6

2005-12-12 26.6 54.0 1012.3 26.7 59.0 1008.1

2005-12-13 27.1 67.0 1005.2 30.8 59.0 1001.2

2005-12-14 32.4 48.0 1006.1 31.5 53.0 1005.8

2005-12-15 27.3 63.0 1014.5 29.2 55.0 1012.5

2005-12-16 26.5 69.0 1011.5 25.5 82.0 1006.0

2005-12-17 28.7 73.0 1003.4 23.1 67.0 1000.4

2005-12-18 29.6 27.0 1003.0 31.7 20.0 1002.4

2005-12-19 25.7 26.0 1010.8 32.1 17.0 1008.4

2005-12-20 26.6 48.0 1017.2 28.7 45.0 1014.6

2005-12-21 27.9 54.0 1018.1 29.9 56.0 1015.4

2005-12-22 29.6 48.0 1017.3 29.6 59.0 1014.3

2005-12-23 29.7 56.0 1016.6 30.4 57.0 1013.3

2005-12-24 28.0 53.0 1013.9 30.8 58.0 1008.5

2005-12-25 32.1 57.0 1009.7 29.4 72.0 1007.9

2005-12-26 28.3 61.0 1013.5 29.3 65.0 1011.6

2005-12-27 30.4 67.0 1016.2 30.6 59.0 1013.1

2005-12-28 29.5 58.0 1016.2 31.0 61.0 1011.6

2005-12-29 30.1 56.0 1014.3 34.7 49.0 1011.0

2005-12-30 31.2 47.0 1017.5 32.0 43.0 1015.5

2005-12-31 28.3 52.0 1020.6 29.0 51.0 1018.3


112

Table 2 Daily Weather Observations for Brisbane, Queensland for January 2006 All data come from Australia Commonwealth Bureau of Meteorology

Date 9am Temperature (°C) 9am relative humidity (%) 9am MSL pressure (hPa) 3pm Temperature (°C) 3pm relative humidity (%) 3pm MSL pressure (hPa) 2006-1-1 27.1 56.0 1019.0 28.6 59.0 1015.0

2006-1-2 29.5 58.0 1012.9 31.6 56.0 1011.5

2006-1-3 30.3 58.0 1011.5 31.4 57.0 1008.1

2006-1-4 28.6 65.0 1012.5 30.0 59.0 1011.0

2006-1-5 26.1 74.0 1011.8 30.0 65.0 1009.4

2006-1-6 29.1 62.0 1010.8 30.7 62.0 1007.6

2006-1-7 24.3 88.0 1010.9 28.0 64.0 1007.7

2006-1-8 23.3 92.0 1009.1 25.9 71.0 1008.0

2006-1-9 25.7 72.0 1010.7 27.9 66.0 1008.6

2006-1-10 28.7 70.0 1014.6 30.1 65.0 1013.0

2006-1-11 26.5 80.0 1014.8 30.7 59.0 1011.9

2006-1-12 29.1 65.0 1014.4 30.6 58.0 1011.3

2006-1-13 29.2 57.0 1013.4 29.6 60.0 1011.4

2006-1-14 27.5 61.0 1014.5 29.1 57.0 1012.3

2006-1-15 28.1 58.0 1015.1 30.1 52.0 1013.6

2006-1-16 27.9 64.0 1016.0 29.9 54.0 1013.8

2006-1-17 27.8 55.0 1014.1 29.8 58.0 1011.1

2006-1-18 29.0 57.0 1011.4 30.3 57.0 1009.2

2006-1-19 27.0 66.0 1013.8 24.8 83.0 1012.9

2006-1-20 25.1 86.0 1014.9 23.3 92.0 1013.8

2006-1-21 25.3 81.0 1014.6 28.7 56.0 1011.9

2006-1-22 26.6 69.0 1011.8 28.2 63.0 1009.9

2006-1-23 27.4 70.0 1010.1 29.9 60.0 1007.4

2006-1-24 28.7 66.0 1009.2 31.3 57.0 1006.7

2006-1-25 28.9 65.0 1014.5 29.2 57.0 1012.9

2006-1-26 27.5 67.0 1014.7 29.2 51.0 1012.5

2006-1-27 27.0 62.0 1011.5 28.9 54.0 1009.2

2006-1-28 25.7 64.0 1011.4 27.8 45.0 1009.7

2006-1-29 27.4 55.0 1010.6 29.0 54.0 1008.8

2006-1-30 24.1 74.0 1009.0 29.0 54.0 1005.3

2006-1-31 28.7 57.0 1007.5 30.0 60.0 1005.7


113

Appendix 2

Coefficient of Performance Calculation Results

ECOP TCOP COPsys Qc Wother e

ECOP=Qc/Wc TCOP=Qc/QT COPsys=1/[(1/TCOP)+(1/ECOP)+(Wother/Qc)] Qc Wother e=Wother/Qc


Equal COP

Use Solar

Energy

Not Use Solar

Energy

Use Solar

Energy

Not Use Solar

Energy Equal ECOP

Use Solar Energy

Not Use Solar Energy

Cooling load (kW)

kW

5.0 6.7 6.7 - - 1.9 5.0 6.2 1.5 8.99 0.1 0.0111

4.2 6.4 6.4 - - 1.9 4.2 5.9 1.5 8.51 0.1 0.0117

4.0 5.6 5.6 - - 2.0 4.0 5.2 1.5 7.69 0.1 0.0130

4.8 5.3 5.3 - - 2.0 4.8 5.0 1.5 7.28 0.1 0.0137

3.1 5.8 5.8 - - 1.9 3.1 5.4 1.4 7.56 0.1 0.0132

3.8 5.5 5.5 - - 1.8 3.8 5.1 1.4 6.91 0.1 0.0145

2.5 4.8 4.8 - - 2.1 2.5 4.6 1.4 6.59 0.1 0.0152

CAC=Conventional Vapour Compression Air Conditioning; HAC=Hybrid Liquid Desiccant Air Conditioning; QC=Total Cooling Load on the system

(W/kW); QT=Total Thermal Energy Input (W/kW); WC=Total Cooling Electrical Input (W/kW); Wother=Total Other Electrical Input (W/kW)


114

Appendix 3

Different Points Calculation Results and DAIKIN Air Conditioning Cooling Capacity

T1 RH1 AH1 m1 QC (W/kW) QT (W/kW) WC (W/kW) Wother (W/kW)

Supply Air Temp. (°CDB)

Supply Air Relative

Humidity (%)

Supply Air Absolute

Humidity (g/kg)

Outdoor Mass Flow Rate

(kg/s)

Cooling Load (W/kW) Thermal Energy

Input (W/kW) Cooling Electrical

Input (W/kW) Total Other Electrical

Input (W/kW)

CAC HAC CAC HAC CAC HAC CAC HAC

23.3 92 16.6 0.0286 12307.5 7549.0 0.0 4000.0 2471.4 1125.0 0.0 100.0

24.8 83 16.4 0.0282 12650.7 7626.9 0.0 4000.0 3048.4 1201.1 0.0 100.0

25.9 71 15.0 0.0408 12255.2 8087.9 0.0 4000.0 3102.6 1446.8 0.0 100.0

27.8 45 10.5 0.0698 9604.0 8079.8 0.0 4000.0 2000.8 1513.1 0.0 100.0

29.0 54 13.6 0.0390 12061.6 7571.6 0.0 4000.0 3903.4 1314.5 0.0 100.0

30.0 60 16.1 0.0275 14202.1 7319.8 0.0 4000.0 3777.1 1338.2 0.0 100.0

31.6 56 16.5 0.0469 15975.8 8315.1 0.0 4000.0 6467.9 1718.0 0.0 100.0

CAC=Conventional Vapour Compression Air Conditioning; HAC=Hybrid Liquid Desiccant Air Conditioning; QC=Total Cooling Load on the system (W/kW); QT=Total Thermal Energy Input (W/kW); WC=Total Cooling Electrical Input (W/kW); Wother=Total Other Electrical Input (W/kW)


115

T1 RH1 m1 QC (W/kW) QT (W/kW) ECOP TCOP COPsys

Supply Air Temp. (°CDB)

Supply Air Relative Humidity

(%)

Outdoor Mass

Flow Rate (kg/s)

Cooling Load (W/kW)

Thermal Energy Input (W/kW)

ECOP=Qc/Wc TCOP=Qc/QT COPsys=Qc/(QT+Wc+Wother)

CAC HAC CAC HAC


Equal COP

Use Solar

Energy

Not Use Solar

Energy

Use Solar

Energy

Not Use Solar

Energy

Equal ECOP

Use Solar

Energy

Not Use Solar

Energy

23.3 92 0.0286 12307.5 7549.0 0.0 4000.0 5.0 6.7 6.7 - - 1.9 5.0 6.2 1.5

24.8 83 0.0282 12650.7 7626.9 0.0 4000.0 4.2 6.4 6.4 - - 1.9 4.2 5.9 1.5

25.9 71 0.0408 12255.2 8087.9 0.0 4000.0 4.0 5.6 5.6 - - 2.0 4.0 5.2 1.5

27.8 45 0.0698 9604.0 8079.8 0.0 4000.0 4.8 5.3 5.3 - - 2.0 4.8 5.0 1.5

29.0 54 0.0390 12061.6 7571.6 0.0 4000.0 3.1 5.8 5.8 - - 1.9 3.1 5.4 1.4

30.0 60 0.0275 14202.1 7319.8 0.0 4000.0 3.8 5.5 5.5 - - 1.8 3.8 5.1 1.4

31.6 56 0.0469 15975.8 8315.1 0.0 4000.0 2.5 4.8 4.8 - - 2.1 2.5 4.6 1.4

CAC=Conventional Vapour Compression Air Conditioning; HAC=Hybrid Liquid Desiccant Air Conditioning; QC=Total Cooling Load on the system (W/kW); QT=Total Thermal Energy Input (W/kW); ECOP=Qc/Wc; TCOP=Qc/QT; COPsys=Qc/(QT+Wc+Wother); Use Solar Energy: HAC use solar energy to work as thermal energy input


116

Daikin RMK140J Outdoor Units Cooling Capacity (50Hz 220-240V, 60Hz 220-230V)

Outdoor Air Temp. °CDB

Combination 100%(14.5kW) Combination 90%(13.1kW) Combination 80%(11.6kW)

Indoor Air Temp. 18.0°CWB Indoor Air Temp. 18.0°CWB Indoor Air Temp. 18.0°CWB

TC (kW) PI (kW) COP TC (kW) PI (kW) COP TC (kW) PI (kW) COP

23°C 15.31 4.05 3.78 13.83 3.17 4.36 12.25 2.46 4.98

25°C 15.10 4.21 3.59 13.64 3.29 4.15 12.08 2.55 4.74

27°C 14.89 4.36 3.42 13.46 3.41 3.95 11.92 2.64 4.52

29°C 14.69 4.52 3.25 13.27 3.53 3.76 11.75 2.74 4.29

31°C 14.48 4.68 3.09 13.08 3.66 3.57 11.58 2.83 4.09

33°C 14.27 4.83 2.95 12.89 3.78 3.41 11.41 2.93 3.89

35°C 14.06 4.99 2.82 12.70 3.90 3.26 11.25 3.02 3.73

37°C 13.85 5.14 2.69 12.51 4.02 3.11 11.08 3.12 3.55

Outdoor Air Temp. °CDB

Combination 70%(10.2kW) Combination 60%(8.7kW) Combination 51.7%(7.5kW)

Indoor Air Temp. 18.0°CWB Indoor Air Temp. 18.0°CWB Indoor Air Temp. 18.0°CWB

TC (kW) PI (kW) COP TC (kW) PI (kW) COP TC (kW) PI (kW) COP

23°C 10.77 1.93 5.58 9.19 1.48 6.21 7.92 1.18 6.71

25°C 10.62 2.00 5.31 9.06 1.54 5.88 7.81 1.23 6.35

27°C 10.48 2.08 5.04 8.94 1.60 5.59 7.70 1.27 6.06

29°C 10.33 2.15 4.80 8.81 1.65 5.34 7.60 1.32 5.76

31°C 10.18 2.23 4.57 8.69 1.71 5.08 7.49 1.37 5.47

33°C 10.04 2.30 4.37 8.56 1.77 4.84 7.38 1.41 5.23

35°C 9.89 2.37 4.17 8.44 1.83 4.61 7.27 1.46 4.98

37°C 9.74 2.45 3.98 8.31 1.88 4.42 7.16 1.50 4.77


117

Appendix 4

Ecotect 520 Weather Data (http://ecotect.com/downloads/weatherdata)

The hourly weather data files provided on this page are .WEA files (139 Kb) and are

Binary. They are intended as example data for use in building simulation and analysis

-- if you require specific year data or data source information, it is recommended that

you source your own (and use the Weather Tool to import the data) as we are unable

to provide this information.

The majority of files listed below, contain most if not all of the following:

Dry Bulb Temperature

Relative Humidity

Direct Solar Radiation

Diffuse Horizontal Solar Radiation

Wind speed (not essential but preferable)

Wind direction (not essential but preferable)

Cloudiness

Rainfall (not essential but preferable)

These files are for use with ECOTECT and the Weather Tool. To view the .WEA files

you will need to download a copy of the Weather Tool. Other data Formats and

sources. EPW Energy Plus Weather Data

The most reliable and comprehensive source of international weather data at the

moment would have to be the EnergyPlus Weather Data site. With v1.20+ of the

Weather Tool you can simply drag and drop the .EPW files from this site in to the

Weather Tool to load them. Data for over 900+ international locations are available...

Generic ASCII data

Any format of ASCII (text only, human readable) format can be imported into the

Weather Tool. For more detailed information about this process please refer to the

Import Data section of the Weather Tool help file, and specifically the included

tutorials:

Importing Fixed Format Data

Importing CSV Data

'tas' Weather Data

Some users might have access to 'tas' weather data. If so then this Excel Macro for

preparing your data should be useful. Read this tutorial from the Weather Tool help to

view the code for the macro in full, as well as an explanation for its use.

Import Format Files

In addition to the already converted .WEA files provided below, the following

Custom Column Format (.CCF) files may be useful to some users.

http://ecotect.com/weathertool

http://ecotect.com/glossary/12

http://ecotect.com/ecotect


http://ecotect.com/downloads

http://www.eere.energy.gov/buildings/energyplus/cfm/weather_data.cfm


http://squ1.org/wiki/WeatherTool/Importing_Data

http://squ1.org/wiki/WeatherTool

http://squ1.org/wiki/WeatherTool/Tutorial_Fixed

http://squ1.org/wiki/WeatherTool/Tutorial_CSV

http://companyshed.com/downloads/weatherdata/tasCSVtoWEA.zip

http://squ1.org/wiki/WeatherTool/Tutorial_tasXLStoWEA


118

EnergyPlus CCF

HTB2_import CCF

HTB2_export CCF

Meteonorm CCF

tasXLStoCSVtoWEA CCF

The EnergyPlus conversion format is particularly useful with the data available from

the EnergyPlus Weather Data site (with v1.10 or earlier of the Weather Tool).

A Donated/Updated Date

Australia - All (.ZIP 2,951Kb) 19 May '02

Australia - Adelaide SA - 1 19 May '02

Australia - Adelaide SA - 2 19 May '02

Australia - Albany WA 19 May '02

Australia - Alice Springs NT 19 May '02

Australia - Amberley QU 19 May '02

Australia - Bordertown SA 19 May '02

Australia - Brisbane QU - 1 19 May '02

Australia - Brisbane QU - 2 19 May '02

Australia - Cairns QU - 1 19 May '02

Australia - Cairns QU - 2 19 May '02

Australia - Canberra ACT - 1 19 May '02

Australia - Canberra ACT - 2 19 May '02

Australia - Carnarvon WA 19 May '02

Australia - Darwin NT - 1 19 May '02




Australia - Geraldton WA 19 May '02

Australia - Halls Creek WA 19 May '02

Australia - Hobart TAS 19 May '02

Australia - Kalgoorlie WA 19 May '02

Australia - Launceston TAS 19 May '02

Australia - Longreach QU 19 May '02

http://companyshed.com/downloads/weatherdata/EnergyPlus.zip

http://companyshed.com/downloads/weatherdata/HTB2_import.zip

http://companyshed.com/downloads/weatherdata/HTB2_export.zip

http://companyshed.com/downloads/weatherdata/Meteonorm.zip

http://companyshed.com/downloads/weatherdata/tasCSVtoWEA.zip

http://www.eere.energy.gov/buildings/energyplus/cfm/weather_data.cfm


http://companyshed.com/downloads/weatherdata/Australia.zip

http://companyshed.com/downloads/weatherdata/Australia-AdelaideSA-1.wea

http://companyshed.com/downloads/weatherdata/Australia-AdelaideSA-2.wea

http://companyshed.com/downloads/weatherdata/Australia-AlbanyWA.wea

http://companyshed.com/downloads/weatherdata/Australia-AliceSprings%20NT.wea

http://companyshed.com/downloads/weatherdata/Australia-AmberleyQU.wea

http://companyshed.com/downloads/weatherdata/Australia-BordertownSA.wea

http://companyshed.com/downloads/weatherdata/Australia-BrisbaneQU-1.wea

http://companyshed.com/downloads/weatherdata/Australia-BrisbaneQU-2.wea

http://companyshed.com/downloads/weatherdata/Australia-CairnsQU-1.wea

http://companyshed.com/downloads/weatherdata/Australia-CairnsQU-2.wea

http://companyshed.com/downloads/weatherdata/Australia-CanberraACT-1.wea

http://companyshed.com/downloads/weatherdata/Australia-CanberraACT-2.wea

http://companyshed.com/downloads/weatherdata/Australia-CarnarvonWA.wea

http://companyshed.com/downloads/weatherdata/Australia-DarwinNT-1.wea




http://companyshed.com/downloads/weatherdata/Australia-GeraldtonWA.wea

http://companyshed.com/downloads/weatherdata/Australia-HallsCreekWA.wea

http://companyshed.com/downloads/weatherdata/Australia-HobartTAS.wea

http://companyshed.com/downloads/weatherdata/Australia-KalgoorlieWA.wea

http://companyshed.com/downloads/weatherdata/Australia-LauncestonTAS.wea

http://companyshed.com/downloads/weatherdata/Australia-LongreachQU.wea


119

Australia - Melbourne VIC - 1 19 May '02

Australia - Melbourne VIC - 2 19 May '02

Australia - Mildura VIC 19 May '02

Australia - Moree NSW 19 May '02

Australia - Mt Gambia SA 19 May '02

Australia - Perth WA - 1 19 May '02



Australia - Port Hedland WA 19 May '02

Australia - Richmond NSW - 1 19 May '02

Australia - Richmond NSW - 2 19 May '02

Australia - Rockhampton QU 19 May '02

Australia - Sydney NSW - 1 19 May '02



Australia - Tamworth NSW 19 May '02

Australia - Townsville QU - 1 19 May '02

Australia - Townsville QU - 2 19 May '02

Australia - Wagga Wagga NSW - 1 19 May '02

Australia - Wagga Wagga NSW - 2 19 May '02

Australia - Whyalla SA 19 May '02

Australia - Williamtown NSW 19 May '02

Australia - Wiluna WA 19 May '02

Austria - Vienna 19 May '02

B Donated/Updated Date

Belgium - Brussels 19 May '02

Belgium - Oostende 19 May '02

Belgium - St Hubert 19 May '02

C Donated/Updated Date

Canada - All (.ZIP 1,144Kb) 19 May '02

Canada - Abbotsford BC 19 May '02

Canada - Comox BC 19 May '02

Canada - Edmonton AB 19 May '02

http://companyshed.com/downloads/weatherdata/Australia-MelbourneVIC-1.wea

http://companyshed.com/downloads/weatherdata/Australia-MelbourneVIC-2.wea

http://companyshed.com/downloads/weatherdata/Australia-MilduraVIC.wea

http://companyshed.com/downloads/weatherdata/Australia-MoreeNSW.wea

http://companyshed.com/downloads/weatherdata/Australia-MtGambiaSA.wea

http://companyshed.com/downloads/weatherdata/Australia-PerthWA-1.wea



http://companyshed.com/downloads/weatherdata/Australia-PortHedlandWA.wea

http://companyshed.com/downloads/weatherdata/Australia-RichmondNSW-1.wea

http://companyshed.com/downloads/weatherdata/Australia-RichmondNSW-2.wea

http://companyshed.com/downloads/weatherdata/Australia-RockhamptonQU.wea

http://companyshed.com/downloads/weatherdata/Australia-SydneyNSW-1.wea



http://companyshed.com/downloads/weatherdata/Australia-TamworthNSW.wea

http://companyshed.com/downloads/weatherdata/Australia-TownsvilleQU-1.wea

http://companyshed.com/downloads/weatherdata/Australia-TownsvilleQU-2.wea

http://companyshed.com/downloads/weatherdata/Australia-WaggaWaggaNSW-1.wea

http://companyshed.com/downloads/weatherdata/Australia-WaggaWaggaNSW-2.wea

http://companyshed.com/downloads/weatherdata/Australia-WhyallaSA.wea

http://companyshed.com/downloads/weatherdata/Australia-WilliamtownNSW.wea

http://companyshed.com/downloads/weatherdata/Australia-WilunaWA.wea

http://companyshed.com/downloads/weatherdata/Austria-Vienna.wea

http://companyshed.com/downloads/weatherdata/Belgium-Brussels.wea

http://companyshed.com/downloads/weatherdata/Belgium-Oostende.wea

http://companyshed.com/downloads/weatherdata/Belgium-StHubert.wea

http://companyshed.com/downloads/weatherdata/Canada.zip

http://companyshed.com/downloads/weatherdata/Canada-AbbotsfordBC.wea

http://companyshed.com/downloads/weatherdata/Canada-ComoxBC.wea

http://companyshed.com/downloads/weatherdata/Canada-EdmontonAB.wea


120

Canada - Fort St John BC 19 May '02

Canada - Kamloops BC 19 May '02

Canada - Montreal QU 19 May '02

Canada - Port Hardy BC 19 May '02

Canada - Prince George BC 19 May '02

Canada - Prince Rupert BC 19 May '02

Canada - Sandspit BC 19 May '02

Canada - Smithers BC 19 May '02

Canada - Summerland BC 19 May '02

Canada - Toronto OT 19 May '02

Canada - Vancouver BC - 1 19 May '02

Canada - Vancouver BC - 2 19 May '02

Canada - Victoria BC 19 May '02

Canada - Winnipeg MA 19 May '02

China - Fushun Xian 19 May '02

China - Hong Kong 19 May '02

D Donated/Updated Date

Denmark - Copenhagen - 1 19 May '02

Denmark - Copenhagen - 2 19 May '02

F Donated/Updated Date

France - All (.ZIP 449Kb) 19 May '02

France - Carpentras 19 May '02

France - Limoges 19 May '02

France - Macon - 1 19 May '02

France - Macon - 2 19 May '02

France - Nancy 19 May '02

France - Nice 19 May '02

France - Trappes 19 May '02

G Donated/Updated Date

Germany - All (.ZIP 723Kb) 19 May '02

Germany - Augsburg 19 May '02

Germany - Berlin 19 May '02

Germany - Bremen 19 May '02

http://companyshed.com/downloads/weatherdata/Canada-FortStJohnBC.wea

http://companyshed.com/downloads/weatherdata/Canada-KamloopsBC.wea

http://companyshed.com/downloads/weatherdata/Canada-MontrealQU.wea

http://companyshed.com/downloads/weatherdata/Canada-PortHardyBC.wea

http://companyshed.com/downloads/weatherdata/Canada-PrinceGeorgeBC.wea

http://companyshed.com/downloads/weatherdata/Canada-PrinceRupertBC.wea

http://companyshed.com/downloads/weatherdata/Canada-SandspitBC.wea

http://companyshed.com/downloads/weatherdata/Canada-SmithersBC.wea

http://companyshed.com/downloads/weatherdata/Canada-SummerlandBC.wea

http://companyshed.com/downloads/weatherdata/Canada-TorontoOT.wea

http://companyshed.com/downloads/weatherdata/Canada-VancouverBC-1.wea

http://companyshed.com/downloads/weatherdata/Canada-VancouverBC-2.wea

http://companyshed.com/downloads/weatherdata/Canada-VictoriaBC.wea

http://companyshed.com/downloads/weatherdata/Canada-WinnipegMA.wea

http://companyshed.com/downloads/weatherdata/China-FushunXian.wea

http://companyshed.com/downloads/weatherdata/China-HongKong.wea

http://companyshed.com/downloads/weatherdata/Denmark-Copenhagen-1.wea

http://companyshed.com/downloads/weatherdata/Denmark-Copenhagen-2.wea

http://companyshed.com/downloads/weatherdata/France.zip

http://companyshed.com/downloads/weatherdata/France-Carpentras.wea

http://companyshed.com/downloads/weatherdata/France-Limoges.wea

http://companyshed.com/downloads/weatherdata/France-Macon-1.wea

http://companyshed.com/downloads/weatherdata/France-Macon-2.wea

http://companyshed.com/downloads/weatherdata/France-Nancy.wea

http://companyshed.com/downloads/weatherdata/France-Nice.wea

http://companyshed.com/downloads/weatherdata/France-Trappes.wea

http://companyshed.com/downloads/weatherdata/Germany.zip

http://companyshed.com/downloads/weatherdata/Germany-Augsburg.wea

http://companyshed.com/downloads/weatherdata/Germany-Berlin.wea

http://companyshed.com/downloads/weatherdata/Germany-Bremen.wea


121

Germany - Dresden 19 May '02

Germany - Essen 19 May '02

Germany - Frankfurt 19 May '02

Germany - Freiburg 19 May '02

Germany - Hannover 19 May '02

Germany - Munich 19 May '02

Germany - Trier 19 May '02

Germany - Wuerzburg 19 May '02

Greece - Athens 19 May '02

I Donated/Updated Date

Ireland - Dublin 19 May '02

Ireland - Valentia 19 May '02

Italy - All (.ZIP 686Kb) 19 May '02

Italy - Bolzano 19 May '02

Italy - Cagliari 19 May '02

Italy - Crotone 19 May '02

Italy - Firenze 19 May '02

Italy - Foggia 19 May '02

Italy - Genova 19 May '02

Italy - Milano 19 May '02

Italy - Monte Terminillo 19 May '02

Italy - Rome 19 May '02

Italy - Trapani 19 May '02

Italy - Venice 19 May '02

K Donated/Updated Date

Kenya - Nairobi 19 May '02

M Donated/Updated Date

Malaysia - All (.ZIP 345Kb) 19 May '02

Malaysia - Cameron Highlands 19 May '02

Malaysia - Kota Kinabulu 19 May '02

Malaysia - Kuala Lumpur 19 May '02

Malaysia - Kuantan 19 May '02

Malaysia - Kuching 19 May '02

http://companyshed.com/downloads/weatherdata/Germany-Dresden.wea

http://companyshed.com/downloads/weatherdata/Germany-Essen.wea

http://companyshed.com/downloads/weatherdata/Germany-Frankfurt.wea

http://companyshed.com/downloads/weatherdata/Germany-Freiburg.wea

http://companyshed.com/downloads/weatherdata/Germany-Hannover.wea

http://companyshed.com/downloads/weatherdata/Germany-Munich.wea

http://companyshed.com/downloads/weatherdata/Germany-Trier.wea

http://companyshed.com/downloads/weatherdata/Germany-Wuerzburg.wea

http://companyshed.com/downloads/weatherdata/Greece-Athens.wea

http://companyshed.com/downloads/weatherdata/Ireland-Dublin.wea

http://companyshed.com/downloads/weatherdata/Ireland-Valentia.wea

http://companyshed.com/downloads/weatherdata/Italy.zip

http://companyshed.com/downloads/weatherdata/Italy-Bolzano.wea

http://companyshed.com/downloads/weatherdata/Italy-Cagliari.wea

http://companyshed.com/downloads/weatherdata/Italy-Crotone.wea

http://companyshed.com/downloads/weatherdata/Italy-Firenze.wea

http://companyshed.com/downloads/weatherdata/Italy-Foggia.wea

http://companyshed.com/downloads/weatherdata/Italy-Genova.wea

http://companyshed.com/downloads/weatherdata/Italy-Milano.wea

http://companyshed.com/downloads/weatherdata/Italy-MonteTerminillo.wea

http://companyshed.com/downloads/weatherdata/Italy-Rome.wea

http://companyshed.com/downloads/weatherdata/Italy-Trapani.wea

http://companyshed.com/downloads/weatherdata/Italy-Venice.wea

http://companyshed.com/downloads/weatherdata/Kenya-Nairobi.wea

http://companyshed.com/downloads/weatherdata/Malaysia.zip

http://companyshed.com/downloads/weatherdata/Malaysia-CameronHighlands.wea

http://companyshed.com/downloads/weatherdata/Malaysia-KotaKinabulu.wea

http://companyshed.com/downloads/weatherdata/Malaysia-KualaLumpur.wea

http://companyshed.com/downloads/weatherdata/Malaysia-Kuantan.wea

http://companyshed.com/downloads/weatherdata/Malaysia-Kuching.wea


122

Malaysia - Subang 19 May '02

Netherlands - De Bilt 19 May '02

Netherlands - Eelde 19 May '02

Netherlands Vlissingen 19 May '02

O Donated/Updated Date

Oman - Musqat 19 May '02

P Donated/Updated Date

Pakistan - Karachi 19 May '02

Poland - Warsaw 19 May '02

R Donated/Updated Date

Russia - Moscow 19 May '02

Russia - St Petersburg 19 May '02

S Donated/Updated

Saudi Arabia - Riyadh 19 May '02

Singapore - Singapore 19 May '02

South Korea - Seoul 19 May '02

Spain - Palma de Mallorca 19 May '02

Sweden - Goteborg 19 May '02

Switzerland - Berne 19 May '02

Switzerland - Lausanne 19 May '02

U Donated/Updated Date

UK - All (.ZIP 1,908Kb) 4 March '02

UK - Aberdeen Scotland 19 May '02

UK - Aberporth Wales - 1 19 May '02

UK - Aberporth Wales - 2 19 May '02

UK - Aldergrove Northern Ireland 19 May '02

UK - Birmingham England 19 May '02

UK - Brighton England 19 May '02

UK - Bristol England 19 May '02

UK - Camborne England 19 May '02

UK - Cambridge England 19 May '02

UK - Cardiff Wales Dr. Ian Knight 23 Oct '06

UK - Cornwall England 19 May '02

http://companyshed.com/downloads/weatherdata/Malaysia-Subang.wea

http://companyshed.com/downloads/weatherdata/Netherlands-DeBilt.wea

http://companyshed.com/downloads/weatherdata/Netherlands-Eelde.wea

http://companyshed.com/downloads/weatherdata/NetherlandsVlissingen.wea

http://companyshed.com/downloads/weatherdata/Oman-Musqat.wea

http://companyshed.com/downloads/weatherdata/Pakistan-Karachi.wea

http://companyshed.com/downloads/weatherdata/Poland-Warsaw.wea

http://companyshed.com/downloads/weatherdata/Russia-Moscow.wea

http://companyshed.com/downloads/weatherdata/Russia-StPetersburg.wea

http://companyshed.com/downloads/weatherdata/SaudiArabia-Riyadh.wea

http://companyshed.com/downloads/weatherdata/Singapore-Singapore.wea

http://companyshed.com/downloads/weatherdata/SouthKorea-Seoul.wea

http://companyshed.com/downloads/weatherdata/Spain-PalmadeMallorca.wea

http://companyshed.com/downloads/weatherdata/Sweden-Goteborg.wea

http://companyshed.com/downloads/weatherdata/Switzerland-Berne.wea

http://companyshed.com/downloads/weatherdata/Switzerland-Lausanne.wea

http://companyshed.com/downloads/weatherdata/UK.zip

http://companyshed.com/downloads/weatherdata/UK-AberdeenS.wea

http://companyshed.com/downloads/weatherdata/UK-AberporthW-1.wea

http://companyshed.com/downloads/weatherdata/UK-AberporthW-2.wea

http://companyshed.com/downloads/weatherdata/UK-AldergroveNI.wea

http://companyshed.com/downloads/weatherdata/UK-BirminghamE.wea

http://companyshed.com/downloads/weatherdata/UK-BrightonE.wea

http://companyshed.com/downloads/weatherdata/UK-BristolE.wea

http://companyshed.com/downloads/weatherdata/UK-CamborneE.wea

http://companyshed.com/downloads/weatherdata/UK-CambridgeE.wea

http://companyshed.com/downloads/weatherdata/UK-CardiffW.wea

http://companyshed.com/downloads/weatherdata/UK-CornwallE.wea


123

UK - Dundee Scotland 19 May '02

UK - Edinburgh Scotland 19 May '02

UK - Eskdalemuir Scotland - 1 6 August '02

UK - Eskdalemuir Scotland - 2 6 August '02

UK - Exeter England 4 March '02

UK - Glasgow Scotland 19 May '02

UK - Heathrow England 19 May '02

UK - Kew England - 1 19 May '02




UK - Lancashire England 19 May '02

UK - Lerwick Scotland 19 May '02

UK - London England 19 May '02

UK - Manchester England 19 May '02

UK - Newcastle England 19 May '02

UK - Norwich England 19 May '02

UK - Sheffield England 19 May '02

UK - York England 19 May '02

USA - All (.ZIP 1,422Kb) 14 June '05

USA - Achorage Alaska 19 May '02

USA - Atlanta Georgia 19 May '02

USA - Boulder Colorado 19 May '02

USA - Dallas Texas 14 June '05

USA - Denver Colorado 19 May '02

USA - Detroit Michigan 19 May '02

USA - Honolulu Hawaii Olivier Pennetier 4 June '03

USA - Houston Texas 19 May '02

USA - Las Vegas Nevada 19 May '02

USA - Little Rock Arkansas 19 May '02

USA - Los Angeles (LAX Airport) 18 August '03

USA - Medford Oregon 19 May '02

USA - Miami Florida 19 May '02

http://companyshed.com/downloads/weatherdata/UK-DundeeS.wea

http://companyshed.com/downloads/weatherdata/UK-EdinburghS.wea

http://companyshed.com/downloads/weatherdata/UK-EskdalemuirS-1.wea

http://companyshed.com/downloads/weatherdata/UK-EskdalemuirS-2.wea

http://companyshed.com/downloads/weatherdata/UK-ExeterE.wea

http://companyshed.com/downloads/weatherdata/UK-GlasgowS.wea

http://companyshed.com/downloads/weatherdata/UK-HeathrowE.wea

http://companyshed.com/downloads/weatherdata/UK-KewE-1.wea




http://companyshed.com/downloads/weatherdata/UK-LancashireE.wea

http://companyshed.com/downloads/weatherdata/UK-LerwickS.wea

http://companyshed.com/downloads/weatherdata/UK-LondonE.wea

http://companyshed.com/downloads/weatherdata/UK-ManchesterE.wea

http://companyshed.com/downloads/weatherdata/UK-NewcastleE.wea

http://companyshed.com/downloads/weatherdata/UK-NorwichE.wea

http://companyshed.com/downloads/weatherdata/UK-SheffieldE.wea

http://companyshed.com/downloads/weatherdata/UK-YorkE.wea

http://companyshed.com/downloads/weatherdata/USA.zip

http://companyshed.com/downloads/weatherdata/USA-AchorageAlaska.wea

http://companyshed.com/downloads/weatherdata/USA-AtlantaGeorgia.wea

http://companyshed.com/downloads/weatherdata/USA-BoulderColorado.wea

http://companyshed.com/downloads/weatherdata/USA-DallasTexas.wea

http://companyshed.com/downloads/weatherdata/USA-DenverColorado.wea

http://companyshed.com/downloads/weatherdata/USA-DetroitMichigan.wea

http://companyshed.com/downloads/weatherdata/USA-HonoluluHawaii.wea

http://companyshed.com/downloads/weatherdata/USA-HoustonTexas.wea

http://companyshed.com/downloads/weatherdata/USA-LasVegasNevada.wea

http://companyshed.com/downloads/weatherdata/USA-LittleRockArkansas.wea

http://companyshed.com/downloads/weatherdata/USA-LosAngelesCalifornia.wea

http://companyshed.com/downloads/weatherdata/USA-MedfordOregon.wea

http://companyshed.com/downloads/weatherdata/USA-MiamiFlorida.wea


124

USA - Nashville Tenessee 19 May '02

USA - New Orleans Lousiana 19 May '02

USA - New York New York 19 May '02

USA - Phoenix Arizona 12 May '03

USA - Salt Lake City Utah 19 May '02

USA - San Francisco California 19 May '02

USA - Seattle Washington - 1 19 May '02

USA - Seattle Washington - 2 19 May '02

http://companyshed.com/downloads/weatherdata/USA-NashvilleTenessee.wea

http://companyshed.com/downloads/weatherdata/USA-NewOrleansLousiana.wea

http://companyshed.com/downloads/weatherdata/USA-NewYorkNewYork.wea

http://companyshed.com/downloads/weatherdata/USA-PhoenixArizona.wea

http://companyshed.com/downloads/weatherdata/USA-SaltLakeCityUtah.wea

http://companyshed.com/downloads/weatherdata/USA-SanFranciscoCalifornia.wea

http://companyshed.com/downloads/weatherdata/USA-SeattleWashington-1.wea

http://companyshed.com/downloads/weatherdata/USA-SeattleWashington-2.wea

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