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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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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© Copyright 2009
By
Zhipeng ZHOU(Joe ZOE)
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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.
Zhipeng ZHOU(Joe ZOE)
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
4.9 Concluding Remarks .................................................................................... 63
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
5.6 Concluding Remarks .................................................................................... 88
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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,
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
Performance Analysis of Hybrid Liquid Desiccant Solar 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,
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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‖
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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).
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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).
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar 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).
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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%).
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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).
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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).
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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:
Performance Analysis of Hybrid Liquid Desiccant Solar 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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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).
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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:
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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:
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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)
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
63
Figure 4-13 Hybrid Cooling system COPsys with Solar Energy and without Solar
Energy
4.9 Concluding Remarks
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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,
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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 )
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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:
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
5.6 Concluding Remarks
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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)
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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)
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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)
Absolute Humidity (7g/kg) Line
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)
Absolute Humidity (7g/kg) Line
Air Temp (27 Degree) Line
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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)
Absolute Humidity (7g/kg) Line
Air Temp (27 Degree) Line
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
96
(a) Brisbane Whole Year Relative Humidity (Daily Trend)
(b) Brisbane Whole Year Relative Humidity (Weekly Trend)
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
97
(c) Brisbane Whole Year Dry Bulb Temperature (Daily Trend)
(d) Brisbane Whole Year Dry Bulb Temperature (Weekly Trend)
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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:
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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)
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
6.5 Concluding Remarks
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
107
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Subramanyam, N., Maiya, M. P. and Murthy, S. S. (2004) Parametric studies on a
desiccant assisted air-conditioner. APPLIED THERMAL ENGINEERING, 24, 2679-
2688.
Tashtoush, B., M. Molhim and M. Al-Rousan (2005) Dynamic model of an HVAC
system for control analysis. Energy, 30(10): 1729-1745.
Trott, A. R. and Welch, T. (2000) Refrigeration and Air-Conditioning. Great Britain:
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Wang, S. K. (2000) HANDBOOK OF AIR CONDITIONING AND REFRIGERATION
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Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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9 APPENDICES
Appendix 1
Daily Weather Observations for Brisbane, Queensland for December 2005/January 2006
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
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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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
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
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)
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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)
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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.
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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 - Darwin NT - 2 19 May '02
Australia - Darwin NT - 3 19 May '02
Australia - Darwin NT - 4 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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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 - Perth WA - 2 19 May '02
Australia - Perth WA - 3 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 - Sydney NSW - 2 19 May '02
Australia - Sydney NSW - 3 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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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 - Kew England - 2 19 May '02
UK - Kew England - 3 19 May '02
UK - Kew England - 4 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
Performance Analysis of Hybrid Liquid Desiccant Solar Cooling System
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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