An Absorption Chiller in a Micro BCHP Application - Carnegie ...

An Absorption Chiller in a Micro BCHP Application:
Model based Design and Performance Analysis
Hongxi Yin
Carnegie Mellon University
School of Architecture
Ph.D. Committee
Prof. Volker Hartkopf, Ph.D. (Chair)
Prof. David Archer, Ph.D.
Prof. David Claridge, Ph.D.

Copyright Declaration
I hereby declare that I am the sole author of this thesis.
I authorize Carnegie Mellon University, Pittsburgh, Pennsylvania to lend this thesis to other
institutions or individuals for the purpose of scholarly research.
I authorize Carnegie Mellon University, Pittsburgh, Pennsylvania to reproduce this thesis by photo
copying or by other means, in total or in part, at the request of other institutions or individuals for the
purpose of scholarly research.
Copyright © 2006 by Hongxi Yin
i

Acknowledgment
It has been a long journey to complete my Ph.D. thesis with the objective of making myself more
capable of dealing with the increasing complexity of building-related technical issues. The scientific
research in the Intelligent Workplace (IW) starts my academic career and a brand new professional
practice. In the future, I shall see myself as an “engineered architect”, who could help the building
industry create healthy, efficient, and economical and ultimately sustainable environments.
I wish to express my sincere appreciation and gratitude to my advisor, Professor Volker Hartkopf, for
his invaluable vision, support, and encouragement. His enthusiasm and inspiration were essential to
the success of this research, and his wisdom and insights will serve as a source of ideas for my future
endeavors.
Let me extend my profound gratitude to Professor David Archer who has played a pivotal role in this
thesis. He has far exceeded his duty as an advisor, a loyal colleagues and an enthusiastic partner in this
endeavor. Furthermore, and more importantly, he has given me a deep understanding of building
energy systems, and has also implanted his rigorous method of thinking and effective way of working.
I would like to thank Mr. Zhang Yue, CEO of Broad Air Conditioning Co., and his colleagues for their
generous support, diligent work, and warm cooperation over the past several years. Mr. Zhang Yue
spent much time on the design, test, and commercialization of this chiller. His strong motivation and
ability to convert scientific research into commercial products is one of the essential lessons he taught
me.
It gives me great pleasure to thank Professor David Claridge of Texas A&M University for providing
valuable suggestions and clarifications and Professor Richard Christensen of Ohio State University for
his careful review of the draft and his constructive critique of this work.
I also voice my appreciation to Nancy G. Berkowitz for her diligent guidance on writing skills and
editing efforts. Above all are these life-long experiences that are important for my future endeavors. I
am indebted to my colleague and lovely wife, Ming Qu, who gave me unconditional support and took
the responsibility for caring for our baby, Ryan, who fills us with joys every day. This thesis is also
dedicated to my parents in their confidence, their high expectations, and their hearty blessing.
ii

An Absorption Chiller in a Micro BCHP Application:
Model based Design and Performance Analysis
iii

Table of Contents
Copyright © 2006 by Hongxi Yin.........................................................................................................i
Acknowledgment.................................................................................................................................ii
List of Figures...................................................................................................................................viii
List of Tables........................................................................................................................................x
Abstract...............................................................................................................................................xi
1 Introduction ....................................................................................................................... 1
1.1 Background and Motivation .................................................................................................2
1.1.1 CHP Systems ....................................................................................................................3
1.1.2 BCHP Systems..................................................................................................................3
1.1.3 Heat Utilization.................................................................................................................4
1.2 Overview of Absorption Chiller Technology........................................................................5
1.2.1 Absorption Cycle Analysis ...............................................................................................6
1.2.2 Absorption Refrigeration Working Fluids ........................................................................8
1.2.3 Absorption Refrigeration Operating Conditions...............................................................9
1.2.4 Absorption Chiller Cycle Modifications...........................................................................9
1.3 Research Objectives............................................................................................................ 11
1.4 Research Approach .............................................................................................................12
1.4.1 The Planning and Installation of Experimental Equipment ............................................12
1.4.2 The Test Program and Experimental Data ......................................................................13
1.4.3 The Development of Computational Performance Model..............................................13
1.4.4 The Analysis of the Experimental Data ..........................................................................14
1.5 Current Absorption Chiller Modeling Studies ....................................................................14
1.5.1 Absorption Chiller Modeling Approaches ......................................................................14
1.5.2 The Insufficiencies of Current Absorption Chiller Modeling Studies ............................15
1.6 The Comprehensive Performance Model and its Applications...........................................16
1.6.1 The Chiller Model Description .......................................................................................16
1.6.2 Applications of the Chiller Performance Design Model.................................................18
1.6.2.1 Preliminary Design Computations..............................................................................18
iv

1.6.2.2 Detailed Design and Performance Computations .......................................................19
1.6.3 Data analysis ...................................................................................................................20
1.7 Chapter Overview...............................................................................................................20
2 Chiller Test System and Program .................................................................................. 22
2.1 Absorption Chiller ..............................................................................................................22
2.1.1 System Descriptions........................................................................................................22
2.1.2 Evaporator and Chilled-Water Pump ..............................................................................26
2.1.3 Absorber and Solution Pump ..........................................................................................27
2.1.4 High-Temperature Regenerator.......................................................................................27
2.1.5 Low-Temperature Regenerator .......................................................................................29
2.1.6 Condenser .......................................................................................................................29
2.1.7 Heat Recovery Devices...................................................................................................30
2.1.8 Cooling Tower ................................................................................................................30
2.1.9 Vacuum System...............................................................................................................31
2.2 Absorption Chiller Test Systems ........................................................................................32
2.2.1 System Description .........................................................................................................32
2.2.1.1 Steam Supply System..................................................................................................33
2.2.1.2 Variable Cooling Load System ...................................................................................34
2.2.2 Instrumentation, Control, and Data Acquisition System.................................................35
2.2.2.1 Structure of Instrumentation Control System .............................................................35
2.2.2.2 Data Acquisition and Display .....................................................................................36
2.2.2.3 Instrumentation for the Chiller....................................................................................38
2.2.2.4 Instrumentation for the Auxiliary Systems .................................................................40
2.2.2.5 Instrumentation Calibration ........................................................................................40
2.2.3 Controls for the Chiller ...................................................................................................41
2.3 Chiller Performance and Test Program...............................................................................43
2.3.1 Chiller Testing.................................................................................................................43
2.3.2 Conduct of the Testing Program .....................................................................................45
2.4 Chiller Performance............................................................................................................45
2.4.1 Chiller Performance Calculations ...................................................................................46
2.4.2 Chiller Performance under Design Condition.................................................................47
2.4.3 Chiller Performance at Reduced Capacity Condition .....................................................51
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2.5 Further Information from Chiller Testing ...........................................................................54
3 Chiller Design and Performance Model........................................................................ 55
3.1 Flow Diagram .....................................................................................................................55
3.2 Dűhring Chart Representation ............................................................................................57
3.3 T-Q Diagram.......................................................................................................................59
3.4 Calculation Procedure.........................................................................................................59
3.4.1 Mass Balance ..................................................................................................................60
3.4.2 Energy Balance ...............................................................................................................60
3.4.3 Thermodynamic Property and Equilibrium Relations ....................................................61
3.4.4 Heat Transfer Models......................................................................................................61
3.4.5 Overall Heat Transfer Coefficient Model .......................................................................62
3.4.6 Mass Transfer Models.....................................................................................................65
3.4.7 Model Assumptions ........................................................................................................65
3.5 Model Steps ........................................................................................................................66
4 Model-based Experimental Data Analysis.................................................................... 69
4.1 Analytical Method ..............................................................................................................69
4.1.1 Statistical Analysis Procedure.........................................................................................70
4.1.2 Absorption Cycle at Design Condition ...........................................................................72
4.1.3 Overall Deviation............................................................................................................74
4.2 Model Analysis ...................................................................................................................75
4.2.1 Analysis of Cooling-Load Variation ...............................................................................75
4.2.2 Performance Curve .........................................................................................................77
4.2.3 Flow Rate Variations.......................................................................................................79
4.2.4 Temperature Variations ...................................................................................................81
4.2.5 Composition Variations...................................................................................................82
4.2.6 Vapor Quality Variations.................................................................................................83
4.2.7 Heat Transfer Area Variations.........................................................................................84
4.2.8 Deviation Variations........................................................................................................85
4.2.9 Analysis of Other Test Data ............................................................................................86
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5 Contributions and Areas of Future Research............................................................... 87
5.1 Contributions ......................................................................................................................87
5.2 Areas of Future Research....................................................................................................89
5.2.1 Extended Chiller Model for Multi-Heat Resources ........................................................89
5.2.1.1 Hot Water Absorption Chiller .....................................................................................90
5.2.1.2 Natural Gas Absorption Chiller...................................................................................90
5.2.1.3 Exhaust Gas Absorption Chiller..................................................................................91
5.2.2 System Integration and Application................................................................................91
5.2.2.1 Chiller Performance Tables for Building Simulation Tools........................................92
5.2.2.2 Cost Model..................................................................................................................92
References ............................................................................................................................... 93
Appendix 1A ........................................................................................................................... 97
Appendix 2A ......................................................................................................................... 102
Appendix 2B ..........................................................................................................................118
Appendix 3A ......................................................................................................................... 130
Appendix 4A ......................................................................................................................... 150
Acronyms .............................................................................................................................. 194
vii

List of Figures
Figure 1-1: Gross estimation of annual rejected heat in the U.S., 2004 ..................................................2
Figure 1-2: Conceptual Diagram for System Integration in Buildings....................................................3
Figure 1-3: Schematic diagram of BCHP systems...................................................................................4
Figure 1-4: Basic vapor compression chiller cycle..................................................................................7
Figure 1-5: Basic LiBr absorption chiller cycle.......................................................................................7
Figure 1-6: Typical two-stage parallel flow absorption chiller configuration .......................................10
Figure 1-7: Typical two-stage series flow absorption chiller configuration .......................................... 11
Figure 2-1: Absorption chiller installed in the IW.................................................................................22
Figure 2-2: Schematic diagram of the absorption chiller.......................................................................23
Figure 2-3: Structure of the absorption chiller.......................................................................................25
Figure 2-4: Configuration of the lower vessel .......................................................................................26
Figure 2-5: Configuration of the upper vessel .......................................................................................28
Figure 2-6: Configuration of cooling tower...........................................................................................31
Figure 2-7: Simplified flow diagram of the chiller test system .............................................................33
Figure 2-8: Site views of the absorption chiller test system ..................................................................34
Figure 2-9: Control and instrumentation structure.................................................................................36
Figure 2-10: Absorption chiller monitoring software ............................................................................37
Figure 2-11: Test system monitoring software.......................................................................................38
Figure 2-12: PI&D diagram of the absorption chiller............................................................................39
Figure 2-13: Typical start-up of the chiller test system .........................................................................47
Figure 2-14: Steady-state operation of the chiller under design load condition ....................................48
Figure 2-15: Steady-state operation of the chiller under design load condition ....................................49
Figure 2-16: Chiller performance under various load conditions..........................................................53
Figure 2-17: Chiller power consumption under various load conditions...............................................53
Figure 2-18: Comparison of chiller performance ..................................................................................54
Figure 3-1: Simplified flow diagram for chiller model .........................................................................56
Figure 3-2: Dűhring chart at design condition.......................................................................................58
Figure 3-3: T-Q diagram for the heat transfer components....................................................................59
Figure 3-4: Steps in the use of the performance model .........................................................................67
Figure 3-5: Structure of the design model .............................................................................................68
Figure 3-6: Structure of performance model .........................................................................................68
viii

Figure 4-1: Data analytical procedure flow diagram .............................................................................70
Figure 4-2: Absorption cycle at design load condition ..........................................................................73
Figure 4-3: Dűhring chart at 55% design load condition.......................................................................76
Figure 4-4: Absorption cycle variations with load changes...................................................................77
Figure 4-5: Chiller performance curve under various load conditions ..................................................78
Figure 4-6: Heat transfer load on each component under various load conditions................................78
Figure 4-7: Steam supply flow rate under various load conditions .......................................................79
Figure 4-8: Sorbent solution flow rate under various load conditions...................................................80
Figure 4-9: Sorbent solution split ratio under various load conditions..................................................80
Figure 4-10: Refrigerant regeneration rate under various load conditions ............................................81
Figure 4-11: Refrigerant vaporization temperature under various load conditions ...............................82
Figure 4-12: Sorbent solution composition changes under various load conditions .............................82
Figure 4-13: Refrigerant vapor quality leaving the LTRG under various load conditions ....................83
Figure 4-14: UA changed for the 5 major components under various load conditions .........................84
Figure 4-15: Surface contact area changes under various load conditions............................................85
Figure 4-16: Overall and weighted deviations under various load conditions ......................................86
Figure 5-1: Simplified HTRG configurations for natural-gas-driven absorption chiller.......................91
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List of Tables
Table 1-1: Power generation equipment rejected heat temperature ranges..............................................5
Table 1-2: Water-LiBr absorption chiller thermal energy types and temperature ranges ........................5
Table 2-1: Component names and corresponding abbreviations ...........................................................23
Table 2-2: Specifications of the absorption chiller ................................................................................25
Table 2-4: Control points of the chiller..................................................................................................41
Table 2-3: Instrumentations of the chiller test systems..........................................................................42
Table 2-5: Input and primary output of the test program.......................................................................45
Table 2-6: Measurement data of the chiller under design condition......................................................50
Table 2-7: Comparison of chiller performance under design conditions...............................................51
Table 2-8: Primary measurement for chiller input and output...............................................................52
Table 3-1: Chiller model state point descriptions ..................................................................................57
Table 3-2: Physical features of heat and mass transfer components......................................................63
Table 3-3: Heat and mass transfer correlations used in the performance model ...................................64
Table 4-1: Measured values and model calculations for 100% and 55% of design load conditions .....71
Table 5-1: Heat transfer features of the HTRG of different heating media ...........................................90
x

Abstract
Developments in absorption cooling technology present an opportunity to achieve significant
improvements in microscale building cooling, heating, and power (BCHP) systems for residential and
light commercial buildings that are effective, energy efficient, and economic. However, model based
design and performance analysis methods for micro scale absorption chillers and their applications
have not been fully developed; particularly considering that thermal energy from a wide variety of
sources might be used to drive the chiller in a residential or light commercial building. This thesis
contributes important knowledge and methods for designing and integrating absorption chillers in
BCHP systems that reduce energy consumption, decrease operational costs, and improve
environmental benefits in residential and light commercial buildings.
To be more specific, this thesis contributes the development and application of absorption chiller and
the computational model in the following areas:
1) establishment of a unique experimental environment and procedures for absorption chiller
tests under various conditions
2) conduct of a comprehensive testing program on a microscale absorption chiller
3) construction of a comprehensive chiller model based on the pertinent scientific and
engineering principles adapted to the design of a chiller and to the analysis of extensive,
detailed test data obtained from the test program
4) analysis of the measured data, refinement of the model, and improvement of the chiller design
on the basis of the data analysis process
The model is now being used as a tool to adapt the chiller to various heat sources and sinks and to
carry out performance simulations of micro BCHP system.
xi

1 Introduction
In the United States, residential and commercial buildings – more than 107 million households (2001)
[1] and 71.7 billion square feet of commercial floor space (2003) [2] – account for more than one-third
of the total energy consumption of the country. Significant energy efficiency improvements in heating,
ventilation, air conditioning and refrigeration (HVAC&R) systems for residential and light
commercial buildings might be achieved by the application of microscale heat-driven absorption
chillers for space and ventilation air cooling.
Absorption chillers are key components in a building cooling heating and power (BCHP) system to
cool space in buildings. They can be driven directly by the thermal energy and heat recovered from
various sources, including power generation equipment and solar receiving devices. The combination
of heat recovery equipment and heat-driven absorption chillers provides significantly increased overall
energy efficiency. Most of today’s heating and cooling technologies for buildings, however, are not
designed to make use of rejected heat. Performance modeling studies of heat-driven absorption chillers
are accordingly limited, contributing to the difficulty of preparing and applying building simulation
programs for BCHP system design and performance analysis.
This thesis contributes important knowledge and methods for designing and integrating absorption
chillers in BCHP systems that reduce energy consumption, decrease operational costs, and improve
environmental benefits in residential and light commercial buildings.
The gap between experiment and simulation is closed in this thesis because of the availability of a
unique microscale absorption chiller and an associated experimental setup. By developing and
applying a numerical performance model, a refined understanding of a particular chiller and its
operation can provide improved design and modeling tools for heat-driven absorption chillers in
general. The approach developed in this thesis will allow developers to simulate the interaction of the
BCHP components as a system along with its interactions with:
• power and other energy supply systems
• electricity grids
• indoor air conditions
• various load profiles
1

The modeling tool will also allow engineers to assess different operating strategies of such a system to
find the most economic operating conditions, based on the idealized nonlinear systems with only a few
degrees of freedom.
1.1 Background and Motivation
In the United States, approximately two-thirds of the energy of the fuel used to generate electricity is
wasted as rejected heat. Annually, 28.8 to 34.0 quadrillion Btu of thermal energy are rejected to the
atmosphere, lakes, and rivers from power generation, building equipment operations, and industrial
processes, Figure 1-1, [3, 4].
Figure 1-1: Gross estimation of annual rejected heat in the U.S., 2004
National
total energy
consumption
(99.74 Quads)
Power
generation
(40.77 Quads)
Residential
sector
(6.92 Quads)
Commercial
sector
(4.02 Quads)
Industrial
sector
(21.18 Quads)
Transportation
(27.79 Quads)
Electricity
(14.2 Quads)
HVAC, lighting, and others
(4.84-5.54 Quads)
HVAC, lighting, and others
(2.81-3.21 Quads)
Manufacturing processes
(16.94-19.06 Quads)
Power production
waste heat
(24.5 -26.5 Quads)
Residential sector
waste heat
(1.38-2.08 Quads)
Commercial sector
waste heat
(0.8-1.2 Quads)
Industrial sector
waste heat
(2.12-4.24 Quads)
National total
waste heat
(28.8-34.02 Quads)
Rejected heat from power generation can be used for building operations. Renewable energy sources
(such as solar thermal energy to drive absorption chillers and boilers) combined with advanced
distributed electric energy generation can also be used in buildings. Figure 1-2 illustrates the system
integration concepts that Volker Hartkopf put forward for the first time [5], for the opportunities of
simultaneously achieving energy conservation, using renewable resource, and deploying distributed
electricity generation technologies. The building of the future is conceived as a power plant (BAPP)
that would generate more energy on site than is brought to it in the form of non-renewable resources.
The surplus of energy (power, heating, and cooling) could export to the utility grids or neighboring
buildings.
2

Figure 1-2: Conceptual Diagram for System Integration in Buildings
Resource conservation:
Energy, water,
material, and so forth
Source: Volker Hartkopf [5]
1.1.1 CHP Systems
Renewables: solar, wind, bio-gas,
day-lighting, natural ventilation,
passive/active heating/cooling
System
Integration
Distributed generation:
engine generator,
gas turbine, and fuel cell
Combined heating and power (CHP) systems are based on the concept of producing electrical energy
and recovering rejected heat for useful purposes. Compared with conventional power plants, CHP
systems can improve overall energy efficiency from 30% to 70% or more. CHP is effective in largescale
industrial plants, hospitals, university campuses, and urban district energy systems. Recent
developments in small-scale power generation, heat recovery, and heat-driven refrigeration
technologies make possible the installation and effective operation of CHP in residential and small
commercial applications.
1.1.2 BCHP Systems
In BCHP systems, the electrical energy generated on site is used to meet the demands of lighting and
electrical equipment. The rejected heat in power generation is used to provide space ventilation,
cooling, heating, dehumidification, and domestic hot water for the building, Figure 1-3.
Various technologies can be used to configure a BCHP system. The power generation equipment, as
illustrated at the top of the figure, could be a steam turbine, combustion turbine, reciprocating spark
ignition, Diesel engine, or fuel cell. These power generators produce power and reject heat in various
quantities at various temperatures that can be used for the building operation. Heat recovery
exchangers/boilers, absorption chillers, and desiccant dehumidifiers are equipment that can deliver
heating, cooling, or ventilation to the building space. As indicated in Figure 1-3, the thermal input can
also be provided directly from solar thermal receivers. Finally, a capable, robust control system is
needed to integrate the operation of all equipment to meet the needs of the building and its occupants
3

and to achieve the full benefits of system efficiency and economy. Heat-driven absorption chiller
technology plays a prominent role in making use of the reject as well as solar energy, for space and
ventilation air cooling, and thus in the design and operation of overall BCHP systems.
Figure 1-3: Schematic diagram of BCHP systems
Traditionally, CHP systems with power generation capacities below 500 kW are categorized as
microscale systems. With the development of compact, microscale absorption chillers, more reliable,
lower-emitting reciprocating engines, and high-temperature fuel cell power supplies, BCHP is feasible
for packaged systems in residential and light commercial buildings having power requirements less
than 15 kW. This introduction of micro-BCHP systems presents many technical and commercial
challenges, but the production of heat-driven absorption chillers and their integration in BCHP
systems can assist the nation in
• increasing energy efficiency
• integrating renewable forms of energy
• eliminating transmission and distribution costs and losses
• increasing reliability by combining distributed with centralized utility power supplies
1.1.3 Heat Utilization
Table 1-1 illustrates the temperature range of rejected thermal energy from typical power generators
and heat recovery units. Among them, a solid oxide fuel cell (SOFC) gives the highest exhaust gas
4

temperature for heat recovery and utilization. The hot water temperature from solar collectors varies
with the type of collector. Solar collectors with parabolic trough reflectors can generate hot water up
to 180 o C; integrated compound parabolic collectors, ICPC’s, 140 to 160 o C; flat plate collectors, 65 to
90 o C.
Table 1-1: Power generation equipment rejected heat temperature ranges
No. Power Generation Equipment, Waste Stream Temperature ( o F) Temperature ( o C)
1 Solid Oxide Fuel Cell Exhaust 1300 700-800
2 Reciprocal Engine Exhaust 1100-1200 600-650
3 Molten Carbonate Fuel Cell Exhaust 1100 600
4 Gas Turbine Exhaust 950-1000 510-540
5 Microturbine Exhaust 450-600 230-315
6 HRSG Exhaust 350 175
7 Reciprocal Engine Jacket Water 180-200 80-95
8 Phosphoric Acid Fuel Cell 180 80
9 Solar Thermal Collector 150-250 65-180
Table 1-2 shows typical temperature ranges for the heating medium to drive a water-lithium bromide
(LiBr) absorption chiller [6]. A single-stage hot-water-driven chiller can use heat at a temperature as
low as 75 o C. Tables 1-1 and 1-2 show that an absorption chiller can be found to use heat from a wide
range of sources. Because of its higher thermal efficiency, this study focuses on a two-stage absorption
chiller and its appropriate sources of rejected heat.
Table 1-2: Water-LiBr absorption chiller thermal energy types and temperature ranges
No. Heat-driven Absorption Chiller Type Pressure (kPa) Temperature ( o C)
1 Direct-fired fossil fuel (natural gas, oil, LPG etc.) - 1,000 – 1,800
2 Double-stage exhaust gas - 400 - 600
3 Single-stage exhaust gas - 230 - 350
4 Double-stage steam 400 – 1,000 144 - 180
5 Single-stage steam 100 - 400 103 - 133
6 Double-stage hot water 350 – 1,100 140 - 200
7 Single-stage hot water 40 - 200 75 - 120
8 Other fuel/steam/hot water/exhaust gas Same as above Same as above
1.2 Overview of Absorption Chiller Technology
An absorption chiller is a machine that, driven by heat, produces chilled water for space and
ventilation air cooling. Little or no mechanical energy is consumed in an absorption chiller, and little
or no electric power is required. A great variety of hot media, gases and liquids, over a broad range of
temperatures above ambient can be used. The chiller must also reject an amount of heat equal to that
provided in driving it plus that absorbed in producing the chilled water. Ammonia-water (NH3-H2O)
5

absorption refrigeration technology has been used for more than 150 years. As a refrigerant, ammonia
has high latent heat and excellent heat transfer characteristics, but its toxicity has limited its use in this
technology.
Since 1945, water-LiBr absorption chillers have achieved widespread use. This trend reached its peak
in the 1960s, and then diminished in the late 1970s. The technology has since revived in Asia, because
the rapidly increasing electricity demand has limited the application of electrically driven vapor
compression chillers. The sales data of a leading absorption chiller manufacturer, presented in
appendix 1A, shows several new developments in the current absorption chiller market. Today, water-
LiBr absorption chiller technology is returning to the United States with the increasing application of
CHP systems.
In the past three years, heat-driven water-LiBr absorption chillers have been used widely both in large
commercial buildings combined with advanced power generation equipment and in individual houses
driven directly by fossil fuels or by other heat sources. The cooling capacity of chillers can vary from
greater that 1,000 refrigeration ton (3,561.85 kW) to as low as a microscale, 4.5 refrigeration ton (16
kW). This thesis will focus on microscale water-LiBr absorption chiller research, development, and
demonstration in residential and light commercial applications.
1.2.1 Absorption Cycle Analysis
A chiller produces chilled water by removing heat from it and transferring this heat to a vaporizing
refrigerant. The process is illustrated in Figure 1-4 for a conventional vapor compression chiller and
in Figure 1-5 for an absorption chiller. In both, the refrigerant liquid flows into an evaporator,
evaporates at a reduced pressure and temperature, and absorbs heat from chilled water flowing in a
tube through the evaporator. In the vapor compression process, the refrigerant vapor is compressed
and condensed at a high-pressure and temperature, transferring heat to cooling water or to the
surroundings in a condenser. The high-pressure condensed refrigerant is then returned through the
expansion valve to a low-pressure evaporator, once again to absorb heat from the chilled-water flow.
6

Figure 1-4: Basic vapor compression chiller cycle
Refrigerant
expansion valve
P
Evaporator
Heat absorbed
from chilled water
T
Heat rejected
to cooling water
Condenser
Compressor
In the absorption process shown in Figure 1-5, the refrigerant vapor from the evaporator is absorbed at
low pressure into a sorbent solution in the absorber. Heat is released as the refrigerant vapor is
absorbed. This heat is removed by cooling water flowing through the absorber. The sorbent solution
is then pumped to the regenerator, where refrigerant vapor is driven from the sorbent solution by the
addition of heat at high temperature and pressure. The refrigerant vapor is condensed at high pressure
and temperature with the removal of heat to ambient or to cooling water. The liquid refrigerant is
returned to the evaporator through the expansion valve.
Figure 1-5: Basic LiBr absorption chiller cycle
P
Refrigerant
expansion valve
Heat absorbed
from chilled water
T
Heat rejected
to cooling water
Condenser
Solution
pump
Evaporator Absorber
Heat rejected
to cooling water
Work
Heat input
Regenerator
Solution
expansion valve
7

This basic absorption chiller cycle shown in Figure 1-5 is similar to the traditional vapor compression
chiller cycle in Figure 1-4 in that
• refrigerant vapor is condensed at high pressure and temperature, rejecting heat to the
surroundings
• refrigerant vapor is vaporized at low pressure and temperature, absorbing heat from the chilled
water flow
The chiller cycles differ in that
• the pumped circulation of a sorbent solution replaces the compression of the refrigerant vapor
The energy, work, required by the pump is significantly less than that required by the
compressor
• heat must be supplied in the regenerator to release refrigerant vapor at high pressure for
condensation, and heat must be removed from the absorber
From the standpoint of thermodynamics, the vapor compression chiller is a heat pump, using
mechanical energy and work, to move heat from a low to a high temperature. An absorption chiller is
the equivalent of a heat engine – absorbing heat at a high temperature, rejecting heat at a lower
temperature, producing work – driving a heat pump.
1.2.2 Absorption Refrigeration Working Fluids
An absorption chiller requires two working fluids, a refrigerant and a sorbent solution of the
refrigerant. In a water-LiBr absorption chiller, water is the refrigerant; and water-LiBr solution, the
sorbent. In the absorption chiller cycle the water refrigerant undergoes a phase change in the
condenser and evaporator; and the sorbent solution, a change in concentration in the absorber and
evaporator.
Water is an excellent refrigerant; it has high latent heat. Its cooling effect, however, is limited to
temperatures above 0 o C because of freezing. The sorbent, LiBr, is nonvolatile, so a vapor phase in
the absorption chiller is always H2O. The sorbent solution, water-LiBr, has a low H2O vapor pressure
at the temperature of the absorber and high H2O vapor pressure at the temperature of the regenerator,
facilitating design and operation of the chiller. The advantage of the water-LiBr pair includes its
stability, safety, and high volatility ratio. It has no associated environmental hazard, ozone depletion,
or global warming potential.
8

1.2.3 Absorption Refrigeration Operating Conditions
The choice of the refrigerant, water, and sorbent, water-LiBr solution, along with the designation of a
chilled-water outlet temperature and cooling-water inlet temperature determines the operating
temperatures and pressures in the evaporator, absorber, regenerator, and condenser of the LiBr
absorption chiller as illustrated in Figure 1-5.
• In the evaporator, low operating temperature and pressure are required to vaporize refrigerant
to absorb heat from the chilled water.
• In the absorber, the cooling-water temperature determines the composition of the sorbent
solution so that it absorbs the refrigerant vapor, as required, at the pressure determined by the
evaporator.
• In the regenerator, the pressure is that of the condenser. An elevated value is required to
condense the refrigerant vapor at the temperature of the cooling water. The temperature in the
absorber is that required to vaporize the refrigerant from the sorbent solution.
The low operating pressure in the evaporator and absorber requires high equipment volume and a
special means for reducing pressure loss in the refrigerant vapor flow. Preventing the leakage of air
into the evaporator and the absorber is one of the main issues in operating an absorption chiller. A
special purge device removes air and other noncondensable gases, and an external vacuum pump is
used periodically to maintain low operating pressure. The high operating pressure in the regenerator
and condenser requires the use of heavy-walled equipment and a pump to deliver the sorbent solution
from the low-pressure absorber to the high-pressure regenerator. Crystallization, the deposition of
LiBr from the sorbent solution at high concentrations and low temperatures, can block the sorbent
flow and cause the chiller to shut down. Controls are usually necessary to prevent crystallization.
1.2.4 Absorption Chiller Cycle Modifications
Several modifications can be made in the basic absorption chiller cycle to reduce the heat required to
operate the chiller and to reduce the extent of heat transfer surface incorporated in the machine.
• Countercurrent heat interchange can be arranged between the two sorbent solution flows
connecting the low-temperature absorber and the high-temperature regenerator. This
interchange can significantly reduce the heat quantities involved in the operation of both; less
heat will need to be supplied to the regenerator, and less heat will need to be removed form the
absorber.
9

• The refrigerant vapor leaving the high-temperature and -pressure regenerator can be used to
vaporize an equal quantity of refrigerant from the sorbent solution in a second regenerator
operating at a lower temperature and pressure. This second stage of regeneration reduces the
heat requirement of the absorption chiller by a factor approaching 2.
• Heat transfer between the vaporizing refrigerant and the chilled water in the evaporator can be
facilitated by recirculating the refrigerant liquid over the heat transfer surface, reducing the
temperature difference and the heat transfer area.
Figure 1-6: Typical two-stage parallel flow absorption chiller configuration
Refrigerant
combiner
Refrigerant
expansion valve
P
Recirculation
pump
Heat absorbed
from chilled water
Heat rejected
to cooling water
T
Condenser
Solution
splitter
Solution
pump
Evaporator Absorber
Heat rejected
to cooling water
Condenser
Heat to
LTRG
Low-temp.
regenerator
Low-temp.
heat exchanger
Solution
combiner
Solution
expansionvalve
Heat input
Regenerator
High-temp.
heat exchanger
The revised flow diagrams illustrating these absorption chiller flow diagrams are shown in Figures 1-6
and 1-7. The flow of the sorbent solution from the absorber to the two regenerators can be either
parallel or in series. In a parallel flow arrangement, the dilute solution from the absorber is pumped to
both the high-temperature and the lower-temperature regenerators in parallel, as shown in Figure 1-6.
Concentrated solutions from both regenerators are recombined and returned to the absorber. In a
series flow arrangement, the solution from the absorber is first pumped to the high-temperature, highpressure
regenerator; and the partially concentrated sorbent solution then flows to the lower-pressure,
lower-temperature regenerator, as shown in Figure 1-7.
10

Figure 1-7: Typical two-stage series flow absorption chiller configuration
P
Refrigerant
combiner
Refrigerant
expansion valve
Recirculation
pump
Heat absorbed
from chilled water
Heat rejected
to cooling water
T
Condenser
Solution
Pump
Evaporator Absorber
Heat rejected
to cooling water
Condenser
Heat to
LTRG
Low-temp.
heat exchanger
High-temp.
heat exchanger
Low-temp.
regenerator
Solution
expansion valve
Heat input
Regenerator
A parallel flow configuration has several advantages over the series flow configuration. The sorbent
solution flow in each heat interchanger is only half that of the series flow configuration. In general,
the parallel configuration has a lower heat input requirement than the series flow configuration.
1.3 Research Objectives
The objective of this research is to develop methods for the effective design and evaluation of
absorption chiller-based micro-BCHP systems that reduce energy consumption, decrease operational
costs, and improve environmental benefits in residential and light commercial buildings. The methods
demonstrated in the thesis can be widely used in building energy system design and evaluation; they
can also be broadly applied in an absorption chiller and other BCHP system equipment design, and in
system integration. The analytical methods also provide the basis for diagnosing and optimizing the
operation of absorption chiller-based micro-BCHP systems.
Four research areas are involved in this work on microscale absorption chiller system evaluation and
performance simulation:
11

1) establishment of a unique experimental environment and procedures for absorption chiller
tests under various conditions
2) conduct of a comprehensive testing program on a microscale absorption chiller
3) construction of a comprehensive chiller model based on the pertinent scientific and
engineering principles adapted to the design of a chiller and to the analysis of extensive,
detailed test data obtained from the test program
4) analysis of the measured data, refinement of the model, and improvement of the chiller design
on the basis of the data analysis process
The model is now being used as a tool to adapt the chiller to various heat sources and sinks and to
carry out performance simulations of micro BCHP system. In both its theoretical and practical aspects,
this study contributes important knowledge for the development and application of micro-BCHP
systems in residential and light commercial buildings. The improvements in BCHP system analytical
methods lay the groundwork for developing of overall BCHP system performance assessment tool; the
practical progress in microscale-BCHP system experiment and evaluation setups establishes the
threshold for an efficient and integrated microscale building energy supply, distribution, and delivery
system. These contributions are made possible by close cooperation in research and development
(R&D) with a leading manufacturer; in turn, some of the research achievements of this study have
been promptly incorporated into the emerging technology and product.
1.4 Research Approach
To achieve the research objectives, this thesis focuses on equipment installation and test, model
development, data analysis, and system simulation of a microscale, steam-driven, two-stage LiBr
absorption chiller for an energy supply system in Carnegie Mellon University (CMU)’s Robert L.
Preger Intelligent Workplace (IW). Experimental data and a computational model are the two basic
components of this work. The experience gained provides the framework for other BCHP component
studies and system integration. The research has been carried out in the following several steps: some
in parallel, others sequentially:
1.4.1 The Planning and Installation of Experimental Equipment
A microscale BCHP energy supply system (ESS) has been designed for the IW, a 6,500 ft 2 office
environment at CMU, to provide power and space cooling heating, and ventilation. As the first stage
in realizing this overall system, a 16kW steam-driven water-LiBr absorption chiller was installed in
the south section of the IW. This chiller
12

• is driven by steam, reducing summer electrical peak demands and leveling the year round
demand for natural gas and other fuels
• is flexible in adapting to thermal recovery equipment associated with various prime movers
• provides a cooling capacity and compactness appropriate for residential, small commercial, and
institutional buildings
• incorporates a cooling tower to reject the heat from its operation as required
The chiller was installed together with its auxiliary steam and chilled-water supply, and test load
systems in the IW. A web-based chiller automation system (CAS) was also installed to operate the
chiller with its auxiliary systems, monitor the overall system status, and collect the experimental data.
In this test-bed the absorption chiller was also integrated into the IW and campus chilled-water system,
so when the test was over, the chiller could provide chilled water to the IW and the campus.
Experiments were carried out under a broad range of system operating parameters.
In this work, both equipment testing and mathematical model simulation of the chiller were combined
to provide a detailed understanding of the equipment, to analyze the test data, to discover possible
chiller design improvements and modifications, and to provide a method to design and evaluate
overall BCHP systems.
1.4.2 The Test Program and Experimental Data
The chiller was tested by varying six operating parameters in turn: the chilled-water return temperature
and flow rate, the cooling-water supply temperature and flow rate, and the steam pressure. In the test
program, only one parameter was adjusted at a time, and the others were kept at design conditions.
Additional sensors were installed in the chiller beyond those provided by the manufacturer to operate
the chiller and its auxiliary system to calculate chiller performance such as the coefficient of
performance (COP) and cooling capacity, and to observe chiller internal conditions. Experimental data
obtained from 11 temperature sensors in the chiller were used to verify the predictions of the
performance model.
1.4.3 The Development of Computational Performance Model
On the basis of scientific and engineering principles and the specific configurations of the chiller, a
detailed computational performance model was constructed to evaluate the chiller performance under
various operating conditions. This model was developed for the chiller to further refine the
13

understanding of the principles of the chiller, to analyze the experiment data from the test program, to
assist in the equipment design, and to evaluate the performance of BCHP systems.
The basic equation types incorporated in the model include: mass and energy balances,
thermodynamic property relations, thermal and phase equilibrium relations, and heat and mass transfer
coefficient correlations. The variables in these equations are the operating conditions – pressures,
temperatures, compositions, and flows – throughout the chiller. The model includes 416 variables and
409 equations. If seven operating conditions are specified, the model can be solved and all the
operating conditions throughout chiller can be calculated.
1.4.4 The Analysis of the Experimental Data
To assess the performance data collected, an analytical method was developed that minimizes the
deviations between the experimental measurements and the model solutions. Several model
assumptions were adjusted to improve the agreement between the experimental measurements and the
model calculations. These adjustments significantly improved the agreement between the calculated
and measured variables.
1.5 Current Absorption Chiller Modeling Studies
The microchiller performance model is one of the major efforts of this research. The literature for
absorption chiller model studies has been reviewed; the existing model studies are categorized and
summarized in the following sections.
1.5.1 Absorption Chiller Modeling Approaches
In the past decades, computer models have been developed to investigate the performance of various
water-LiBr absorption chiller cycles. Among these models, some [8, 9] are system specific for
particular machines, flow configurations, and working materials. Others [10, 11, 12] are generic to
handle various potential absorption cycles with one modularized model. The system specific models
are performance models aimed at simulating a specific design and investigating its performance under
various operation conditions; the generic models are aimed at exploring novel absorption cycles and
evaluating their performance under various boundary conditions.
The advantage of system specific or performance models is that the model simulates the configuration
of absorption chiller systems in detail. Thermodynamic cycle, heat, and mass transfer characteristics
can be investigated on the basis of the physical details of the chiller. In these studies the simulation
14

esults are verifiable through the chiller operations under various conditions. The difficulty of this type
of model is that the accurate details of chiller configuration and design are not always available from
the manufacturer. In most cases a simplified approach is adopted to solve the models, such as a
specified heat transfer coefficient of specific chiller components provided by the manufacturer.
The advantage of the generic cycle model is that programming effort is reduced by modular structure.
A generic model is normally developed on the basis of the thermodynamic theory to investigate the
performance of different absorption cycles and working fluids. This type of model is used in the
conceptual design of an absorption machine. It can be used effectively to predict the performance of
different design configurations, but because of its generic characteristics, it is difficult to investigate
the details of the physical configuration of the chiller and its components.
Beyond absorption cycle simulations, modeling efforts [13, 14, 15, 16, 17] focus mainly on chiller
component design. Numerous modeling studies and experimental efforts have been made on combined
heat and mass transfer, working fluid additives, noncondensable gas measures, and other features of
absorption chillers. These studies have advanced the capability for modeling absorption chillers. Some
simulation results were found to be in good agreement with the experiments. On the basis of the
experiments, some empirical correlations for combined heat and mass transfer have been proposed for
several typical absorber configurations and working fluids. The methods and results of these prior
studies have been applied in the modeling efforts of this thesis.
1.5.2 The Insufficiencies of Current Absorption Chiller Modeling Studies
First, the existing simulation models of water-LiBr absorption chillers focus on relatively large-scale
installations for commercial buildings or for district energy centers. None of the studies consider
microscale absorption chillers with a cooling capacity less than 17 kW for residential or light
commercial applications. There are, theoretically, no distinctions between the large-scale and the
microscale absorption chillers in terms of scientific and engineering principles, but the design criteria
and operating conditions for microscale absorption chillers are different from those for the large
capacity chillers. For instance, microscale absorption chillers for residential application must provide a
more compact design and include a heat rejection unit, such as a cooling tower.
Second, at present, nearly all performance models of absorption chillers have been numerical
simulations without significant experimental validation under design and off-design conditions. It has
been difficult to install a commercial absorption chiller in a university laboratory because of their large
capacity. The requirements for operation and test of commercial chillers and their limited
15

instrumentations greatly restrict their accessibility for the experiments. The small cooling capacity of a
microscale chiller, however, makes it possible to provide a test cooling load and to simulate a wide
range of operation conditions for the chiller.
Third, the model validation method has been simplified in the past studies. The deviations between
the experimental and the performance simulation results for the COP and the cooling capacity at a
single given operational condition are used to judge the overall quality of the model.
Finally, the available packaged absorption chiller models lack the flexibility to be integrated into
building simulation tools to support the design and analysis of absorption chiller-based BCHP systems.
The work reported in this thesis addresses these insufficiencies.
1.6 The Comprehensive Performance Model and its Applications
In this work, a steady-state performance model has been developed for the Broad BCT16 absorption
chiller to further refine the understanding of the principles of this chiller, to analyze the experiment
data from the test program, to assist in the equipment design, and to evaluate the performance of
BCHP systems.
1.6.1 The Chiller Model Description
In the model, the absorption chiller is composed of the following components:
• an evaporator: a countercurrent two-phase coiled tube heat exchanger
• an absorber: a countercurrent two-phase coiled tube mass and heat exchanger
• two regenerators: one high temperature, one intermediate temperature: well mixed, two-phase
boiling coiled tube heat exchangers
• a condenser: a countercurrent heat exchanger
• two plate heat interchangers: countercurrent single-phase heat exchangers
• two tube and shell heat recovery exchangers: countercurrent single-phase exchangers
• three pumps: a sorbent pump, a refrigerant pump, and a chilled-water pump
• associated spray nozzles, trap, valves, and pipe fittings
The cooling tower associated with this chiller includes the following components:
• a countercurrent plate column two-phase mass and heat exchanger
16

• a cooling-water pump
• an air fan
The complete steady-state chiller model is composed of the following nonlinear algebraic equations
applicable to each of the above chiller and cooling-tower components:
• two mass balances, water and LiBr
• an energy balance
• thermodynamic property relations for stream enthalpies as a function of pressure, temperature,
and composition
• phase equilibrium relations among pressure, temperature, and compositions of the coexisting
phases
• the appropriate heat transfer (and for the absorber and cooling tower, mass transfer) relations
• correlations of overall heat and mass transfer coefficients, U and K, for the respective
components based on their specific design and operating conditions, (see chapter 3)
• work computations for the pumps and fan
These equations involve, as variables, the properties – pressure, temperature, composition, and flow –
of all the phases present in and flows among the chiller components. The completed chiller model
interrelates variables of all these equations based on the configuration and the flow diagram, of the
chiller. In general it has been assumed that:
• The properties of a stream leaving a component to an interconnected component are those of
eithera liquid or a vapor, thus the quality of the stream is either 1.0 or 0.0
• There is no pressure loss and no heat loss/gain in the lines connecting the components
• Tthe sorbent solution charged to the chiller has a concentration of 55% LiBr. Once the chiller
operates under design conditions, the concentration difference of the sorbent solutions flow in
and out of the high-temperature regenerator is roughly at 5%; that of the intermediate
temperature regenerator is approximately 4%. Dilute sorbent is distributed to the two
regenerators in approximately equal quantities.
The completed chiller model involves 416 variables and 409 nonlinear algebraic equations. Solving
the model and determining values for all the chiller variables therefore requires specifying values for
seven operating parameters. In this work, the specified operating parameters are: the chilled water
17

inlet and outlet temperatures and flow, the cooling-water supply temperature and flow, the steam
supply pressure and flow.
1.6.2 Applications of the Chiller Performance Design Model
This chiller performance model has been used in various forms for various applications: preliminary
design, detailed design, and performance data analysis by
• excluding or including various model equations
• making various assumptions relating to the model equations
• specifying various input and corresponding output variables or operating conditions
1.6.2.1 Preliminary Design Computations
The steam flow and the pump work for a given cooling load – chilled-water inlet and outlet
temperatures and flow – and the internal operating conditions throughout the chiller can conveniently
be estimated from a simplified form of the model by
• excluding heat and mass transfer relations and correlations
• fixing the composition of the circulating sorbent solution
• assuming that
o the operating temperatures (and the corresponding equilibrium pressures) of the
evaporator, absorber, high and intermediate temperature regenerators, respectively, are
those of the outlet chilled water, the inlet cooling water, the steam supply, the condensing
temperature of the refrigerant vapor from the high-temperature regenerator.
o the operating pressure of the condenser (with its corresponding pressure) is that of the
intermediate temperature regenerator.
o heat transfer in the countercurrent interchangers and heat recovery exchangers is
maximized by equal stream temperatures at one end of the exchanger.
Preliminary design computations have proved useful in exploring the effects of various chiller
configurations, component characteristics, and external operating conditions on the heating and
cooling requirements, internal conditions, and power requirements of a chiller.
A preliminary design model was programmed to estimate the heat/mass transfer areas of the chiller
components; this was a first step in constructing a comprehensive performance model. If the design
18

conditions, the desired performance, the specific configuration, and the reasonable assumptions are
incorporated in the model, the heat transfer area, UA, of chiller components can be calculated. The UA
is defined as the product of overall heat transfer coefficient (U) and the total internal contact area (A):
• The design conditions are these specified conditions (temperature, pressure, and flow) of
chilled water, cooling water, and heat sources at a specified load condition
• The performance parameters are the values of COP and cooling capacity
• The operating parameters are the conditions (temperature, pressure, and flow) of chilled water,
cooling water, and heat source at any operating conditions
• The specific chiller configuration includes the information such as one-stage or two-stage,
parallel- or series-sorbent flow (for a two-stage absorption chiller), heat source types, working
fluids, and other details of the chiller
1.6.2.2 Detailed Design and Performance Computations
On the basis of the design model, the performance model was constructed to predict chiller
performance and to analyze the measured experimental data. First, the performance model took the
initial UA estimations from the design model to predict chiller performance under design conditions;
then, the actual Us and As were calculated from the actual chiller physical configurations and from the
heat and mass transfer correlations from the literature. The heat transfer correlations were corrected by
comparing the actual Us and As and the UA solutions from the design model, and then, the corrected
UA correlations were used to predict the chiller performance for design and off-design operations.
Heat (and mass) transfer areas required in the various components of the chiller can be estimated by
the performance model by
• applying known conditions for steam flow, pump work, etc.
• including heat and mass transfer relations and correlations from the literature in the model
• fixing the composition of the circulating sorbent solution
• assuming “approach” temperatures (and pressures) for heat (and mass) transfer occurring in
each of the various chiller components.
The calculated transfer area values for the given design values of the external operating conditions –
the chilled-water inlet and outlet temperatures and flow, the steam conditions (temperature and
pressure), the cooling water inlet temperature and flow – can then be used in the model to determine
the effects of off design external operating conditions on chiller performance.
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1.6.3 Data analysis
The outputs of performance model were also compared with test data under various conditions by
changing the operation parameters. Based on the performance model, the accuracy and reliability of
the experimental data were assessed, and the model assumptions were validated.
Measured chiller external and internal operating conditions can be used to compare those calculated
from the chiller model with the results when the model is supplied with the external operating
conditions and component areas.
These comparisons can be used to evaluate the accuracy of measurements and to consider the validity
of the model including the assumptions on which the model is based. Such comparisons and
conclusions based on those comparisons are discussed in detail in chapter 3. Several measurements
that differ significantly from model predicted values have been analyzed, and the procedures for
correcting these measurements have been proposed and applied.
On the basis of the model developed in this thesis, the validated model can then be extended to
incorporate the following heat sources:
• hot water from solar thermal or heat recovery equipment
• natural gas
• exhaust gas from gas turbine, engine generator, and fuel cells
The validated models, as a tool, can be integrated with the IW model to evaluate overall BCHP system
performance incorporating with a cost model.
1.7 Chapter Overview
This thesis contains five chapters followed by references, and appendixes, and list of abbreviations.
Chapter 1, Introduction introduces the background and motivation of this dissertation and
summarizes the research objectives of this chapter. The emerging features of the modern absorption
chiller industry are summarized in appendix 1A.
Chapter 2, Chiller Test System and Performance introduces the chiller and examines the overall
experimental system setups. It presents detailed information concerning the instrumentation and
control for the chiller and its auxiliary systems. The chiller testing program, measured experimental
20

data, and chiller performance are presented. The chiller internal control principles and the system
operation instructions are presented in appendixes 2A and 2B, respectively.
Chapter 3, Computational Model describes the framework of the performance model within which
the absorption chiller component modules are developed. It provides an in-depth presentation of the
governing equations and modeling assumptions. The computational and numerical issues are
addressed in the various stages of the absorption chiller component modeling in appendix 3A; the
source code of the performance model is attached in appendix 3B.
Chapter 4, Model-based Data Analysis assesses the model calculations and experimental data
accuracy and reliability to learn how to validate the model as well as improve the equipment designs.
The analysis results presented regard the test programs that vary for five operating parameters: chilledwater
supply temperature and flow, cooling-water supply temperature and flow, and steam supply
pressure. When analyzing the experimental data, opportunities to improve the accuracy of the model
became apparent. Consequently, the adjustments to model assumptions significantly improved the
agreement between the calculated and the measured variables.
Chapter 5, Contributions and Areas for Future Research summarizes the contributions of this
thesis and suggests future areas for research and the issues involved, including: extension of the
validated steam-driven absorption chiller model to several other heat sources: hot water, natural gas,
and exhaust gases. The chiller performance models can be integrated and evaluated into overall BCHP
system configurations on an annual basis.
21

2 Chiller Test System and Program
As a first step in providing an energy supply system for CMU’s IW, a 16kW, steam-driven, two-stage
absorption chiller was installed together with an auxiliary steam supply and a variable load for the
chiller test and performance evaluation. A web-based data acquisition and control system was
developed to operate the chiller and its auxiliary equipment while storing and displaying the test
measurement data. The chiller was tested at various operating conditions in accordance with a test
program. In the future, the chiller and its control system will be incorporated in the cooling system of
the IW and connected with the campus chilled-water supply system.
2.1 Absorption Chiller
2.1.1 System Descriptions
The absorption chiller installed in the IW is a steam-driven, two-stage, water-LiBr, parallel-sorbentflow
series-cooling-water flow chiller with a cooling tower. This chiller, provided by Broad Co., has a
16kW rated cooling capacity. It is the smallest absorption chiller available in the existing market and
the only steam-driven absorption chiller of such capacity in the world.
Figure 2-1 shows the absorption chiller installed on
a platform adjacent to the IW. The chilled-water
supply and return, steam supply, condensate return,
power, and city water lines connect with the chiller
at the bottom left. Figure 2-2 is a schematic flow
diagram recreated from the manufacturer’s brochure
for a commercial natural-gas direct-fired chiller; this
flow diagram shows all the heat and mass transfer
components, pumps, and pipe fittings. It also
indicates the design values for temperatures
throughout the chiller. The measurement and
control features of the chiller will be discussed in
conjunction with a detailed process and
instrumentation (P&I) diagram in the section that
follows. The components and parts indicated in
Figure 2-2 are listed in Table 2-1.
Figure 2-1: Absorption chiller installed in the IW
22

Figure 2-2: Schematic diagram of the absorption chiller
Table 2-1: Component names and corresponding abbreviations
Abbreviation Name Abbreviation Name
ABS Absorber EVP Evaporator
BPHX By-pass heat exchanger HTRG High-temperature regenerator
CHSV Cooling/heating switch valve HRHX Heat recovery heat exchanger
CHWBPV Chilled-water by-pass valve HTHX High-temperature heat exchanger
CHWP Chilled-water pump LTHX Low-temperature heat exchanger
COND Condenser LTRG Low-temperature regenerator
CT Cooling tower RBPSV Refrigerant by-pass solenoid valve
CTOF City-water overflow RP Refrigerant pump
CTWS City-water switch RPH Refrigerant pump heater
CWBPV Cooling-water by-pass valve SF Steam filter
CWDD Cooling-water drain device SP Solution pump
CWDV Cooling-water detergent valve ST Steam trap
CTF Cooling-tower fan SV Steam valve
CWP Cooling-water pump
23

The absorption chiller in Figure 2-2 consists of five major and four minor heat transfer components,
three pumps, a cooling tower, an automatic inert gas purge device, and the associated valves and pipe
fittings. Specifically, the five major components are:
• an evaporator, a countercurrent two-phase heat exchanger
• an absorber, a countercurrent two-phase heat and mass exchanger
• a high-temperature regenerator (HTRG), a well-mixed, two-phase, boiling heat exchanger
• a low-temperature regenerator (LTRG), a well-mixed, two-phase boiling heat exchanger
• a condenser, a countercurrent heat exchanger
The four minor components are:
• a high-temperature heat interchanger (HTHX), a countercurrent, single-phase heat exchanger
• a low-temperature heat interchanger (LTHX), a countercurrent, single-phase heat exchanger
• a heat recovery heat exchanger (HRHX), a countercurrent, single-phase heat exchanger
• a refrigerant by-pass heat exchanger (BPHX), a countercurrent, single-phase heat exchanger
The three pumps are:
• a solution pump (SP), a variable-speed pump
• a chilled-water pump (CHWP), a single-speed pump
• a refrigerant pump (RP), a single-speed pump
The cooling tower (CT) includes:
• a countercurrent vertical plate column; a two-phase, mass and heat exchanger
• a cooling-water pump (CWP); a single-speed pump
• a cooling-tower fan (CTF); a three-speed air fan
• associated valves and drain devices
Other associated components include:
• an automatic gas purge device (AGPD)
• associated valves, spray nozzles, and pipe fittings
24

Figure 2-3: Structure of the absorption chiller
The physical arrangement of the absorption chiller is shown in Figure 2-3. The main body of the
chiller consists of two sealed vessels: the upper one at an elevated pressure, the lower vessel at a high
vacuum. The upper vessel includes the HTRG, the LTRG, and the condenser. The lower vessel
includes the absorber, the evaporator, the BPHX, the LTHX, and the HTHX. The flows of sorbent
solutions, refrigerant, and cooling water penetrate the vessel walls in pipes between the two vessels.
The high vacuum in the lower vessel is maintained by the AGPD and a manual vacuum pump
independent of the chiller. The chilled water and cooling water are circulated by the CHWP and the
CWP, respectively. The inclusion of the cooling tower enables chiller installation where cooling water
may be unavailable.
Table 2-2: Specifications of the absorption chiller
Chilled water
Steam
Power
Solution
Name Quantity Unit
Cooling capacity 16 kW
Chilled-water return temperature 14
o
C
Chilled-water supply temperature 7
Chilled-water flow rate 2 m 3 /h
Chilled-water pump head 8 mH2O
Rated steam pressure, absolute 0.7 mPa
Steam pressure limit, absolute 0.9 mPa
Maximum steam consumption 24 kg/h
Power voltage 220 V
Power frequency 60 Hz
Maximum power consumption 1 kW
Water-LiBr sorbent solution mass 65 Kg
Water-LiBr sorbent concentration 55 %
o C
25

Table 2-2 lists the chiller specifications from the manufacturer; these are the only published
performance data for this unique chiller. A test program was developed to investigate chiller
performance and to provide additional measurements of chiller operating conditions. The chiller
specification data are useful in evaluating the results of the chiller tests. The chiller working principles
are described in the following sections.
2.1.2 Evaporator and Chilled-Water Pump
The evaporator of the chiller, shown in Figures 2-2 and 2-4, occupies the lower vessel. The evaporator
tube bank comprises two parallel tubes spiraling 18 times from the bottom to the top of the coil. Water
refrigerant is distributed evenly over the tubes in the bank by nozzles spraying water from the
condenser. Water that was not evaporated in the first pass collects in the refrigerant tray at the base of
the evaporator and is recirculated by the refrigerant pump. The refrigerant vaporizes in the evaporator
at low pressure, about 0.8-1.0 kPa, and low temperature, about 3-4 o C. The vaporization absorbs heat
from the chilled water flowing through the evaporator coil, cooling this flow from 14 o C to 7 o C.
Figure 2-4: Configuration of the lower vessel
26

At a constant flow rate of 2 m 3 /h and a head of 8 mH2O to overcome the pressure loss, the evaporator
functions as a countercurrent, two-phase heat exchanger. The steam flow to the HTRG is adjusted to
maintain a constant refrigerant level water tray reservoir; a low level requires an increase in the steam
flow to provide more refrigerant. The chiller control system is discussed in appendix 2.A
2.1.3 Absorber and Solution Pump
The absorber, shown in Figures 2-2 and 2-4, maintains the low operating pressure required in the
evaporator. It is a spiral tube bank, consisting of two tubes spiraling from the bottom to the top. The
coil surrounds the evaporator but is separated from it by a chevron separator to prevent carryover of
refrigerant liquid. Concentrated water-LiBr sorbent solution is distributed evenly over the tubes of the
absorber coil by nozzles spraying sorbent from the two regenerators, cooled in the HTHX and the
LTHX. The water refrigerant vapor from the evaporator passes through the chevron separator, enters
the absorber, and is absorbed in the water-LiBr sorbent flowing 5 m 3 /h over the coil. The heat released
by the sorption of the refrigerant in the sorbent is transferred to the cooling water flowing in the tubes
of the coil, increasing its temperature of 30 o C. The cooling water circulates to the condenser and then
to the cooling tower of the chiller where the sorption heat is rejected to the surroundings by
evaporation. The concentrated sorbent solution becomes dilute by absorbing the refrigerant vapor. The
dilute sorbent solution, collected in the solution reservoir at the bottom of the lower vessel, is pumped
back to the HTRG and LTRG with pressure about 10 kPa and 100 kPa, respectively, either in series or
in parallel by the solution pump for regeneration.
2.1.4 High-Temperature Regenerator
The water-LiBir sorbent solution, diluted by absorbed water refrigerant vapor, is pumped in the Broad
chiller to the two regenerators in parallel: the HTRG and the LTRG. In each regenerator, the
refrigerant water vapor added to the sorbent in the absorber is removed by evaporation at elevated
temperature and pressure. Approximately equal quantities of sorbent solution are fed to each
regenerator controlled by a flow restriction device in the pipe leaving the solution pump. In the
HTRG, steam in a coil is used to boil off refrigerant vapor from the sorbent. The temperature and
pressure of the refrigerant vapor produced in the HTRG is high enough to generate an approximately
equal quantity of refrigerant vapor from the sorbent in the LTRG operating at a lower temperature and
pressure. The driving heat provided to the HTRG is thus cascaded and used twice. This makes the
absorption cycle a two-stage process. The generation of additional refrigerant from a given heat input,
improves significantly the cycle performance.
27

The design of the HTRG differs depending both on the heating medium, gas, or liquid, and on its
temperature. Many forms of thermal energy can be used in the HTRG to drive a two-stage absorption
chiller, such as steam, hot water, exhaust gas, natural gas, oil, and liquid pressurized gas. In this
section, only the steam-driven HTRG is discussed; other kinds of heat sources - natural gas, hot water,
and exhaust gas - are discussed in the sections that follow.
The water-LiBr sorbent, reconcentrated in the regenerators, returns to the absorber through flow
restrictions that assist in maintaining appropriate liquid levels to submerge the heat transfer coils in the
regenerators. The solution pump frequency is adjusted to maintain a constant level in the HTRG.
Both the HTRG and the LTRG use water vapor as
a heat resource; they have similar functions and
structure. The heat transfer process includes
condensation inside the tubes and boiling on the
outer surface of these tubes.
The configuration of the upper vessel for the
absorption chiller installed in the IW is similar to
that of a natural gas direct-fired absorption chiller
of the same capacity shown in Figure 2-5. The
combustion chamber and convection chamber of
the natural-gas-fired HTRG are replaced by a
spiral tube bank in the steam-driven HTRG to
vaporize water refrigerant from the water-LiBr
sorbent.
Figure 2-5: Configuration of the upper vessel
The major part of the HTRG is a spiral tube bank with three parallel tubes spiraling eleven rounds
from the top to the bottom. Steam supply flows in parallel through the tubes from top to bottom. The
dilute sorbent solution is pumped into the HTRG from the bottom of the tank, and the concentrated
sorbent solution leaves the HTRG from the bottom of the tank at a distant point. The vigorous mixing
resulting from the boiling in the regenerator minimizes sorbent concentration differences in the HTRG.
While mass transfer is involved as water diffuses to and is evaporated from the sorbent-vapor
interface, the vigorous mixing minimizes mass transfer resistance. The HTRG thus functions as a
well-mixed two-phase boiling heat exchanger.
28

At design conditions, the HTRG requires a steam supply at 0.7 mPa; the maximum steam supply
pressure is 0.9 mPa, and the maximum flow rate is 24 kg/h. An elevated pressure, typically at a
saturated vapor pressure of 100 kPa, is maintained in the HTRG to provide a condensing temperature
of about 100 o C.
2.1.5 Low-Temperature Regenerator
The LTRG is a staggered tube bank with 14 parallel tubes circulating once around. Vapor from the
HTRG enters at one end of each parallel tube, and condensate leaves the other end of the tubes and
enters the condenser. One end of each tube is connected to the HTRG, the other, to the condenser. The
refrigerant, water, and vapor from the HTRG passes through the LTRG tubes and transfers the heat of
condensation to the sorbent solution surrounding the tube bank. The dilute sorbent solution enters the
LTRG on the top; the concentrated sorbent solution leaves from the bottom. Refrigerant vapor is
boiled off; the dilute sorbent solution is concentrated. Similar to the HTRG, the boiling process in the
LTRG is violent; bubbles stir the sorbent solution. The concentration of the sorbent in the LTRG is
therefore nearly uniform, close to the exit value, and mass transfer is not a limiting process. Similar to
the HTRG, the LTRG functions as a well-mixed, two-phase boiling heat exchanger.
The LTRG has a lower boiling temperature and pressure than the HTRG. At design conditions, a
medium pressure, typically at a saturated vapor pressure of 10 kPa, is maintained to provide an
evaporating temperature of about 45 o C. The LTRG has no solution level control like the HTRG, but
the maximum solution level is measured in the LTRG to prevent crystallization in the LTHX. The
details of the control principles are discussed in appendix 2.A.
2.1.6 Condenser
The condenser and the LTRG are housed in the same vessel with the HTRG, and they operate at the
same intermediate pressure. The condensate from the LTRG flashes into the condenser operating at
intermediate pressure. The condenser then condenses both the vapor produced in this flashing and the
water vapor from the LTRG, transferring heat into cooling water flowing into the condenser coil. This
condensate is returned to the evaporator.
The condenser is a spiral copper tube bank with three parallel tubes spiraling three rounds from the
bottom to the top. The cooling water flowing from the absorber enters the condenser from the bottom
and leaves the condenser to the cooling tower at the top. The liquid condensed from the vapor as a
29

film on the surface of tube bank drips down to a drain pan that separates the condenser from the LTRG.
The condenser functions as a two-phase, countercurrent heat exchanger.
2.1.7 Heat Recovery Devices
In Figure 2-2, the four minor heat transfer components in the chiller are used to recover thermal
energy by heat exchange between the various refrigerant, sorbent, and steam condensate streams. All
these exchangers are single-phase, countercurrent heat exchangers that recover heat from a hot stream
and deliver it to a cold stream. One is the LTHX, and the other is the HTHX. These interchangers
reduce the heat requirements of the regenerators and the cooling requirement of the absorber.
In the chiller, the temperature of the condensate leaving the HTRG is high enough to be used to
preheat the dilute solution from the LTHX before it enters the LTRG. A heat recovery exchanger
between the steam condensate and the sorbent stream entering the LTRG reduces the heat requirement
of the LTRG and the temperature of the steam condensate, avoiding its flashing in the condensate tank.
A heat recovery exchanger between the water refrigerant leaving the condenser and the sorbent pool in
the absorber, called the by-pass heat exchanger (BPHX), increases cooling in the evaporator. Broad
terms it an elbow-heat exchanger. In the elbow, the liquid refrigerant condensed from the condenser
releases a small amount of heat to the dilute solution in the absorber.
2.1.8 Cooling Tower
A cooling tower is widely used to dissipate reject heat from a water-cooled air-conditioning system to
the surroundings. This Broad absorption chiller has a built-in cooling tower, as shown in Figures 2-2
and 2-6. Its compact design facilitates chiller installation and operation. The cooling water in the
chiller flows in series through the absorber, the condenser, and then through the cooling tower. This
arrangement provides for a minimum operating temperature in the absorber that is required to achieve
a low chilled-water temperature; the high flow in both the absorber and the condenser provides for
high heat transfer coefficients in these components. The recirculating cooling water flows down
vertical plates in countercurrent contact with upward-flowing ambient air.
Evaporation of a small portion of the water flowing downward through the cooling tower reduces its
temperature; makeup water added to the cooling tower replaces that evaporated. The air temperature is
also reduced, but the humidity increases markedly. Thus, the cooling tower functions as a two-phase
countercurrent heat and mass exchanger.
30

As illustrated in Figure 2-6, the cooling tower attached
to the chiller comprises spray nozzles, vertical PVC
plates, a PVC mist collector, a cooling-water tank, a
cooling-water pump, a cooling-water by-pass valve,
and a cooling-air fan along with devices for water
drain and city-water supply and detergent addition.
The major components of the cooling tower are the
PVC vertical plates (a heat and mass transfer medium)
that increase water/air contact surface as well as the
duration of contact. The closely packed vertical PVC
plates are spaced with staggered bars installed below
the spray nozzles in the air path. At design conditions,
the cooling water is distributed from the top of the
tower through spray nozzles at a temperature of 35.5
o
C. The speed of the cooling tower air fan is varied to
maintain the cooling-water supply to the chiller at 30
o C.
Figure 2-6: Configuration of cooling tower
From chiller
City water
To chiller
CWP
Drain
CWBPV
CWDV
CTWS
CWDD
Air outlet
Air inlet
PVC mist
collector
Spray
nozzles
PVC Plates
CTF
Water tank
As illustrated in Figure 2-6, the cooling tower attached to the chiller comprises spray nozzles, vertical
PVC plates, a PVC mist collector, a cooling-water tank, a cooling-water pump, a cooling-water bypass
valve, and a cooling-air fan along with devices for water drain and city-water supply and
detergent addition. The major components of the cooling tower are the PVC vertical plates (a heat and
mass transfer medium) that increase water/air contact surface as well as the duration of contact. The
closely packed vertical PVC plates are spaced with staggered bars installed below spray nozzles in the
air path. At design conditions, the cooling water is distributed from the top of the tower through spray
nozzles at a temperature of 35.5 o C. The speed of the cooling-tower air fan is varied to maintain the
cooling-water supply to the chiller at 30 o C.
2.1.9 Vacuum System
The pressure of the evaporator and the absorber is significantly below atmospheric pressure, and air
can leak into the absorption chiller. Corrosion can also occur in the chiller, generating another
noncondensable gas, H2. Air and other noncondensable gases in the evaporator and absorber can
seriously reduce the rate of heat and mass transfer processes there and thus reduce the overall cooling
31

capability of the chiller. An appropriate means for removing noncondensable gases is essential to the
operation of microscale absorption chillers.
An automatic gas purge device (AGPD) has been provided in the chiller to continuously remove
noncondensable gases from the absorber and the evaporator to maintain the required vacuum. The
vacuum can be maintained through the AGPD and/or by periodic manual vacuum removal. The
advantage of using the AGPD is that the noncondensable gases are continuously removed from the
refrigerant vapor, so the pressure in the absorber and the evaporator vessel remains constant until the
storage chamber is full. Noncondensable gas is generated in the upper vessel (the HTRG and the
LTRG), but is hard to remove through the AGPD. Even if an automatic purge unit is installed,
therefore, manual vacuum removal is still required to purge the noncondensable gas from the storage
chamber and the upper vessel. The detailed mechanisms for controlling noncondensable gas are
described in appendix 2.A.
2.2 Absorption Chiller Test Systems
2.2.1 System Description
A system was set up to test the Broad BCT 16 absorption chiller and to evaluate its performance under
a wide range of external and internal operating conditions. This test system, shown in Figure 2-7,
comprises the following equipment in addition to the chiller:
• a steam supply
• a variable cooling load
• an instrumentation, control, and data acquisition system
In Figure 2-7, the absorption chiller is in the middle. It is connected with the steam supply system on
the left and the variable cooling load system on the right. The necessary control and instrumentation
for the overall system has been installed to operate this test system and to process the data it provides.
The measurement data are used both to monitor the system status and to calculate chiller performance.
32

Figure 2-7: Simplified flow diagram of the chiller test system
CTW
WS BFT
Absorption chiller
2.2.1.1 Steam Supply System
BFP
P7 T25
ESB
Steam supply system
P4
CR
T22
SS
T23
P5
ALC
Additional
sensors to
chiller
F2 P3 T21
Absorption
chiller
F1
Cooling loads Steam supply
T20
P2
BHWSV
HWS
IV5
IV2
HWR
TLHX
IV4
IV3
Variable cooling load
To conduct the tests, steam is generated on site by a steam supply system. The CMU campus has a
steam supply grid, but the closest possible point of connection is remote from the chiller (six floors
below). The campus steam supply is used mainly for the building heating system; the steam supply
pressure is high in the winter, about 0.7 mPa, but low in the summer, about 0.4 mPa – lower than that
specified to operate the chiller, 0.7 mPa. An electric steam boiler (ESB) was procured and installed
along with its auxiliaries to supply steam for testing the chiller. The boiler auxiliaries, shown in
Figure 2-7, include:
• a boiler feed receiver tank (BFT)
• a boiler feed pump (BFP)
• a boiler blowdown separator (BBDS)
• a boiler system chemical feeder (BSCF)
• a water softener
• a boiler chemical treater (BCT)
The ESB is capable of providing steam at a maximum pressure of 1.0 mPa; its rated capacity is 24
kW. The boiler capacity and pressure range is sufficient to drive the absorption chiller with a rated
cooling capacity of 16 kW. Steam pressure is adjusted by on/off control of the two horizontal
33

electrical heaters mounted in the base of the boiler. The water level in the boiler is maintained around
a set point that submerges the heaters by on/off control of the boiler feed pump that delivers
condensate from the receiver. A water level set point in the receiver controls the input of water from
the tap through an ion exchange water treatment system. Chemicals are added to the condensate in a
treater to avoid corrosion and deposits in the system.
At design conditions, the ESB provides steam with a pressure of 0.7 mPa to the absorption chiller. In
the chiller, the steam is condensed in the HTRG; and the condensate subcooled in the HRHX at 0.1
mPa. The condensate from the chiller is then collected in the BFT at atmospheric pressure to serve as
feedwater to the boiler. The other source of feedwater in the BFT is city water, which is pretreated
through the water softener. The water softener includes two water treatment tanks filled with ion
exchange resin and a sodium chloride salt tank. The ion exchange resin in each of the tanks is
regenerated periodically by the sodium chloride salt solution. The boiler feedwater is delivered to the
ESB by the BFP.
Figure 2-8: Site views of the absorption chiller test system
2.2.1.2 Variable Cooling Load System
Systematic testing of the chiller requires a load that can be adjusted independently and maintained
constant during a test run. This load is provided by a shell-and-tube heat exchanger fed with water at
34

80 o C to the shell from the building hot water grid. The flow of chilled water from the chiller outlet to
the tubes of the load exchanger is controlled by a valve to achieve a desired flow set point. The flow
of hot water to the exchanger is also controlled by a valve (BHWSV) to maintain a desired set point
temperature for the chilled water at the inlet to the chiller. Under design conditions, the chilled water
flows in the chiller at a rate at 2 m 3 /h and a temperature of 14 o C. The chiller cools the chilled water to
7 o C. Figure 2-8 shows photographs of the overall system. The photo on the left indicates the ESB,
BFT, and some chilled-water supply and return pipes. The photo on the top right shows the absorption
chiller installed on the deck adjacent to the IW. The picture on the bottom right shows the variable
cooling load system.
2.2.2 Instrumentation, Control, and Data Acquisition System
For operation of the chiller test system, an instrumentation, control, and data acquisition
system has been provided by the Automated Logic Co. (ALC). It collects the measurement data
from the operation of the absorption chiller and the auxiliary steam supply and cooling-load systems to
evaluate the chiller under various conditions and to assess chiller internal working conditions. The
system also displays the operational data in various forms such as trends and bar charts, and stores
data for future analysis. The ALC system is a web-based control and data display system, so the
operators can operate the system and access the experimental data via the Internet.
2.2.2.1 Structure of Instrumentation Control System
The structure of the overall control and instrumentation system is illustrated in Figure 2-9. The left
side is the internal control system from the chiller manufacturer, and the right side is the ALC control
system for the steam supply and variable cooling load systems. The additional sensors used to monitor
the chiller internal conditions were also implemented through the ALC control system. The ALC
control system is one of the basic platforms for building an integrated BCHP system. It can not only
perform regular control and data acquisition functions, but it can also perform more complicated tasks
such as system diagnostics and optimization when the overall system becomes more complicated.
With this automation control system, the overall chiller test system can be started up, shut down, and
adjusted automatically or manually through a computer. The operating conditions can be displayed on
a graphic interface, and the measured data can be collected for further analysis.
35

Figure 2-9: Control and instrumentation structure
The sensors installed through the ALC system include:
• the additional sensors (surface type) for the chiller
• the sensors for the steam supply system
• the sensors for the variable cooling load system
The chiller internal control system receives the measurement signals from the sensors and sends
commands to the control points (components) through the chiller control panel. The control algorithm
provides for startup, shutdown, and operation of the chiller on the basis of the sensor information. The
chiller control will be discussed in appendix 2.A. The chiller operational status is monitored through a
remote control device or a computer; the measurement data are stored in the computer.
2.2.2.2 Data Acquisition and Display
User friendly interfaces are important to operate the chiller, the auxiliary steam supply, and the
variable cooling load systems. The operators can operate the chiller automatically or manually in
startup, shutdown, and adjustment of the system. The measurement data are displayed instantaneously
on the monitor and are stored in the computer for future analysis.
The chiller monitoring interface is illustrated in Figure 2-10. The monitoring interface is a small and
independent software package with data collection functions. It is a good tool for monitoring the status
of the chiller instantly and displaying the information graphically. Through the interface, the operator
36

can perform startup, shutdown, and other operational actions to the chiller automatically and can
adjust chiller operational parameters such as set points. The chiller operation status (on or off) is
displayed in the top row; the temperature measurements are displayed on the left of the interface. The
solution level in the HTRG and the solution pump status are displayed in the middle of the interface,
and the valve positions and other pump states are displayed at the lower part of the interface. On the
right of the interface are warning and alarming messages.
Figure 2-10: Absorption chiller monitoring software
Parallel to the computer monitoring interface, an on-site key pad monitoring system is mounted near
the chiller. This system has the same functions as the control software installed in the central computer
in Figure 2-10 that can save the measurement data in an Excel spreadsheet.
The ALC control software is called the web control server (WCS). As an example, Figure 2-11
displays additional sensor measurements for the chiller and for the steam supply and variable load
systems through the ALC WCS data display system.
The WCS plots historical data in various forms, such as graphics, trends, and spreadsheets. The
measurement data can be sampled in any time step from a second to a year. The ALC control system
has the potential to communicate with the chiller control module directly through a standard
37

communication port; this function is usually called “the third party integration”. Third-party
equipment, like the absorption chiller, becomes a subsystem that can receive commands such as
startup and shutdown from the WCS, a primary control system. The manufacturer of the absorption
chiller, however, uses nonstandard control protocol, so a software driver to translate the chiller control
protocol into a commercial standard was required to implement “the third party integration”. The
chiller control system and the ALC control system were installed and worked as two independent
systems in parallel.
Figure 2-11: Test system monitoring software
2.2.2.3 Instrumentation for the Chiller
Figure 2-12 is the process and instrumentation (P&I) diagram for the chiller. This figure illustrates
chiller components, piping, and the measurement and control points. The chiller components and
configurations have been described in Figure 2-2, Table 2-1, and section 2.1. This section discussed
the instrumentation and control of the chiller. The absorption chiller has its own controls and
instrumentation from the manufacturer; additional sensors are installed for the chiller by the ALC
during chiller installation to study its internal operation conditions. In Figure 2-12, the sensors from
38

the manufacturer are indicated in green. The temperature sensors and flow meter provided by the ALC
are indicated in blue. The red lines indicate the control points of the chiller from the manufacturer,
and the pink ones refer to the controls for the steam supply and cooling-load systems.
The instruments and sensors installed in the chiller by its manufacturer indicate chiller operating
conditions. These temperature, level, flow, and electric power measurements are listed in Table 2-3.
There was a total of 16 measurements from the chiller manufacturer. The configurations and the
functions of these sensors are discussed in appendix 2.A.
Figure 2-12: PI&D diagram of the absorption chiller
To auxilary system
ME-LGR 25
Steam
Condensate
CHWS
CHWR B1 T1 B2
Drain
City water
T9 T2
T6
T10
Controller
T7
SF
T14
P1
SV
T8
RBPSV
HRHX
L2
RP
CHWP
L3
CHSV
T11
CTWS
Evaporator
T33 T13
RPH
SP
LTRG
HTRG
Condenser
T19
HTHX LTHX
Absorber
T12
T3
T17 T18
T5
L1
T15
T32
L4
CWP
T16
L5
F6
CWBPV
DV
CWDD
Hot Humid Air
CTF
Air
CWF
39

The ALC installed 11 additional temperature sensors and a flow meter in the chiller for this study to
obtain further information on its operation, Table 2-3. The temperature sensors were mounted on the
surface of the chiller vessel and piping. This is an economical and convenient method, but the heat
conduction through the pipe and heat loss to the surroundings affects the accuracy of the
measurements.
Serious consideration was given to installing of three pressure sensors to indicate pressure levels in the
evaporator and in each of the two regenerators of the chiller. Broad advised against penetrating the
chiller housing because of the possible introduction of air leakage or corrosion at the point of sensor
installation.
2.2.2.4 Instrumentation for the Auxiliary Systems
The steam supply system and the variable cooling-load system were discussed in sections 2.2.1 and
2.2.2. The system configuration is indicated in Figure 2-7. Table 2-3 lists the seven measurements of
the steam system provided by the ALC system, which are also indicated in Figure 2-7. Among these
sensors, the steam flow rate, steam supply temperature, and condensate return temperature are used to
calculate the quantity of heat input to the chiller. These sensors measure the fluid directly.
Table 2-3 also lists the six sensors installed for the variable cooling load system. There are a total of
six temperature sensors, two flow meters, and four pressure transducers. Among these sensors, the
chilled-water flow rate (F1), chilled-water supply temperature (T21), and chilled-water return
temperature (T20) were used to calculate the cooling capacity of the chiller. All these sensors and
meters measure the fluid directly.
2.2.2.5 Instrumentation Calibration
These sensors provided by the ALC were calibrated on site by the following methods:
• the temperature sensor readings were calibrated in ice water and boiling water.
• the condensate from the chiller was collected in a barrel; the weight of the condensate was
measured every 15 minutes for 2 hours. The weight of condensate was compared with the
measured values of the steam flow meter.
• the chilled-water flow meter piping configuration was sent back to the flow meter
manufacturer for calibration; the suggested deviation has been applied to the chilled-water flow
measurement.
40

The calibration results are indicated in the last column of Table 2-3. The accuracies of the resistance
temperature detectors (RTD) type temperature sensors are within ± 0.2%. The surface temperature
sensors calibrated in ice water are accurate to ± 1.5% but have an accuracy of ± 0.5% in boiling
water. The calibration offset values are assigned to the sensors listed in the last column of Table 2-3.
The steam flow meter gives higher accuracy at high flow rate and pressure. On average, when the
steam flow rate is higher than 12 kg/h, the steam flow meter accuracy is within 10%, but when the
steam flow rate drops below 12 kg/h, the steam flow meter accuracy is within roughly 50%. At design
condition, the steam flow meter indicates a value only 2% lower than the condensate weight. Because
of these inaccuracies, the chiller heat input was calculated by measuring the power input to the boiler.
According to the manufacturer, the boiler efficiency is 98% to 99% under design load and off-design
load conditions. In calculating the chiller performance, therefore, the power measurements of the
boiler are more reliable than the steam flow measurements.
2.2.3 Controls for the Chiller
On the basis of sensor inputs, the chiller control algorithms determine the outputs to the actuators and
control points on various system components. The chiller has a total of 12 control components listed in
Table 2-4, and the sensor locations, types, and the configurations are indicated in Figure 2-12. The
features of these control components are discussed in appendix 2.A.
Table 2-3: Control points of the chiller
Abbrev. Name Signal Category
CHWP Chilled-water pump Digital On/off
CWP Cooling-water pump Digital On/off
SP Solution pump Analog Quantity control
RP Refrigerant pump Digital Quantity control
CTF Cooling-tower fan Analog Temperature control
SV Steam valve Analog Operation time control
RBPSV Refrigerant by-pass valve Digital On/off
CTS City-water switch Digital Quantity control
CWDD Cooling-water drain device Digital On/off
CWDV Cooling-water detergent valve Digital On/off
CWBPV Cooling-water by-pass valve Analog Temperature control
CTF Cooling-tower fan Analog Temperature control
RPH Refrigerant-pump heater Digital Temperature control
41

Table 2-4: Instrumentation of the chiller test systems
Absorption Chiller (manufacturer)
Absorption Chiller (ALC)
Steam System
Cooling Load System
Temp.
Level
Power Flow
Temp.
Misc.
Temp.
Pres.
Misc.
Temp.
Flow
Pres.
Label Sensor location Medium Range Manufacturer
Accuracy
T1 Chilled-water return Water (-15 o C) to 110 o C ± 0.1%
T2 Chilled-water supply Water (-15 o C) to 110 o C ± 0.1%
T3 Cooling-water supply Water (-15 o C) to 110 o C ± 0.1%
T5 High-temperature regenerator Solution (-15 o C) to 210 o C ± 0.1%
T6 Ambient Air (-15 o C) to 110 o C ± 0.1%
T7 Steam supply Steam (-15 o C) to 210 o C ± 0.1%
T8 Condensate return Water (-15 o C) to 210 o C ± 0.1%
T9 Chilled-water return 2 Water (-15 o C) to 110 o C ± 0.1%
L1 HTRG solution level probe Solution 4 pins
L3 LTRG upper-limit level probe Solution 1 pin
L4 Auto
probe
vacuum device level Solution 1 pin
L5 Cooling-water level probe Water 1 pin
B1 Chilled-water flow detector Water on/off
B2 Chilled-water flow detector Water on/off
D1 Solution pump frequency, Electricity
amps, and voltage
On-site
Calibration
T11 Condensate after HTRG Surface 0-400 o F ± 0.1% 0.5 o F
T12 Solution in Absorber Surface 0-400 o F ± 0.1% 0.5 o F
T13 Solution entering HRHX Surface 0-400 o F ± 0.1% 2 o F
T14 Solution leaving HRHX Surface 0-400 o F ± 0.1% 1.5 o F
T15 Cooling water after absorber Surface 0-400 o F ± 0.1% 1.5 o F
T16 Cooling water after condenser Surface 0-400 o F ± 0.1% 1.0 o F
T17 Low-temperature
(LTRG)
regenerator Surface 0-400 o F ± 0.1% 0.3 o F
T18 Refrigerant after condenser Surface 0-400 o F ± 0.1% 1.5 o F
T19 Refrigerant from evaporator Surface 0-400 o F ± 0.1% 2.0 o F
T32 Cooling water after cooling
tower
Surface 0-400 o F ± 0.1% 0.0 o F
T33 HTRG temperature Surface 0-400 o F ± 0.1% 1.5 o F
F6 Cooling-water flow Water 0 to 30 gpm ± 1% -15%
E1 Electric power of absorption
chiller
Electricity 0-2400 amps ± 1% 0.0%
T22 Steam-supply temperature Steam (50 o F) to 250 o F ± 0.1% 0.0 o F
T23 Condensate-return
Water (50 o F) to 250 o F ± 0.1% 0.0 o F
temperature
T25 Feed-water temperature Water (50 o F) to 250 o F ± 0.1% 0.0 o F
P4 Steam supply pressure Steam 0 to 40 gpm ± 0.13%
P5 Condensate return pressure Water 0 to 100 psi ± 0.13%
P7 Feed-water pressure Water 0 to 50 psi ± 0.13%
F2 Steam flow Steam 0 to 75 lb/h ± 0.5% 0-10%
E2 Electric power of steam boiler Electricity 0-2400 amps ± 1% 0.0%
T20 Chilled-water supply Water (-10 o F) to 110 o F ± 0.1% 0.6 o F
T21 Chilled-water return Water (-10 o F) to 110 o F ± 0.1% -0.4 o F
T31 Ambient temperature Air (-58 o F) to122 o F ± 0.1%
P2 Chilled-water inlet Water 0 to 100 psi ± 0.13%
P3 Chilled-water outlet Water 0 to 50 psi ± 0.13%
F1 Chilled water Water 0 to 20 gpm ± 1% 0%
42

Compared to traditional pneumatic or electric controls, the use of electronic controls with advanced
control algorithms makes the complicated absorption chiller more efficient and reliable. Control
categories for the chiller are:
• startup and shutdown
• chilled-water supply temperature control
• cooling-water supply temperature control
• vacuum maintenance
• crystallization judgment and de-crystallization
• safety and diagnostics
The details of the chiller control principles of the six categories are discussed in appendix 2.A.
Knowledge of the chiller controls greatly improves the understanding of the chiller, which, in turn,
assists in improving the accuracy of the computational model discussed in chapter 3.
With the control, instrumentation, and data acquisition systems, the absorption chiller can be tested
under various load conditions. On the basis of a test program, the chiller performance was investigated
by varying the operational parameters individually. The testing approaches and results will be
discussed in the following sections.
2.3 Chiller Performance and Test Program
The chiller was first tested at design condition and then under off-design conditions on the basis of a
test program.
2.3.1 Chiller Testing
2.3.1.1 Chiller Test
An individual chiller test was conducted by setting the six operating conditions that are the primary
input to the test system, all external to the chiller:
• the pressure of the saturated steam supply
• the flow rate, inlet , and outlet temperature of the chilled water
• the flow rate and inlet temperature of the cooling water. Ordinarily, the chiller cooling- water
pump maintains a constant flow; and the air fan maintains a constant supply temperature by
varying its speed in response to the cooling load and the ambient air conditions. To test the
43

chiller over a broader range of operating conditions, however, measures adjust cooling-water
flow and temperature were taken
The chilled-water outlet temperature setting in the chiller control system was maintained constant at 7
o
C throughout the test program. While this setting remained constant, the measured value of chilledwater
outlet temperature varied ±2 o C from the set point, depending on the test conditions. At a given
setting of operating conditions, the chiller was allowed to reach steady-state operation. Three primary
performance conditions were measured:
• the chilled-water outlet temperature
• the steam flow
• the power consumption of the chiller in its pumps, fan, heater, and controls
Steady-state was established by observing that these conditions had a constant average value over a
period of 20 minutes or longer. The chiller load, COP, and power consumption were calculated for the
test. The chiller load is the product of the chilled-water flow, the temperature difference between the
inlet and outlet chilled-water temperature, and the specific heat of the chilled water. The COP is the
quotient of the chiller load and the enthalpy difference of the inlet steam and the outlet condensate
from the chiller. In addition, all the input data from sensors and output signals to actuators in the
chiller, steam supply, and variable load for each test were recorded and stored in the data acquisition
system for further consideration and analysis as described in chapter 3.
2.3.1.2 The Chiller Test Program
A chiller test program was planned and executed. Each of the six operating conditions, identified
above, was varied one at a time over a range of design values, as indicated in Table 2-5. Within its
range each operating condition was tested at 5 to 10 values. Ranges of each of these six variables are
indicated in Table 2-5. Each test collected 20 to 200 data sets obtained at 2-minute intervals during
steady-state operation of the chiller. A total of 38 tests were conducted over an estimated 220 hours of
chiller operation.
The results of these tests in terms of the steam flow and chilled-water outlet temperature, the chiller
load and the coefficient of performance, is reported and discussed in subsection 2.4 below.
44

Table 2-5: Input and primary output of the test program
CHW
return T
CHW
flow
Inputs Primary Outputs Calculated performance
CW
supply T
CW
flow
Steam
pressure
Steam
Flow
CHW
supply COP
Cooling
load
o C kg/s o C kg/s kPa kg/s o C kW
Design
condition 13.9 0.5616 30.78 1.546 600 0.00727 7 1.03 16.63
CHW
0.00382–
return T 8–14 Design Design Design Design 0.00771 6.25–6.9 0.95–1.03 9–18
CHW
0.531–
0.00705– 6.42–
flow Design 0.864 Design Design Design 0.00771 8.97 0.93–1.02 16.56–17.96
CW
0.00609– 6.13–
supply T Design Design 27.5–36 Design Design 0.00764 9.26 1.11–0.91 19.22–13.29
0.784–
0.00618–
CW flow Design Design Design 1.547 Design 0.00709 7.52–9 0.81–0.98 11.73–15.23
Steam
0.00556– 6.34–
pressure Design Design Design Design 360–700 0.00748 10.27 0.65–0.99 8.41–17.47
2.3.2 Conduct of the Testing Program
In the testing program, various procedures were used to adjust the six operating conditions.
• The chilled-water return temperature was varied by adjusting the hot water supply flow to the
variable load.
• The chilled-water flow was varied by adjusting a ball valve in the chilled-water supply pipe,
but when the chilled-water flow was reduced below the design flow rate for the chiller, a
“water cut” warning was reported, and the chiller automatically executed its shutdown
procedure. Only design and higher chilled-water flows have been tested.
• The cooling-water supply temperature was varied by adjusting the air fan speed.
• The cooling-water flow rate was varied by blocking a portion of the cooling-water filter
located at the bottom of the cooling tower.
• The steam supply pressure was varied by adjusting the pressure set point in the boiler control
system.
Fewer tests have been performed at operating conditions that result in loads below the design values.
When the cooling load is below 50% of the design load, it has proved difficult to obtain stable data for
analysis because of on-off cycling of the steam valve.
2.4 Chiller Performance
The chiller performance has been calculated on the basis of the resulting measurements. The chiller
performance has been compared with the chiller specification data Broad provided. Results show that
45

the chiller cooling capacity is higher than its rated capacity of 16 kW. The chiller has a COP of 1.0,
slightly below its specified value.
2.4.1 Chiller Performance Calculations
The performance of an absorption chiller is determined by its cooling capacity, coefficient of
performance, and electric power consumption. These quantities are defined in the following equations
The cooling capacity or chiller load is
cooling
CHW
p
( T T )
Q = m&
× C × −
(Equation 2-1)
CHWR
CHWS
Where m& CHW is the flow rate of the chilled water
T CHWR is the temperature of the chilled water entering the chiller
T is the temperature of the chilled water leaving the chiller
CHWS
C is the specific heat of water.
The COP of the chiller is conventionally
thermal
p
Q
Q
cooling
COP = (Equation 2-2)
heat
Where, Q m × ( h − h )
heat
= & = the heat delivered to the chiller by steam
steam
steam
condensate
condensation; this quantity can also be estimated from the electrical power
consumption in the steam boiler.
m& steam is the flow rate of the steam supply
h steam and h condensate are the enthalpies of the steam supply and the condensate,
respectively.
An overall COP can be defined that includes both the thermal and the electrical energy supplied to the
chiller
Q
COP = (Equation 2.3)
energy
cooling
overall
Qenergy
Q Q + E
= (Equation 2-3)
heat
E =
V * I
power
power
power
power
46

Where, the power consumed in the pumps, fan, heater, and control of the chiller.
The power consumption of the chiller is approximately 8% of the total energy supplied; the thermal
COP ( COP thermal ), usually used to represent chiller performance. is therefore only slightly greater than
the (COPoverall ).
2.4.2 Chiller Performance under Design Condition
The chiller has been tested under the design conditions indicated in Table 2-5. In the test, the chiller
was started up and operated for a period of time before the steam supply system was started. This
procedure is called a “cold” start of the chiller. The “cold” start of the chiller provides an opportunity
to check the accuracy of the sensors. For example, there are two measurements of the cooling water
temperature: at the cooling tower outlet after the absorber and after the condenser. In the “cold” start,
these two sensors should indicate the same, ambient temperature. Similarly, the sorbent solutions in
the absorber and the high-temperature and low-temperature regenerators should be equal. Figure 2-13,
showing a typical cold start, startup, and steady-state operation of a chiller test, confirms these
expectations. Once the steam supply system is provided, all these stream temperatures diverge as
steady-state operation is approached.
Figure 2-13: Typical start-up of the chiller test system
47

The steady state is defined by observing the monitoring interface. Figure 2-14 shows steady-state
measurement data over a period of 1 hour and 45 minutes. The time step for the measurements was
preset at 2 minutes; a total of 67 data samples were collected for this design load test.
The steam flow fluctuated around an average value within ± 20%
. A drop in steam flow occurs
because the boiler feed pump periodically feeds condensate to the boiler at a temperature lower than
the boiling point. The boiler temperature, pressure, and steam flow are consequently reduced until the
electrical elements of the boiler heat its water content to the preset boiling pressure. This periodic
action of the boiler feed system repeats as shown in Figure 2-14. The cooling capacity of the chiller is
rather stable around 17.5 kW; but the calculated thermal COPthermal, directly dependent on the steam
flow, fluctuates around an average value of 1.0.
Figure 2-14: Steady-state operation of the chiller under design load condition
Steam supply flow rate (kg/h),
Cooling load (kW)
30
25
20
15
10
5
0
Cooling load (Q cooling)
Steam flow rate (F1)
COP thermal
21:30 21:38 21:46 21:54 22:02 22:10 22:18 22:26 22:34 22:42 22:50 22:58 23:06 23:14
Steam supply flow rate Cooling load COP
Figure 2-15 shows the stream temperatures for the same test as Figure 2-14. The data indicate the
chilled-water supply and return temperatures at the bottom of the chart. The chilled-water supply
temperature is very stable compared with the steam supply temperature on the top of the chart and the
cooling water temperatures in the middle. The sorbent solutions in both regenerators are affected by
the temperature of steam supply. The refrigerant temperature after the condenser is plotted also; this
temperature is apparently affected by the temperature of the steam supply, but the curve is flatter than
the sorbent solution in the HTRG and the LTRG.
4
3
2
1
0
Time
3.5
2.5
1.5
0.5
Cofficient of Performance (COP)
48

The cooling water fluctuates some because of the operation of the cooling-tower fan. The coolingwater
supply temperature is consistently maintained at 30 o C within ± 1
o C deviations. Although the
conditions of the outside stream (steam, cooling water, chilled water) vary as indicated in Figure 2-14,
the absorption chiller reduces these effects by its internal control system. As a result, it is more
convenient to use the temperatures in the chiller as an indicator of steady-state operation.
Figure 2-15: Steady-state operation of the chiller under design load condition
Temperature ( o C)
180
160
140
120
100
80
60
40
20
0
HTRG sorbent
LTRG sorbent
Cooling water supply and return
Refrigerant after Condenser
Steam supply
Chilled water supply
Chilled water return
21:30 21:38 21:46 21:54 22:02 22:10 22:18 22:26 22:34 22:42 22:50 22:58 23:06 23:14
CHWR CHWS SS
CWS CW after ABS CW after COND
LTRG HTRG Refrigerant after COND
Table 2-6 presents the average measured data of the chiller operating at design condition before and
after sensor calibration. Before the calibration means that the test data were collected prior to the
instrumentation processes, and after the calibration means that the test data were collected after the
sensor were relocated, well insulated, and the offsets of calibration are applied through the data
acquisition systems. The differences of the measurement values between the two tests are small
because the sensor calibration offset values in Table 2-6 are relatively small.
Table 2-6 shows that all temperature data collected from the data acquisition system of the absorption
chiller that Broad provided (T1 to T9) show only small differences from similar measurements
collected from the ALC system (T11 to T33). The chiller performance has been calculated on the
basis of the measurement data from the ALC system. The measurements in bold are those used in
calculating the chiller performance. It is notable that T11 and T33 have higher deviations before and
after the calibration, about 6.5 and 3 o C, respectively.
Time
49

Table 2-6: Measurement data of the chiller under design condition
Absorption chiller (manufacturer)
Absorption chiller (ALC)
Steam system
Cooling load system
Temp.
Level
Flow
Ele.
Temp.
Flow
Temp.
Pres.
Flow
Temp.
Pres.
Flow
Label Sensor location Medium Unit Before calib. After calib.
T1 Chilled-water return Water
o
C - 13.83
T2 Chilled-water supply Water
T3 Cooling-water supply Water
T5 High-temperature regenerator Solution
T6 Ambient Air
T7 Steam supply Steam
T8 Condensate return Water
T9 Chilled-water return 2 Water
L1 HTRG solution level probe Solution
L3 LTRG upper-limit level probe Solution
L4 Auto vacuum device level probe Solution
L5 Cooling-water level probe Water
B1 Chilled-water flow detector Water
B2 Chilled-water flow detector Water
E1 Solution pump frequency, amps, and voltage Electricity
T11 Condensate after HTRG Surface
T12 Solution in absorber Surface
T13 Solution entering HRHX Surface
T14 Solution leaving HRHX Surface
T15 Cooling water after absorber Surface
T16 Cooling water after condenser Surface
T17 Low-temperature regenerator (LTRG) Surface
T18 Refrigerant after condenser Surface
T19 Refrigerant from evaporator Surface
T32 Cooling water after cooling tower Surface
T33 HTRG temperature Surface
o C - 6.65
o C - 30.01
o C - 154.87
o C - 24.26
o C - 158.56
o C - 99.00
o C - 6.58
o C 151.23 157.65
o C 36.7 36.63
o C 74.99 75.02
o C 90.54 90.37
o C 37.46 37.7
o C 40.06 40
o C 92.71 93.21
o C 43.29 44.08
o C 127.78 129.24
o C 30.78 30.74
o C 150.54 153.44
F6 Cooling-water flow Water kg/s 1.55 1.45
T22 Steam-supply temperature Steam
T23 Condensate-return temperature Water
T25 Feedwater temperature Water
o
C 163.56 164.31
o
C 99.43 99.3
o C 70.09 59.1
P4 Steam supply pressure Steam psi 698.04 713.87
P5 Condensate return pressure Water psi 101.75 100.04
P7 Feedwater pressure Water psi
F2 Steam flow Steam kg/s 0.00727 0.00734
T20 Chilled-water supply Water
T21 Chilled-water return Water
T31 Ambient temperature Air
o C 13.9 13.88
o C 6.83 6.39
o C 28.49 23.2
P2 Chilled-water inlet Water psi 298.58 154.23
P3 Chilled-water outlet Water psi 372.32 233.33
F1 Chilled water Water kg/s 0.56 0.56
50

Broad specified and measured chiller performance data at design conditions are presented in Table 2-
7. The chiller specified cooling capacity is 16 kW, but the cooling capacity of the chiller after the
calibration is calculated to be 17.6 kW. The measured steam flow is about 10% higher than
specifications; but the load is also 10% higher. The COPthermal of the chiller, in either case, is 1.03. The
power consumption of the chiller is 0.82 kW. The results of the two tests, before and after the
calibration of sensors, are quite similar; they confirm the specifications provided by the manufacturer.
Table 2-7: Comparison of chiller performance under design conditions
Chilled water
Steam
Power
Perform.
Name Unit Specification Before calibration After calibration
Cooling capacity kW 16 16.62 17.62
Chilled-water return temperature
o
C 14 13.9 13.88
Chilled-water supply temperature
o C 7 6.83 6.39
Chilled-water flow rate m 3 /h 2 2.02 2.02
Chilled-water pump head mH2O 8 7.9 7.9
Rated steam pressure, absolute mPa 0.7 0.698 0.714
Steam consumption* kg/h 24 26.18 26.43
Power voltage V 220 220 220
Power frequency Hz 60 60 60
Power consumption* kW 1 0.856 0.823
COP (Thermal) 0.98 1.038
COP (overall) 0.93 0.99
* The maximum values for the specification data.
2.4.3 Chiller Performance at Reduced Capacity Condition
The chiller performance at the design condition is reported in subsection 2.4.2 above. To obtain chiller
test data at reduced capacity, the chilled-water return temperature was decreased from 14 o C to 8 o C in
a series of nine tests. In these tests the chilled- and cooling-water flows and the cooling water inlet
temperature were maintained constant. The test conditions and results are summarized in Table 2-8.
As the chilled-water return temperature and, thus, the chiller capacities were reduced, the saturated
steam pressure and temperature to the chiller were also reduced. This reduction was imposed to avoid
on-off cycling of the steam valve of the chiller and the consequent erratic chiller conditions that were
observed if the steam supply pressure was maintained at 700 kPa.
Figure 2-16 shows the COP of the chiller as the cooling capacity (load) was varied from 21% to 100%
of the capacity at design conditions. As the cooling load was varied from 4 kW to 18 kW, the thermal
51

COP varied from 0.7 to 1.04. The overall COP, which considers the power consumption of the chiller,
is, on average, less than the thermal COP by 5%.
Figure 2-17 shows the power consumption of the chiller at various loads. The bold curve is the
measured value for the chiller, and the second curve is the data from Broad’s brochure for a naturalgas-driven
chiller with the same cooling capacity and a similar configuration.
Figure 2-18 compares the measured heat input, Qheat, of this steam-driven chiller with the
manufacturer’s specified heat input from the heat of combustion of the fuel in their natural-gas-driven
chiller. The ordinate is the actual heat input, and the abscissa is the actual cooling load. In the figure,
the bold curve is the steam-driven chiller measured in the IW, and the curve below is the rated
performance curve from the manufacturer. The two curves have similar trends, but the steam-driven
chiller uses more thermal energy than does the natural-gas-driven chiller.
Table 2-8: Primary measurement for chiller input and output
Chilled-
water
flow
Measurement values for chiller inputs Measurement values for chiller outputs
Cooling-
water
flow
Chilled-
water return
temp.
Cooling-
water supply
temp.
Steam
supply
temp.
Condensate
return
temp.
Steam
flow
Chilled-
water supply
temp.
Cooling
load
Test F1 F6 T20 T32 T22 T23 F2 T21 Q cooling COP
No. m 3 /h kg/s
o C
o C
o C
o C kg/h
o C kW
1 2.02 1.45 13.88 30.72 164.33 99.31 27.42 6.40 18.91 1.04
2 2.01 1.45 14.12 30.71 164.57 99.33 25.90 6.74 17.61 1.01
3 2.01 1.45 13.26 30.56 164.68 99.31 23.10 6.36 16.12 0.99
4 2.00 1.46 12.85 30.35 158.63 99.23 20.60 6.57 15.25 1.01
5 2.02 1.46 11.67 30.38 153.17 92.32 19.44 6.15 13.59 1.03
6 2.01 1.46 10.41 30.96 150.30 76.34 15.56 5.72 10.69 0.99
7 2.02 1.46 10.54 30.40 140.34 69.46 12.04 7.09 8.14 0.87
8 2.02 1.46 9.60 30.51 138.72 61.58 9.96 6.89 6.40 0.80
9 2.04 1.47 8.02 30.24 133.03 54.07 5.30 6.39 3.86 0.67
52

Figure 2-16: Chiller performance under various load conditions
Coefficient of Performance (COP) .
1.3
1.2
1.1
1.0
0.9
0.8
0.7
0.6
0.5
0.4
Thermal COP
(measurement)
Overall COP
(measurement)
0 2 4 6 8 10 12 14 16 18 20
Thermal COP (measurement) Overall COP (measurement)
Figure 2-17: Chiller power consumption under various load conditions
Chiller power consumption (kW) .
2.0
1.6
1.2
0.8
0.4
0.0
Measured power consumption
for natural gas and hot water
Measured power consumption
for steam driven chiller
Actual cooling load (kW)
Rated power consumption for
natural gas driven chiller
0% 10% 20% 30% 40% 50% 60% 70% 80% 90% 100% 110%
Steam chiller
Natural gas & hot water chiller
Natural gas chiller
Actual load / design load
53

Figure 2-18: Comparison of chiller performance
Actual heat input (kW) .
20
16
12
8
4
0
Steam-driven chiller
Natural-gas-driven chiller
0 4 8 12 16 20
Steam-driven chiller Natural gas-driven chiller Actual cooling load (kW)
2.5 Further Information from Chiller Testing
The performance curves for the chiller at part-load conditions, resulting from varied chilled-water
return temperatures, have been plotted and compared with Broad’s published information in the
figures above. The chiller performance based on varying others of the primary operating condition
parameters in Table 2-4 has also been determined using the test approach indicated in subsections
2.3.1 and 2.3.2; the results are summarized in appendix 2.B. The data on chiller internal conditions at
the various operating conditions will be used and analyzed in the next two chapters based on a
comprehensive performance model.
54

3 Chiller Design and Performance Model
As described in the previous chapter, the chiller was tested at various operating conditions, and chiller
performance was calculated on the basis of the measurements. This chapter describes how we
developed a comprehensive performance model to further refine our understanding of the principles of
the chiller, to analyze the experimental data from the test program, to assist in equipment design, and
to evaluate the performance of various BCHP systems. This model is a set of equations consisting of
mass balances, energy balances, relations describing the heat and mass transfer, and equations for the
thermophysical properties of the working fluids.
The model can be solved when appropriate assumptions and a certain number of operating parameters
are assigned, so that the conditions – pressure, temperature, composition, and flow – at each point
within the chiller can be calculated. The model solutions have been used to evaluate the accuracy of
the measurement data and the test program. Heat and mass transfer correlations have been integrated
into the model so that the model cannot evaluate the chiller performance at design conditions only, but
at various off-design conditions also. For the manufacturer and equipment designer, the model can be
used to size the chiller components and to determine configurations; and for the building system
engineer and architect, the model can be used to predict chiller performance under various design
conditions in buildings.
3.1 Flow Diagram
On the basis of the schematic flow diagram in Figure 2-2, we constructed a simplified flow diagram
labeled with corresponding numbered state points as illustrated in Figure 3-1. Each state point is
represented by its pressure, temperature, composition, and flow rate. The bold lines represent the water
refrigerant (vapor/liquid); the other lines refer to the sorbent water-LiBr solutions. Table 3-1 describes
each state point.
55

Figure 3-1: Simplified flow diagram for chiller model
12
Refrigerant
combiner
13
Refrigerant
spray nozzle
17
Evaporator
Sorbent solution
Refrigerant
State points: pressure, temperature,
composition, flow of streams
entering/leaving chiller components
Equilibrium states: thermal and vapor
liquid equilibrium between streams
leaving a chiller component Cooling
water
14
16
18
BPHX
RP
Condenser
Solution
Splitter
2
1
42 43
Heat to
Absorber
32 31 42 41
11
SP
Chilled water CHWP Cooling water
Chiller
25
3
10
Steam
trap
9
91
24
53
HRHX
4
Absorber
54
LTHX
7
LTRG
(inside)
52
Heat to
LTRG
Low-temp. heat exchanger
5
19
23
LTRG (outside)
Solution
8 combiner
Solution
spray nozzle
92
Low-temp. regenerator
6
To HRHX
52
20 21
High-temp. heat exchanger
46
HTHX
22
43
Steam
51
High temp. regenerator
HTRG
from
Condenser
44 41 to
Absorber
45 47
CWP
Air City water
Cooling tower
In Figure 3-1, the model for each chiller component consists of equations representing mass balances
for water and LiBr, the energy balance, the working fluids property relations, and the heat and mass
transfer relations involving the state point conditions of the streams entering and leaving the
component. Once these equations are assembled for all the chiller components and solved, all the state
point conditions of the chiller in terms of temperature, composition, pressure, flow rate, and other
thermodynamic properties will have been determined. The state point conditions of the water
refrigerant and the water-LiBr sorbent solutions can be plotted on a Dűhring diagram as illustrated in
Figure 3-2. Such a plot lacks only the flow quantities to serve as a complete description of the state
points throughout the chiller.
56

Table 3-1: Chiller model state point descriptions
No. Stream description No. Stream description
1 Sorbent solution in the absorber 2 Sorbent solution leaving the solution pump
3 Sorbent solution entering the LTHX 4 Solution leaving the LTHX and entering the HRHX
5 Sorbent solution entering the LTRG 6 Sorbent solution leaving the LTRG and entering the LTHX
7 Sorbent solution leaving the LTHX 8 Sorbent solution after solution combiner
9 Sorbent solution passing spray nozzles 10 Refrigerant vapor leaving the LTRG into the condenser
11 Refrigerant liquid leaving condenser 12 Refrigerant liquid leaving the BPHX
13 Refrigerant liquid after refrigerant combiner 14 Refrigerant leaving spray nozzles
15 Empty 16 Refrigerant liquid entering refrigerant pump
17 Refrigerant liquid leaving refrigerant pump 18 Refrigerant vapor leaving evaporator
19 Sorbent solution entering the HTHX 20 Sorbent solution entering the HTRG
21 Sorbent solution leaving the HTRG 22 Sorbent solution leaving the HTHX
23 Refrigerant vapor leaving the HTRG 24 Refrigerant leaving the LTRG
25 Refrigerant entering the condenser
31 Chilled-water return 32 Chilled-water supply
41 Cooling-water supply 42 Cooling-water leaving absorber
43 Cooling-water return 44 Cooling-water entering cooling water pump
45 Ambient air entering cooling tower 46 Exhaust air leaving cooling tower
51 Steam supply entering the HTRG 52 Condensate leaving the HTRG
53 Condensate leaving the HRHX 54 Condensate leaving steam trap
91 Refrigerant vapor after the spay nozzle 92 Sorbent solution after the spray nozzle
3.2 Dűhring Chart Representation
On the basis of model solutions, Figure 3-3 shows an absorption cycle at design condition with state
points indicated on the Dűhring chart, which visualizes the absorption cycle and associated design
parameters. The ordinate of this plot is the equilibrium vapor pressure of water (kPa), and the abscissa
is corresponding temperature ( o C). The inclined line on the left of the plot represents vapor pressuretemperature
relation for the water refrigerant. The parallel lines within the plot represent the water
vapor pressure of the sorbent solution at various concentrations and temperature. A crystallization line
is located at the bottom of the chart. If the state point of solution drops below this line, sorbent
solution will tend to deposit LiBr solid crystals.
The Dűhring chart is a tool to rapidly perform a number of checks on the measurement data or model
solutions. In such a plot, many design parameters can be illustrated, such as the heat rejection
temperatures, solution concentrations, equilibrium pressures, and pinch point of each heat transfer
component. In Figure 3-2, the connected bold lines in the middle of the figure are the water-LiBr
sorbent solution, and the bold dash lines represent the refrigerant. The lines are connected to form two
complete cycles: the sorbent solution cycle and the refrigerant cycle. The state points indicated on the
57

cycles are identical to those used in the flow diagram in Figure 3-2. The major chiller components are
indicated on the diagram by dashed ellipses.
In the Dűhring chart, a pinch point is one of the notable features of the heat exchanger. The
temperature pinch point is the point of minimum temperature difference between the fluids involved in
the heat transfer process. The pinch point usually occurs at either the inlet or the outlet of the heat
exchanger. Small temperature differences at the pinch point require large heat transfer areas in the
exchanger.
Figure 3-2: Dűhring chart at design condition
Evaporator
20
T31
Steam, cooling water, and chilled water
Water-LiBr sorbent solution
Refrigerant
Condenser
T11
T43
T42 40
T41
30
Pinch point
REFRIGERANT TEMPERATURE, C
50
60
57.39 %
70
61.68 %
80
90
0
30%
40%
T3210
evaporator
1
5
1 7
Pl
T18 0
91 22
10 20 30 40 50 60 70 80 90 100 110 120 130 140 150 160 170 180
Pinch point
for absorber
Absorber
Pinch point
Condenser
LTHX
4
Pinch point for LTHX
Pinch point for HTHX
o
LTRG
5
T53
110
T54
6
100
T1 T91 T7 T4 T5 T6
T20
T21
T22
50%
HTHX
57.39 %
Pinch point
for LTRG
60%
HRHX
120
70%
62.81 %
o
SOLUTION TEMPERATURE, C
Equlibrium Chart for A queous Lithium Bromide Solutions
20
HTRG
T52
21
T51
Ph
Pinch point
for HTRG
Pm
200
150
100
50
40
30
20
10
5
4
3
2
58
SATUATION PRESSURE (P), kPa

3.3 T-Q Diagram
The temperature heat (T-Q) diagram is another tool to indicate the pinch points of each heat transfer
component. The T-Q diagram for heat transfer components in this chiller are illustrated in Figure 3-3.
The ordinate is the stream temperatures along the length of the heat exchanger. The abscissa is the
quantity of heat transferred between the streams up to any given point along the length of the
exchanger.
Figure 3-3: T-Q diagram for the heat transfer components
o
C
160
140
120
100
80
60
40
20
22
1
HTHX
51
21
20
HTRG
20
LTRG
52
23
0
0 5 10 15 20 25 30
7
1
6
5
4
53
LTHX HRHX
3.4 Calculation Procedure
5
43
35
24
Condenser
9
11
42
Absorber
31
18
32
14
40 45 50 55 60 65 70 75 80 85 kW
1
41
Evaporator
This steady-state chiller model is a set of nonlinear algebraic equations, described in subsection 3.1.1
above, programmed in the Engineering Equation Solver (EES), which relate state-point conditions –
pressure, temperature, composition, and flow rates, and equipment design parameters throughout the
chiller. The EES equations of each component with equation type, including mass balance, energy
balance, heat transfer equations, thermal property functions, phase equilibrium equations, and the
assumptions, are annotated in appendix 3A.
The procedure for the EES calculation is straightforward: first, the algebraic equations are entered into
EES. Enough state-point conditions are entered so that the number of equations is equal to the number
of the remaining, unknown state-point conditions. Reasonable estimates are entered for all these
59

unknown conditions. The properties, such as the enthalpy and equilibrium functions, are related to the
pressure, temperature, composition, and vapor quality of water/steam and sorbent solutions by
equations in the EES. The EES then solves the equations by adjusting the estimates to reach a solution
of the equations. The chiller model comprises 416 variables, state-point operating conditions, and 409
equations expressing basic engineering principles. The equations can be solved when seven variable
inputs are provided. Once the model has been solved, the conditions of each state point in the flow
diagram are known. The solutions can be used to check the measurement data from the test program.
3.4.1 Mass Balance
For each chiller component, the steady-state total mass balance equation for refrigerant, chilled water,
cooling water, steam and condensate, and LiBr can be expressed as:
where, m& = mass flow rate.
∑ m& in = ∑ m&
out
.
(Equation 3 - 1)
In equation 3-1, subscripts in and out mean the streams entering and leaving each of the chiller
components. The mass balance of LiBr associated with absorption and regeneration processes can be
expressed in the following equation:
∑
Solution
∑
m& ∗ x = m&
∗ x
(Equation 3 - 2)
in
in
Solution
where, x = the weight concentration of the water-LiBr sorbent solution.
3.4.2 Energy Balance
out
out
As the basic format of energy, heat, and external work associated with the fluids are observed when
they cross the boundaries of each chiller component. In the chiller, except for three pumps, no work is
involved in the components. The steady-state energy balance for each chiller component is expressed
in the equation below:
Q& −W& shaft + ∑ m&
jh
j = 0
(Equation 3 - 3)
where Q& is the quantity of heat transfer to or from the system; Wshaft & is the quantity of shaft work
done by the system; m& is the mass flow of each stream; and h is the enthalpy of each stream.
60

For instance, the cycle involving with absorb process, the energy balance using equation 3-3 can be
written in the following equation:
Q + m&
h + m&
h = m&
h
(Equation 3 - 4)
water
water
conc
conc
dilute
dilute
3.4.3 Thermodynamic Property and Equilibrium Relations
Fluid properties are used widely in the model. In EES, the states and properties of specific fluid can be
related by internal functions. For pure water and steam, the following functions are often used to
provide equilibrium states.
Thermal property relation:
j
( water steam,
T = T P P )
h = enthalpy / , =
(Equation 3 - 5)
Phase equilibrium for saturated liquid and vapor in equilibrium:
j
( water / steam P Pj
)
( water / steam T Tj
)
( water steam,
h = h P P )
j
T = T _ SAT
, =
(Equation 3 - 6)
P = P _ SAT
, =
(Equation 3 - 7)
j
q = quality / , =
(Equation 3 - 8)
j
For the cooling-tower model, moist air properties can be evaluated by the following relations in EES:
3.4.4 Heat Transfer Models
j
= Enthalpy(Air
, P = P , T = T , R = RH )
(Equation 3 - 9)
h j
H2O j j
j
j
= Wetbulb(Air
, T = T , P = P , R = RH ) (Equation 3 - 10)
Tj -wb
H2O j j
j
RH =
, T = T , P = P , w = w )
(Equation 3 - 11)
j
Relhum(AirH2O j j
j
Heat transfer, in some components, coupled with mass transfer, occurs throughout the absorption
chiller. Although a full understanding of both the heat and mass transfer process in the absorber is
necessary and critical, it is usually more convenient to analyze them respectively with heat transfer
study in the first step and coupled mass transfer next. This allows us to handle the problems from the
simple to the complex.
j
61

In the model, the UA-LMTD (UA-log mean temperature difference) values are used to evaluate the
following heat transfer components: absorber, evaporator, condenser, HTRG, LTRG, HRHX, and
BPHX.
In the log-mean temperature difference (UA-LTMD) approach, finding the product of the overall heat
transfer coefficient and the heat exchanger surface area is convenient for specifying the size and
performance of a heat exchanger.
Q
UA = (Equation 3 - 12)
∆T
LMTD
In equation (3-12), Q is the quantity of heat transferred in the components. ∆ TLMTD
is the log-mean
temperature difference; it is expressed below as:
( T − T ) − ( T − T )
hot,
in cold , out hot,
out cold , in
∆ TLMTD
=
(Equation 3 - 13)
Thot,
in − Tcold
, out
ln
T
hot,
out
− T
cold , in
In equation (3-13), subscripts hot and cold refer to the hot and cold streams, respectively; the
subscripts in and out refer to the inlet and outlet of a stream.
The other two minor heat exchangers (HTHX and LTHX) are evaluated by the heat transfer
effectiveness method. For effectiveness-type heat exchanger models, the effectiveness, ε , is defined
as the ratio of the actual heat transfer, Q actual , to the maximum potential heat transfer, Q max , below:
Q
( Tcold
, out − Tcold
, in )
( T − T )
or
( Thot,
in − Tcold
, out )
( T − T )
actual ε = =
(Equation 3 - 14)
Qmax
hot,
in cold , in
hot,
in cold , in
The definition of effectiveness in terms only of the temperatures makes it a convenient heat exchanger
performance parameter.
3.4.5 Overall Heat Transfer Coefficient Model
In practice, the overall heat transfer coefficient, U, is not a constant variable in describing the heat
exchanger but is a function of flow rate, temperature, pressure, and other properties. Physical
information on the heat exchanger configuration and the characteristics of flow in and out of the heat
exchanger must be known to calculate the overall heat transfer coefficient and heat exchanger surface
62

area. Table 3-3 summarizes heat and mass transfer correlations for the various chiller components. The
values of U in the HTRG adapt for different heating media as listed in the table.
Table 3-2: Physical features of heat and mass transfer components
Steam
Chiller Tube bank Surface Inside
Component Material Type Treatment Media Process Media Process
Evaporator Copper Spiral tube Grooved Refrigerant Evaporation Chilled water Convection
Absorber Copper Spiral tube Smooth LiBr Convection Cooling water Convection
Condenser Copper Spiral tube Smooth Refrigerant Condensation Cooling water Convection
LTRG Copper Straight tube Grooved LiBr Boiling Water vapor Condensation
HTRG Copper Spiral tube Grooved LiBr Boiling Steam Condensation
Hot water HTRG Copper Straight tube Grooved LiBr Boiling Hot water Convection
Natural gas
HTRG Steel Comb. chamber Grooved LiBr Boiling Combustion gases Radiation
Steel Straight tube Grooved LiBr Boiling Combustion gases Convection
Exhaust gas HTRG Steel Straight tube Grooved LiBr Boiling Exhaust gases Convection
The overall heat transfer coefficient calculated for circular tube is a function of two heat-transfer
coefficients, hi and h o :
UA
overall
=
1
h A
in
in
1
ln out +
2πkL
( d d )
in
+
h
out
1
A
out
(Equation 3 - 15)
The subscripts in and out refer to the inside and outside of the tube. Ain refers to the inside surface and
Aout refers to the outside surface of the tube. In this model, the heat transfer resistance of the tube wall
is neglected. The area difference between the tube inside surface and outside surface is very small, so
the outside surface area of the tube is used as heat exchanger area A. Heat transfer coefficients
hin and hout can be calculated by the empirical equations in Table 3-3 that relate to steam states, fluid
properties, and physical configurations.
Heat transfer areas, A, sometimes are not constant when partial load operation is considered. For
instance, concentrated solution spray from nozzles on the surface of the absorber tube bank and the
refrigerant from nozzles on the surface of the evaporator tube banks may not cover the tubes at low
flow conditions. In this model all the heat transfer areas, A, are assumed to be constants.
63

Table 3-3: Heat and mass transfer correlations used in the performance model
Component Reference Process Equation Comment
Evaporator
Absorber
Condenser
HTRG
LTRG
Cooling
Tower
HTRG
HTRG
HTRG
Chun and
Seban
Dittus and
Boelter
Water film on
2
1 3
⎛ v ⎞
hevp− film⎜
⎟ 3
evp−
film evp−
tube surface ( ) ( ) 65 . 0
4 . 0
3 −
= . 8×
10 Re Pr
Chilled water in
tube
Vliet et al. LiBr solution film
on tube surface
Dittus and
Boelter
Cooling water in
tube
Vliet et al. Vapor absorption
rate on tube
surface
Nagaoka et al. Vapor absorption
coefficient on
tube surface
Kern D. Q. Condensation
film on tube
surface
Dittus and
Boelter
Jakob and
Hawkins
Cooling water in
tube
Nucleate boiling
Kern D. Q. Condensation
film in tube
Jakob and
Hawkins
h
⎜ g k ⎟
⎝ ⋅ ⎠
d
3 film
evp−chw
30
0.
8 0.
3
= 0. 023Reevp−chw
Prevp−chw
k30
− 1
3
0.
46 1.
5 91 91
− 0.
30 Re − 91 2 ⎟
91
⎟
⎛ µ Γ ⎞
h =
k ⎜
abs film
abs film ⎜ ρ g
h
d
abs−
cw 40 0.
8 0.
4
= . 023Reabs−cw
Prabs−
k40
0 cw
⎝
⎠
For vertical tube
Equation summarized for the ¾-
inch (19 mm) smooth tube bank
with Octhyl Alcohol surfactant
0.
46 q&
abs
m&
abs−
vapor = 0. 30 Reabs−
film
h fg
Equation summarized for the ¾-
inch (19 mm) smooth tube bank
with Octhyl Alcohol surfactant
k abs = 29.571Γ
91 + 0.405
Equation summarized for the ¾-
inch (19 mm) smooth tube bank
with Octhyl Alcohol surfactant
hcond− film
h
2 ⎛ v ⎞
⎜ 11 ⎟
⎜ 3 ⎟
⎝ k11g
⎠
d
1 3
= 1.
51Re
−1
3
11
cond − cw
k49
49 0.
8 0.
4
= 0. 023 Recond
−cw
Prcond
−cw
htrg −nucleate
51 21
0.
4
3⎛
p ⎞ h
⎜
⎟
p0
1
on tube surface h = 1042(
T −T
)
Nucleate boiling
Kern D. Q. Condensation
film in tube
Jakob and
Hawkins
Dittus and
Boelter
Jakob and
Hawkins
hhtrg− film
2 ⎛ v ⎞
⎜ 51 ⎟
⎜ 3 ⎟
⎝ k51g
⎠
1 3
= 1.
51Re
1
on tube surface h = ( T − T )
Nucleate boiling
ltrg −nucleate
⎝
−1
3
51
0.
4
3
1042 24 6 ⎟
0
⎟
⎛ p ⎞ m
⎜
p
2
1 3
−1
3
4
24
24
− 1.
51
3
⎟
24
24
⎟
⎛ v ⎞ ⎛ Γ ⎞
h ⎜ ⎟ = ⎜
ltrg film⎜
k g ⎟ µ
NTU =
⎝
ln
⎠
⎝
( hs43
− h46
) − ( hs44
− h45
)
( ( hs − h − δ ) ( hs − h − δ ) )
43
46
(h
46
ct
- h
1
on tube surface h = 1042(
T −T
)
Hot water in tube
Nucleate boiling
Hottel, et al. Combustion
process in
chamber
Hausen Exhaust gas in
tube
Jakob and
Hawkins
htrg −nucleate
h
htrg −hw
k
55
d
55
=
51
21
⎝
45
3
)
44
⎠
⎠
⎠
45
0.
4
⎛ p ⎞ h
⎜
⎟
⎝ p0
⎠
0.
8 0.
3
0. 023Rehtrg
−hw
Prhtrg
−hw
1
on tube surface h = 1042(
T −T
)
Nucleate boiling
J. P. Holman Exhaust gas in
tube
htrg −nucleate
Q
h
51
21
3
0.
4
⎛ p ⎞ h
⎜
⎟
⎝ p0
⎠
4
4
( T ) σT
α ( T ) σ
radiation = ε g g g − g w Tw
Acc
htrg − gas
k
d 0.
0668
= 3.
66 +
1+
0.
04
1
on tube surface h = ( T − T )
htrg −nucleate
h
htrg −exgas
k
55
d
55
( )
[ ( ) ] 3 2
d L Red
Pr
d L Re Pr
1042 21
d
0.
4
3
⎟
0
⎟
g
⎛ p ⎞ h
⎜
p
=
0.
8
0.
4
0. 023Rehtrg
−exgas
Prhtrg
−exgas
⎝
⎠
ct
Steam as heating medium
See paper
Hot water as heating medium
Natural gas as heating medium
Laminar flow
Exhaust as heating medium
Turbulent flow
64

3.4.6 Mass Transfer Models
The process of mass transfer is complicated by coupled heat transfer and by the properties of the
working fluids. Mass transfer occurs in the absorber, the HTRG, and the LTRG. The boiling processes
in the LTRG and HTRG, however, mix the solution well, and therefore the mass transfer effects are
minimized. In the model only the mass transfer in the absorber is considered. Numerous modeling and
experimental studies found that the absorption process is controlled by the mass transfer resistance on
the liquid side. This is because the refrigerant vapor absorbed at the liquid interface transfers slowly
into the bulk of the liquid. The absorption of additional refrigerant is inhibited. The energy released at
the liquid interface causes an increase in temperature there, and this energy must also transfer through
the liquid film to the bulk of the liquid.
A coupled mass and heat transfer model for the absorber is developed on the basis of a correlation
given by Cosenza and Vliet [4]. They found that the mass transfer rate is a linear relation to the heat
transfer rate. They also observed this relation by experiment on a ¾ -inch (19mm) tube bank. The
details of their studies and how this relation is implemented in the modeling will be discussed further
in appendix 3A.
3.4.7 Model Assumptions
The following assumptions are employed to properly represent the absorption cycle:
• The control of streams between components allows only all liquid or all gaseous flows. The
system operates at steady-state conditions. There is no accumulation/depletion of mass or
energy at any point within the system.
• The overall system is considered a three-pressure system:
o The high-pressure, P h , is determined by the equilibrium water vapor pressure and the
temperature entering the condenser. The pressures in the HTRG and in the heating tubes
of the LTRG are at this high pressure.
o The intermediate-pressure, P m , is determined by the equilibrium water vapor pressure
and temperature of the refrigerant leaving the condenser. The pressure of sorbent solution
in the LTRG is at this intermediate pressure.
o The low-pressure, P l , is determined by the equilibrium of water vapor pressure and
temperature of the refrigerant in the evaporator. The sorbent solution in the absorber is at
65

this same pressure. The pressure difference due to flow from the evaporator to the
absorber is small enough to be neglected. (Herold, Radermacher, and Klein consider the
equilibrium pressure of refrigerant at state point 18 as representing the low pressure in the
absorber and the evaporator.)
• The dilute solution leaving the absorber is in phase equilibrium at the same water vapor
pressure as the refrigerant from the evaporator.
• The temperatures of superheated vapors leaving two regenerators have the same temperature
as the concentrated solution leaving the HTRG and the LTRG. (Koeppel, Klein, and Mitchell
took the following point of view, that refrigerant vapor leaving the regenerator has the
equilibrium temperature of the weak solution at regenerator pressure.)
• The steam input is saturated vapor, and the condensate after the steam trap is saturated liquid.
• There is no liquid carryover between the evaporator and the absorber.
• Flow restrictors, such as expansion valves, spray nozzles, and the steam trap are adiabatic.
• Pump work is isentropic. There are no pressure changes except for flow restrictors and pumps.
Flow head losses in the piping system are negligible.
• There are no convection and radiation heat losses through surfaces to ambient.
3.5 Model Steps
Figure 3-4 illustrates the steps in the use of the performance model to deal with the chiller
performance for design and off-design conditions. These steps are listed below:
1. Estimate the chiller cooling capacity, COP, and heat source conditions theoretically on the
basis of desired sorbent composition, chilled water, and cooling-water conditions.
2. Estimate the UA values (heat transfer areas) of the nine heat transfer components in a design
model when the design operating conditions and the approach temperatures are taken into the
design model. The heat transfer areas can be estimated for design condition.
3. Construct a performance model for design and off-design conditions, the actual U (heat
transfer coefficient) and A (heat transfer surface area) of five major heat exchangers can be
calculated from the chiller physical configurations (from the manufacturer) and heat and mass
transfer correlations. These heat transfer correlations are corrected by the comparison of the
actual U, A, and the solution from the design model.
4. Analyze the accuracy of the measurements and validate the model at design and off-design
conditions. The corrected UA values from step 2 are used to construct a performance model
66

that can predict the chiller performance under design and off-design operation conditions. The
analytical method and results in this last step will be discussed further in chapter 4.
Figure 3-4: Steps in the use of the performance model
The design model and the performance model both use the same equations and assumptions; they are
identical except for the structure of calculation. The design model estimates UA values of the nine heat
transfer components on the basis of desired chiller performance for the design condition; the structure
of a design model is illustrated in Figure 3-5. In the design model, appropriate UA values are
determined from assumed pinch-point temperatures for heat transfer components.
The performance model uses the U and A values by detailed heat transfer coefficient correlations from
the literature and the chiller information from the manufacturer. The U and A values are corrected on
the basis of initial estimation of UA values from the design model. The structure of the performance
model is illustrated in Figure 3-6. In the performance model, the UA solutions of different heat
transfer components replace the pinch-point temperatures in the design model. Up to now, both the
design model and the performance model represent the chiller performance under design conditions.
67

Off-design conditions can be modeled when the physical-mathematical model of heat and mass
transfer characteristics are described in the performance model.
Figure 3-5: Structure of the design model Figure 3-6: Structure of performance model
Once the heat and mass transfer characteristics of the chiller are described in the performance model,
the model can be used to represent the off-design conditions as the chiller tested on the basis of the test
program. The simulation outputs are used to compare with the test data. Discrepancies are acquired to
identify measurement that may be inaccurate. The performance model will be validated by minimizing
the deviations between the model solutions and the test data. An experimental data-driven model
approach will be used to tune up the performance model, particularly those uncertainties that exist in
the model assumptions.
68

4 Model-based Experimental Data Analysis
The comprehensive model developed for the chiller in chapter 3 has been used to analyze the
experimental data from the test program. The computational model has been used to calculate all
chiller internal working conditions from a limited number of measurements. The discrepancies
between the measurements and the model calculations have been minimized by adjusting the model
assumptions. The discrepancies between the measurements and the model solutions are introduced
mainly by the following:
• inaccurate stream flow temperature measurements from sensors mounted on the external pipe
surface
• fluctuating measurements of steam flow due to periodic feedwater addition to the boiler
• imprecise cooling water flow measurements due to space limitations in mounting the flow
sensor
• inaccurate assumptions regarding the quality of the refrigerant flow from various chiller
components
• inaccurate values of heat transfer coefficients calculated from available correlations
The absorption cycle on a Dűhring diagram for each test has been plotted on the basis of the model
calculations. The variation trends of temperature, pressure, and composition of critical state points
have been summarized symmetrically on the Dűhring diagram and other plots. On the basis of the
model analysis results, the strategies to improve the chiller performance (particularly at the partial load
conditions) have been devised and the model has been validated to calculate chiller performance under
various operating conditions.
4.1 Analytical Method
Figure 4-1 illustrates the analytical method used in the model-based data analysis process. First, when
a set of steady-state test data is available, the outlier data are removed. Second, the experimental data
from the chiller test are averaged. Third, the seven operating parameters from the averaged data are
used as the model input to solve the model. The seven input parameters for the model are listed in
Table 4-1 in bold; the values of the measured values are listed under the “sensor” column. The values
of 11 measured operating parameters used for checking model calculations are also listed in Table 4-1.
In the table, two sets of numbering systems are used, one for the sensor, the values under the “sensor”
column being the average value of steady-state data, and the other numbering system for the model
69

state points, the calculated values being those listed in the “model” columns. The measured values
from the chiller tests are compared with the same operating parameters calculated by the engineering
equation solver, EES, from the chiller model.
Figure 4-1: Data analytical procedure flow diagram
The differences between the calculated and measured values of the 11 operating parameters are
weighted and summed in a statistical procedure to arrive at a measure of the data accuracy and model
validity. If this measure is unsatisfactory, the data are examined for possible errors and discrepancies,
and the model assumptions are adjusted to reduce the discrepancies in calculated and measured
operating conditions. If the statistical measure is satisfactory, the calculated and test data are plotted on
a Dűhring chart.
This analytical method is used throughout the data analysis process for all test data. The statistical
analysis procedure and the results of the analysis are presented in the following sections.
4.1.1 Statistical Analysis Procedure
The statistical analysis procedure introduced in Figure 4-1 is used to evaluate the deviations between
the model calculations and the test measurements. The statistical model is based on the following
equation:
2
n ⎛ ∆X
⎞ n
∑ ⎜
⎟
1
=
⎝ X n
σ
⎠ , (Equation 4 - 1)
n −1
70

Table 4-1: Measured values and model calculations for 100% and 55% of design load conditions
Temperature
Flow
Pressure
Power
Calculation
Stream name Label 100% 55%
Condensate after HTRG * T11 T52
Solution in absorber * T12 T1
Solution entering HRHX * T13 T4
Solution leaving HRHX * T14 T5
Cooling-water after absorber * T15 T42
Cooling-water after condenser * T16 T43
Solution leaving LTRG * T17 T6
Refrigerant after condenser * T18 T11
Solution entering HTRG * T19 T20
Chilled-water return T20 T31
Chilled-water supply T21 T32
Steam input T22 T51
Condensate return * T23 T54
Cooling-water supply T32 T41
Solution leaving HTRG * T33 T21
Sensor Model Unit Sensor Model
Absolute
Deviation
Relative
Deviation Weight Sensor Model
Absolute
Deviation
Relative
Deviation Weight
o C 157.5 157.5 0.0 0.0% 10% 133.7 133.7 0.0 0.0% 10%
o C 36.6 36.1 0.5 1.5% 8% 33.8 33.4 0.4 1.1% 8%
o C 75.1 75.1 0.0 0.0% 10% 66.1 66.1 0.0 0.0% 10%
o C 90.5 90.5 0.0 0.0% 10% 81.3 81.4 -0.1 -0.1% 10%
o C 37.7 36.1 1.6 4.3% 6% 33.8 33.8 0.0 0.0% 6%
o C 40.0 38.2 1.8 4.6% 6% 34.9 34.9 0.0 0.1% 6%
o C 93.3 92.7 0.6 0.6% 10% 79.6 79.6 0.0 0.1% 10%
o C 44.1 44.1 0.0 0.0% 20% 37.2 37.2 0.0 0.0% 20%
o C 129.3 128.8 0.5 0.4% 10% 113.8 113.5 0.3 0.3% 10%
o C 13.9 13.9 0.0 -0.1% 0% 10.5 10.6 0.0 -0.1% 0%
o C 6.4 6.4 0.0 0.8% 0% 7.1 7.1 0.0 0.5% 0%
o C 164.3 164.0 0.3 0.2% 0% 140.2 140.0 0.2 0.2% 0%
o C 99.3 100.0 -0.7 -0.7% 2% 69.0 70.2 -1.2 -1.7% 2%
o C 30.7 31.5 -0.8 -2.5% 0% 30.4 31.5 -1.1 -3.6% 0%
o C 153.5 155.8 -2.3 -1.5% 8% 131.7 132.8 -1.1 -0.9% 8%
Chilled-water supply F1 m31 kg/s 0.56 0.55 0.0 2.14% 0% 0.56 0.55 0.0 2.06% 0%
Steam supply F2 m51 kg/s 0.00713 0.00730 -0.00017 -2.33% 0% 0.00334 0.00385 -0.00050 -15.06% 0%
Cooling-water supply F6 m41 kg/s 1.45 1.15 0.3 20.64% 0% 1.46 1.15 0.3 21.09% 0%
Chilled water inlet P2 P31 kPa 54.2 84.5
Chilled water outlet P3 P32 kPa 133.3 164.1
Steam inlet P4 P51 kPa 709.8 683.1 26.7 3.8% 0% 370.4 361.2 9.2 2.5% 0%
Condensate after chiller P5 P54 kPa 101.8 101.3 0.5 0.5% 0% 101.3 101.3 0.0 0.0% 0%
Power for the chiller E1 kW 1.34 1.14
Power for the boiler E2 kW 20.02 10.96
Thermal COP 1.02 1.10 -0.08 -7.9% 0.86 0.91 -0.05 -5.9%
Cooling load 17.39 17.17 0.22 1.3% 8.07 7.89 0.18 2.2%
Overall deviation (σ )
5.0% 5.9%
Weighted deviation (σ ')
1.67% 0.48%
* The 11 measurements inside chiller to check with model calculations
71

where σ is an overall deviation; ∆X n is the deviation between the measured value of a chiller
operating condition value X n'
and the model calculated value X n :
∆ X = X −
n
n
X
n
'
, (Equation 4 - 2)
where the subscript “n” corresponds to each of the 11 test measurements listed in Table 4-1.
A weighted error, σ ' , has also been considered; the weight of each measurement is assigned on the
basis of its perceived accuracy:
∑ ⎟ ⎟
2
n ⎛
⎞
⎜ ⎛ ∆X
⎞ n
σ '= ε
⎜ ⎜
⎟
n *
(Equation 4 - 3)
1 ⎝ ⎝ X n ⎠ ⎠
The summation of weights of all the measurement conditions has been set at 1.
n
∑ ε n = 1
(Equation 4 - 4)
1
The weights of measured conditions are given on the basis of engineering judgment. Higher weights
are assigned to more accurate measurements. All the sensors used in the test program have been
calibrated as presented in chapter 2. The accuracy of the measurement of the sensors discussed in this
chapter is affected by external factors such as location and installation of the sensor. The effects of
these external factors can be minimized when the weights are assigned. If, for instance, a sensor shows
a small discrepancy consistently for all operating conditions, we can conclude that this measurement is
less affected by the external factors. This sensor is treated as an accurate sensor. By using the weights
in the model, the overall deviation between the model solutions and the measurements is less affected
by the uncertainties of the measurements.
4.1.2 Absorption Cycle at Design Condition
The model solution can best represent the absorption cycle at various loads. The model solutions of
major state points for a 100% design load condition are presented in Table 4-1; the solutions are
mapped in the Dűhring chart in Figure 4-2 to form a complete absorption cycle.
72

Figure 4-2: Absorption cycle at design load condition
T32
T31
13.9
20
Steam, cooling water, and chilled water
Water-LiBr sorbent solution
Refrigerant
Model solution
Measurement
+ 5 C
o
T43
38.2
T42 40
36.11
T41
31.5
30
T1 T91 T7 T4 T5 T6
T20
T21
36.1
T11
44.11
10
6.35
1
5 m 1 7
18 =0.007513 kg/s
Pl, 0.7665 kPa
T18 0
91 22
3.154 10 20 30 40 50 60 70 80 90 100 110 120 130 140 150 160 170 180
57.39 %
45.75
REFRIGERANT TEMPERATURE, C
50
T22
46.22
61.68 %
48.1
60
70
m 10 =0.003315 kg/s
1
6
m =0.09622 kg/s
m =0.04434 kg/s
4
o
75.1
80
4
99.97
100
95.26
90
0
30%
40%
m =0.04765 kg/s
T53
109.6
110
T54
5
6
90.65 92.7
50%
57.39 %
60%
120
m =0.0042 kg/s
24
m =0.04857 kg/s
70%
m =0.0073 kg/s
51
62.81 %
128.8
o
SOLUTION TEMPERATURE, C
Equlibrium Chart for A queous Lithium Bromide Solutions
21
20
21
m =0.04437 kg/s
T52
157.5
21
156.2
T51
164
200
150
100
Ph, 85.34 kPa
Pm, 9.155 kPa
The chart visualizes the sorbent solution cycle, the refrigerant cycle, and the conditions of steam,
condensate, cooling water, chilled water, and sorbent solutions at the major state points indicated in
Figure 4-2. In the figure, the measurements of the 11 temperature sensors inside the chiller are also
plotted. The discrepancies are displayed between the model solution and the measurements. The 11
dotted circles are centered at model solutions with a radius of 5 o C. This is a convenient way to
illustrate the discrepancies between the model solutions and the measurement values. The flow rates of
the sorbent solution and the refrigerant are directly labeled in the Dűhring chart above each stream
line.
Among the 11 measurements, the sorbent solution leaving the HTRG at state point 21 (T21) shows the
highest deviation to be about 2.3 o C, where the model solution gives a higher value than the
measurement. The cooling-water temperatures after the absorber (T42) and the condenser (T43) show
50
40
30
20
10
5
4
3
2
73
SATUATION PRESSURE (P), kPa

deviations of about 1.6 o C and 1.8 o C, respectively, where the actual measured values are even higher
if surface temperature measurement are used here. Other sensors agree with the model solutions
within ± 1
o C.
Figure 4-2 shows an average concentration of 60% between the concentrated and dilute water-LiBr
sorbent solutions. The concentration differences between the concentrated solutions and the dilute
solutions are 5.4% and 4.3% for the HTRG and LTRG, respectively. These data check well with the
design conditions Broad provided. The chiller was initially charged with a total of 65 kg sorbent
solution at 55%. The concentration of sorbent solution at the different state points depends on the
inventories of water in the water tray of the evaporator, the drain pan of the condenser, and the pipes.
We have estimated that, at design conditions, 5 kg of refrigerant water are held up in the reservoir of
the evaporator, the condenser, and the pipes. The average concentration of the sorbent solution is then
about 60%.
At off-design conditions, less refrigerant circulates in the chiller. The average sorbent solution in the
chiller may then be more dilute than at design conditions. The sorbent solution concentration changes
corresponding to the chiller load variations will be discussed in the next section.
The dilute sorbent solution flow ratio to the HTRG and to the LTRG is another key variable that
deserves closer consideration. At design condition, the model shows that the flow distribution ratio is
roughly 0.5, which means that the flow is equally distributed to each regenerator. Broad has
confirmed this result for the design condition. The model shows 17 kW cooling load, the same as the
measurements. The model, however, shows higher COP, about 1.10, than the measured value of 1.02.
4.1.3 Overall Deviation
An initial standard and weighted error is calculated by the statistical analysis procedure and then
model assumptions are adjusted one by one to reduce standard and weighted errors. New assumptions
are then applied to the nine test data sets. If consistent improvements are found for all nine tests, the
new assumption is adopted. This has proved effective in improving the model and in identifying ways
in which the chiller performance might be improved.
Overall deviations are calculated on the basis of the statistical analysis procedure (subsection 4.1.1) to
evaluate the model accuracy. Table 4-1 shows the results of the procedures for calculating the absolute
and weighted deviations. The absolute deviations between the two columns are calculated in the third
74

column. The relative deviation is calculated on the fourth column. The weights assigned to each of the
11 measurement points, the cooling capacity, and the COP are listed in the bottom of the table.
The 100% on the first row means the cooling capacity full load is 17.17 kW; partial-load percentage is
calculated by comparing its cooling capacity with this full-load capacity. The measurement data,
model solutions, and the analysis for the 55% cooling-load conditions are also listed in Table 4-2. At
design condition, the overall deviations between the measurements and the model solutions are about
5%, the weighted overall deviation is 1.67%. The major deviation is introduced by the steam flow
meter and the cooling-water flow meter. By using the weights, the systematic deviations introduced by
the two flow meters are deemphasized.
4.2 Model Analysis
Once the deviations between the measurements and the model solution for the design condition are
reduced satisfactorily, the model is used to analyze all the experimental data from the test program by
the same analytical method. The test data for the cooling-load variations are analyzed in the following
subsections.
4.2.1 Analysis of Cooling-Load Variation
In the cooling-load variation test, the chiller cooling load was changed in the model by adjusting the
chilled-water return temperature with a fixed chilled-water supply temperature and flow rate. The
cooling loads were varied in 9 steps from 100% to 35%. The test data were obtained using the same
method as described in chapter 2. As an example of off-design condition analysis, Table 4-1 also
shows the measured values, the model solutions, and the analysis results for a 55% design load
condition. Similarly, the absorption cycle of 55% design load condition is mapped in a Dűhring
diagram in Figure 4-3.
If we compare the 55% of design load with the 100% of design load condition in Figure 4-2, the
pressures in the HTRG, LTRG, and the evaporator all decrease significantly. The concentration of
dilute sorbent solution does not change appreciably, but the concentrated sorbent solutions from the
two regenerators become more dilute. The flow rates of total sorbent solution and refrigerant are
decreased. The dilute sorbent flow distribution ratio to the HTRG and the LTRG is about 0.6. At 55%
design load condition, less refrigerant circulates in the chiller; the average concentration of the dilute
sorbent solution in the chiller is estimated at 58%, which is lower than that of the design condition at
60%. The model calculated COP for the 55% design load condition is 0.91, which is higher than the
75

measured COP at 0.86. The overall deviation of the measurements and the model solutions is about
5.9%, and the weighted overall deviation is about 0.48%.
Figure 4-3: Dűhring chart at 55% design load condition
REFRIGERANT TEMPERATURE, C
60
70
o
80
90
0
30%
40%
50
T11
T43
10
T42 40
Pm, 6.352 kPa
T41
30
5
4
3
20
T31
2
T3210
5
0
10 20 30 40 50 60 70 80 90 100 110 120 130 140 150
1
Pl, 0.6871 kPa
160 170 180
m =0.04701 kg/s
18
Steam, cooling water, and chilled water
Water-LiBr sorbent solution
Refrigerant
Model solution
Measurement
T1 T22
T91 T7
4
56.89 % m =0.03305 kg/s
T54
m =0.01436 kg/s
10
6
59.47 % m =0.03162 kg/s
T4
100
50%
110
56.89 % m =0.04958 kg/s
60%
T6 T5 T20
20
120
m =0.02012 kg/s
59.29 % m =0.04756 kg/s
70%
T21
o
SOLUTION TEMPERATURE, C
Equlibrium Chart for A queous Lithium Bromide Solutions
23
21
m =0.03848 kg/s
51
T52 T51
Ph, 54.4 kPa
Other test data for the cooling-load conditions are analyzed using the same method. Figure 4-4 shows
the model output of the 9 steady-state tests at various load conditions. The trends of the major state
points illustrate the variation of composition, temperature, and equilibrium pressure with the change of
the operating load conditions. The Dűhring plot illustrates the conditions of the chiller under various
loads in one diagram, but it does not indicate the flow rate of each stream, the quantity of heat
transferred, or the heat transfer coefficients in each component. Other plots are needed to supplement
the Dűhring chart.
200
150
100
50
40
30
20
76
SATUATION PRESSURE (P), kPa

Figure 4-4: Absorption cycle variations with load changes
20
Steam, cooling water, and chilled water
Water-LiBr sorbent solution
Model solution for 100%, 83%,
67%, and 34% of design load
Chilled/cooling water
supply temperature set point
Point 4340
Point 41
30
REFRIGERANT TEMPERATURE, C
50
60
70
o
80
100
Point 54
90
0
30%
40%
Point 31
10
Point 32
5
0
10 20
Point 1
30 40
Point 7
Point 22
Point 5091 60 70 80 90 100 110 120 130 140 150 160 170
1
180
4.2.2 Performance Curve
Point 4
Point 5
50%
Point 6
110
60%
120
Point 20
70%
o
SOLUTION TEMPERATURE, C
Equlibrium Chart for A queous Lithium Bromide Solutions
Point 51
Point 21
Figure 4-5 shows that the chiller COP calculated by the model is, on average, higher than the
measurement by 8%. The reason for the discrepancy is the calculation of the heat input. The model
and the measurement share the same steam inlet pressure/temperature and the flow, but in the
measurement the condensate from the chiller is assumed to be saturated water at atmospheric pressure.
The model solution, however, shows that the condensate is partially vaporized when it leaves the
chiller above an 82% design load. In the model, the heat input is defined by the summation of the heat
transferred to the HTRG and the HRHX. The model, therefore, predicts lower quantity of heat
transferred to the chiller than the measurements. Below 82% load conditions, the model predicts
higher condensate return temperature than the measurements. Theoretically, the performance curve
calculated by the model is a better representation of the chiller performance.
200
150
100
50
40
30
20
10
5
4
3
2
77
SATUATION PRESSURE (P), kPa

Figure 4-5: Chiller performance curve under various load conditions
Coefficient of Performance (COP) .
1.30
1.20
1.10
1.00
0.90
0.80
0.70
0.60
0.50
0.40
Thermal COP
(model)
Overall COP
(measurement)
0 2 4 6 8 10 12 14 16 18 20
Thermal COP (measurement) Overall COP (measurement)
Thermal COP (model)
Thermal COP
(measurement)
Figure 4-6: Heat transfer load on each component under various load conditions
Heat transfer on chiller component (kW) .
32
28
24
20
16
12
8
4
Actual cooling load (kW)
0
HRHX
BPHX
0 2 4 6 8 10 12 14 16 18 20
Cooling tower Absorber Evaporator HTRG
LTRG Condenser HTHX LTHX
HRHX BPHX
Cooling tower
Absorber
Evaporator
HTRG
LTRG
HTHX
Condenser
LTHX
Actual cooling load (kW)
The quantity of heat transferred in each component in the chiller is illustrated in Figure 4-6. The load
in the 5 major heat transfer components is linearly related to the cooling load. The heat transferred on
the 4 minor heat recovery exchangers, the HTHX, LTHX, HRHX, and BPHX, are relatively constant.
The heat transfer in the BPHX is negligible.
78

4.2.3 Flow Rate Variations
The steam flow rate is an input parameter to the chiller model; the flow rate has been calculated on the
basis of the power input measurement to the steam boiler. In the chiller operation, the outlier data are
mainly introduced by the steam flow meter. The readings of the steam flow meter may be zero or very
low when the feedwater pump operates to supply feedwater to the boiler; these low readings reduce
the average steam flow measurement below the actual value. In this case, the measured steam flow
meter is not very reliable, so the actual steam flow is calculated using the power measurements to the
steam boiler. The calculated steam flow agrees well with the condensate return measured after the
chiller. The comparison of the calculated steam flow and the measurement readings under various
operation conditions are presented in Figure 4-7. The manufacturer has calibrated the steam flow
orifice meter at design conditions with 700 kPa saturated steam.
Figure 4-7: Steam supply flow rate under various load conditions
Steam flow rate (kg/s)
0.008
0.007
0.006
0.005
0.004
0.003
0.002
0.001
0
Steam flow calculated
from power
Measured steam flow
Measured steam flow
(corrected by the desnsity)
0 2 4 6 8 10 12 14 16 18 20
Steam flow calculated from power measurement
Measured steam flow
Measured steam flow (corrected)
Actual cooling load (kW)
Figure 4-8 shows the flow rates of the dilute sorbent solution from the absorber to the HTRG and
LTRG at various load conditions. The dilute solution flow in the HTRG remains relatively constant.
The sorbent solution flow rates to the LTRG decrease with the drop of cooling loads. This result is
consistent with the chiller control principle that the variable frequency solution pump maintains the
sorbent solution level in the HTRG. Figure 4-9 shows the sorbent solution split ratio that is defined by
the dilute sorbent solution flow to the HTRG over the total solution flow from the absorber. The
chiller does not control sorbent solution distribution ratio, the ratio is preset roughly at 0.5 at design
79

load condition by predetermining the pipe diameters of each sorbent stream. In this case, when the
cooling load decreases, the pressure drops in the HTRG faster than in the LTRG, so more sorbent
solution flows into the HTRG than the LTRG, and the sorbent solution split ratio increases the value
above 0.5.
Figure 4-8: Sorbent solution flow rate under various load conditions
Sorbent solution flow, kg/s .
0.12
0.1
0.08
0.06
0.04
0.02
0
Sorbent solution
to the LTRG
Sorbent solution
from the absorber
Sorbent solution
to the HTRG
0 2 4 6 8 10 12 14 16 18 20
Sorbent solution from absorber
Sorbent solution to HTRG
Sorbent solution to LTRG
Actual cooling load (kW)
Figure 4-9: Sorbent solution split ratio under various load conditions
Dilute sorbent solution distribution ratio, R .
1
0.8
0.6
0.4
0.2
0
Split ratio, R
0 2 4 6 8 10 12 14 16 18 20
Sorbent solution distribution ratio
Actual cooling load (kW)
80

Figure 4-10 shows that the refrigerant flow vaporized in the evaporator is proportional to the coolingload
condition. The HTRG consistently generates more refrigerant than the LTRG. This result agrees
well with the chiller control principle in which, under lower load conditions, the refrigerant level in the
water tray of the evaporator drops, and the refrigerant pump is on/off less frequently than at the higher
load conditions.
Figure 4-10: Refrigerant regeneration rate under various load conditions
Refrigerant flow (kg/s)
0.01
0.008
0.006
0.004
0.002
0
Refrigerant vaporized
in the evaporator
Refrigerant produced
from the HTRG
Refrigerant produced
from the LTRG
0 2 4 6 8 10 12 14 16 18 20
Referigerant from LTRG
Refrigerant vaporizaed in evaporator
Refrigerant from HTRG
Actual cooling load (kW)
4.2.4 Temperature Variations
In Figure 4-11, the refrigerant vaporization temperature in the evaporator is plotted again to illustrate
state point 18 more clearly. The refrigerant vaporization temperature is in equilibrium with the vapor
pressure in the absorber. The refrigerant has a higher vaporization temperature, around 3.2 o C at
design conditions; this temperature drops to 0.5 o C at 34% design load conditions. The result agrees
with the chiller control principle that when the cooling load drops too low, ice may form in the
evaporator. Ice formation affects chiller operation by blocking the spray nozzles; the problem may be
solved automatically when the chiller stops for a short while. To avoid the hazard of ice formation in
the recirculation pump, an electrical heater installed in the evaporator is turned on to protect the
refrigerant pump from freezing.
81

Figure 4-11: Refrigerant vaporization temperature under various load conditions
Vaporization temperature, T18, ( o C ) .
4
3.5
3
2.5
2
1.5
1
0.5
0
T18
0 2 4 6 8 10 12 14 16 18 20
Vaporization temperature in the evaporator
Actual cooling load (kW)
Figure 4-12: Sorbent solution composition changes under various load conditions
Sorbent solution concentration (%) .
64
62
60
58
56
54
4.2.5 Composition Variations
Sorbent from HTRG
Sorbent from Absorber
Sorbent from LTRG
0 2 4 6 8 10 12 14 16 18 20
Sorbent form absorber
Sorbent from HTRG
Sorbent from LTRG
Actual cooling load (kW)
Figure 4-12 illustrates the composition changes of the sorbent solutions with the variations of the load
conditions. The sorbent solution leaving the HTRG shows higher concentration than that of the LTRG
at design load conditions. With the load decrease, the concentration of sorbent leaving the LTRG
becomes higher than that of the HTRG. This result is introduced by sorbent solution split ratio changes
from design load to partial load condition. The concentrations of the dilute solution do not change
82

appreciably. The average concentration of the sorbent solution approaches 57% from the design load
condition to lower load condition. This result checks well with the design parameter from Broad
indicating when the chiller is off, the concentration of the dilute solution is 57%.
The composition differences between the dilute sorbent solution and the concentrated solutions
variations reflect the chiller performance. The wider the discrepancies between the dilute and
concentrated sorbent solution are, the higher the COP values will be.
4.2.6 Vapor Quality Variations
The vapor qualities of the chiller internal conditions at different state points are assumed in the model.
The refrigerant vapor from the HTRG at state point 23 was initially assumed to be completely
condensed in the LTRG, so only saturated water enters the condenser at state point 24, q 0 . This
assumption results in generating high COP values because the LTRG recovers most of the latent heat
by the condensation process. This assumption, however, produces higher overall and weighted
deviations for all 9 data sets. The values of vapor quality q24 have been adjusted as shown in Figure 4-
13. The overall and weighted deviations are reduced dramatically by using these new q24 values. The
chiller performance can be improved by an appropriate measure reducing the vapor carryover.
Figure 4-13: Refrigerant vapor quality leaving the LTRG under various load conditions
Vapor quality (q 24) .
0.4
0.35
0.3
0.25
0.2
0.15
0.1
0.05
0
Vapor quality of
refrigerant entering
Condenser
0 2 4 6 8 10 12 14 16 18 20
Vapor quality of refrigerant entering condenser
Actual cooling load (kW)
24 =
83

4.2.7 Heat Transfer Area Variations
The UA values of the 5 major heat transfer components are plotted in Figure 4-14. The overall heat
transfer coefficients (Us) of the 5 major heat transfer components are functions of the mass flow rates,
inlet and outlet temperature of streams on both sides of the tubes; the effects of the stream flow rates
and temperature are presented in appendix 4A. The model initially assumes that the contact area (As)
of each heat transfer component remains constant. The model analysis, however, shows that the
decrease in surface contact areas for the heat transfer components in the evaporator, and the LTRG
under partial load conditions may contribute to the decrease of UA values. The variations of contact
areas in the evaporator and the LTRG are due to the significant flow rate changes. Figure 4-15 shows
the estimations of the area changes in the evaporator and the LTRG on the basis of the overall
deviations between the measured values and the model solutions.
Figure 4-14: UA changed for the 5 major components under various load conditions
UA, kW/ o C
4
3.5
3
2.5
2
1.5
1
0.5
0
UA for Condenser
UA for Evaporator
UA for Absorber
UA for LTRG
UA for HTRG
0 2 4 6 8 10 12 14 16 18 20
Absorber Evaporator HTRG LTRG Condenser Actual cooling load (kW)
The surface area for the evaporator is indicated in Figure 4-15, The surface contact areas decrease at
partial load conditions by 30-50%. The reason for this change is the significant refrigerant flow
decrease. For instance, Figure 4-10 indicates that this drops from 0.0075 kg/s at design load condition
to 0.0016 kg/s at 34% of design load condition.
The surface variations also exist in the LTRG because of the flow decrease from the design load
condition to the partial load condition. Figure 4-8 indicates that the sorbent solution distributed to the
84

LTRG drops from 0.048 kg/s at design load condition to 0.022 kg/s at 34% of design load condition.
The chiller controls the solution levels in the HTRG but not in the LTRG. The total contact area of the
LTRG is estimated to decrease by 20-30% from the design load to 34% of design load conditions. This
means that under the lower load conditions, some of the tubes in the LTRG may be exposed to the
refrigerant vapor.
The surface contact areas of the HTRG, condenser, and absorber are not significantly affected by the
load variations. First, the control system maintains the solution level in the HTRG. Figure 4-8
indicates that the solution sorbent solution flow rate is relatively constant for all load conditions;
second, the absorber contact area does not vary much from the design load condition to the partial load
condition because, as Figure 4-8 indicates, the dilute sorbent solution circulation rate does not change
significantly for all load conditions; third, the condenser contact area does not change much because
the tubes are consistently exposed in the refrigerant vapor.
Figure 4-15: Surface contact area changes under various load conditions
Surface area, m 2
1.6
1.4
1.2
1
0.8
0.6
0.4
0.2
4.2.8 Deviation Variations
0
Contact area of
LTRG
Contact area of
Evaporator
0 2 4 6 8 10 12 14 16 18 20
Evaporator surface area LTRG surface area
Sorbent solution flow (kg/s)
The overall deviations and the weighted deviations of the measurements and the model solutions are
plotted in Figure 4-16. The overall deviations are below 6% when the cooling loads are below 60% of
the design condition. When the load drops below 60%, the overall deviations increase fast to 13% at
34% of design load condition. The dramatic increase of overall deviation is due to the inaccuracy of
85

steam flow measurements and the relative increasing discrepancy of condensate return temperature
between the measured values and the model solutions.
Figure 4-16: Overall and weighted deviations under various load conditions
Overall and weighted deviations .
16%
14%
12%
10%
8%
6%
4%
2%
0%
Weighted deviation
Overall deviation
0 2 4 6 8 10 12 14 16 18 20
Overall deviationr Weighted deviation
Actual cooling load (kW)
4.2.9 Analysis of Other Test Data
Only the results of the cooling-load variation tests are presented in this chapter. The analyses of other
test data are implemented in the following order:
• chilled-water supply temperature variation
• chilled-water flow rate variation
• cooling-water supply temperature
• cooling-water supply flow rate
• steam supply temperature
The results of the analysis are presented in the appendix 4A.
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5 Contributions and Areas of Future Research
5.1 Contributions
This thesis develops methods for the effective design and evaluation of an absorption chiller for
micro-BCHP systems that reduce energy consumption, decrease operational costs, and improve
environmental benefits in residential and light commercial buildings.
1) Establishment of a unique experimental environment for the equipment tests under various
conditions.
A 16 kW steam-driven, two-stage absorption chiller was installed together with an auxiliary steam
supply and a variable load for the chiller test and performance evaluation. We developed a web-based
data acquisition and control system to operate the chiller and its auxiliary equipment while storing and
displaying the test measurement data. We tested the chiller at various operating conditions in
accordance with a test program. On the basis of the test program, the effects of chilled water, cooling
water, and steam input operating conditions on the chiller performance were examined systematically.
The chiller performance was calculated and presented on the basis of the measurement data gathered
in the test program. The calculated chiller performance data under various load conditions checked and
supplemented the performance data from manufacturer publications. In the future, the chiller and its
control system will be incorporated in the cooling system of the IW and connected with other BCHP
components and the campus chilled-water supply system.
2) Construction of a comprehensive chiller model for the analysis of extensive, detailed test data
obtained from the absorption chiller
A comprehensive computational model was developed to further refine the understanding of the
principles of the chiller, to analyze the experimental data from the test program, to assist in equipment
design, and to evaluate the performance of various BCHP systems. This model is a set of equations
consisting of: mass balances, energy balances, relations describing heat and mass transfer, and
equations for the thermophysical properties of the working fluids. The model can be solved when
appropriate assumptions and a certain number of operating parameters are assigned, so that the
conditions – pressure, temperature, composition, and flow – at each point within the chiller can be
calculated. Heat and mass transfer correlations have been integrated into the model so that it can
evaluate the chiller performance not only at design conditions, but also at various off-design
conditions.
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3) Analysis of the measured data, refinement of the model, and improvement of the chiller design on
the basis of the data analysis process
The comprehensive model has been used to assess the accuracy of the experimental data from the test
program. The discrepancies between the measurements and the model calculations were reduced by
adjusting the model assumptions. The discrepancies between the measurements and the model
solutions are introduced mainly by the following factors:
• inaccurate stream flow temperature measurements from sensors mounted on the external pipe
surface
• fluctuating measurements of steam flow due to periodic feedwater addition to the boiler
• imprecise cooling-water flow measurements because of space limitations in mounting the flow
sensor
• inaccurate assumptions regarding the quality of the refrigerant flow from various chiller
components
• inaccurate values of heat transfer coefficients calculated from available correlations
The absorption cycle for each test has been plotted on a Dűhring diagram based on the model
calculations. The trends of temperature, pressure, and composition of critical state points for each
group of operating parameter tests have been summarized on the Dűhring diagram and other plots. We
have devised a strategy to improve the chiller performance (particularly at partial load conditions) on
the basis of model analysis results and have validated the model for the calculation of chiller
performance under various operating conditions.
These research efforts have provided a solid basis for future studies on microscale absorption chiller
design, application, and simulation. Current work can be extended into the following research areas in
the future:
• Extension of the validated model to various heat sources and sinks and thermal capacities in
microscale BCHP system design evaluations
o The model can be extended to heat sources including natural gas, hot water, and exhaust gases
from engines and gas turbines
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o The model can also be adapted to air cooling, but this adaptation may reduce the capital and
maintenance costs, although the rated COP may drop when compared with the water-cooled
system.
• Integration of the chiller performance and cost models into overall simulations of microscale
BCHP systems to optimize overall system performance and operating strategies
o A cost model including capital cost, operational cost, and maintenance cost can be integrated
with building load simulation tools to evaluate absorption chiller economic performance under
various types of buildings and weather conditions.
o A guideline for applying the microscale absorption chiller in buildings can be proposed on the
basis of the simulation results of the economical evaluation model.
o As a simulation tool, the model should provide a graphic user interface (GUI) and standard
output sheets that can assist the system designers in implanting BCHP system design and
evaluation.
The field of computational support for building an energy system is extensive, and this thesis has
illustrated significant concepts in designing, analyzing, and modeling of microscale absorption chiller
systems and of analyzing extensive test data sets with the support of a detailed model. Some of the
future areas of study have been investigated preliminarily in this thesis along with the chiller
equipment tests and the experimental data analysis processes. The methods and some of the results are
summarized in the following sections.
5.2 Areas of Future Research
5.2.1 Extended Chiller Model for Multi-Heat Resources
Many types of fuel and thermal heat sources can be used to drive a double-effect absorption chiller,
such as steam, hot water, exhaust gas, natural gas, oil, and LPG. Among them the most widely used
heating media are natural gas, steam, and hot water. From the manufacturer’s perspective, the chiller
can be adapted to any heat source with minor changes on the HTRG and internal control system.
From a research perspective, the validated chiller model, based on a modularized structure, allows for
the flexible extension of one specific chiller to other types of chillers with similar flow configurations
but different heating media. The extended absorption chiller models using various heat sources can
meet the total cooling demand of buildings and can better integrate with other BCHP components,
such as solar collectors and various power generators.
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Compared with the steam-driven chiller model, the major difference in the new types of chiller is in
the calculation of heat transfer coefficients for the HTRG. In the steam-driven chiller, the condensate
after the HTRG is recovered by the HRHX. In the other types of chiller, however, this HRHX does not
exist. The heat transfer features of the HTRG for the four types of chiller are listed in Table 5-1. The
common feature of the four HTRGs is the boiling process of LiBr solution. The detailed
configurations of the HTRGs are summarized in Tables 3-2 and 3A-1. The heat transfer coefficient
equations applied for the HTRGs are listed in Table 3-3.
Table 5-1: Heat transfer features of the HTRG of different heating media
Outside tube (combustion chamber)
Heating medium Reference Process Medium
All types Jakob and Hawkins Boiling Water-LiBr sorbent solution
Inside tube (combustion chamber)
Steam Kern D. Q. Condensation Steam, condensate
Hot water Dittus and Boelter Convection Water
Exhaust gas J. P. Holman Convection Combustion gases
Natural gas Hottel, et al. Radiation, convection Combustion gases
5.2.1.1 Hot Water Absorption Chiller
The HTRG using hot water comprises a spiral circular tube bundle with 3 parallel tubes spiraling 8
rounds down to the bottom; hot water is split into three streams at the inlet located at the top and
combined into one stream on the bottom of the HTRG. The hot water flow is regulated by a motorized
hot water valve. The hot water supply temperature is 160 o C, and the temperature difference between
the hot water inlet and the outlet is 10-20 o C. When the hot water leaves the HTRG, it can be reheated
by external heat sources such as solar collectors or other heat recovery systems.
5.2.1.2 Natural Gas Absorption Chiller
In a natural-gas-driven absorption chiller, the HTRG in Figure 5-1 comprises a combustion chamber
(burner) cooled by a radiation convection section and an exhaust gas. The combustion gases exit at the
far end of the combustion chamber at a temperature significantly lower than the adiabatic temperature.
The combustion gases are cooled further in the exhaust gas convector, so that the exhaust gas
temperature approaches the sorbent solution temperature in the HTRG within 30 o C. The natural gas
flow to the burner is regulated by a flow switch. The control logic of natural gas flow is similar to that
of steam, but the natural gas burner has only two stages, high flame and low flame.
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Figure 5-1: Simplified HTRG configurations for natural-gas-driven absorption chiller
Combustion
Chamber
Fuel ejector
200
400
200 200
Exhaust gas
25 Exit
5.2.1.3 Exhaust Gas Absorption Chiller
Exhaust gas
Convector
Using exhaust gas directly from a power generator, such as an engine, gas turbine, or solid oxide fuel
cell, is one of the latest practices in the application of absorption chillers. The structure of the HTRG
for an exhaust-gas-driven absorption chiller is similar to that of the gas convector in the natural-gasdriven
chiller. It is comprises a staggered tube bundle with 22 circular grooved copper tubes; the
exhaust gas is split into 22 streams at the inlet located at one side and combined into one stream on the
other side of the HTRG. High-temperature exhaust gas can be supplied at 520 o C from an engine or a
reciprocated gas turbine, 755 o C directly from a solid oxide fuel cell. When the exhaust gas leaves the
tube bundles, its temperature is usually higher than the solution temperature by 30 to 50 o C.
5.2.2 System Integration and Application
Many cogeneration concepts are conceivable with absorption chiller systems, but the selection of one
over another requires detailed study of long-term technical and economic performance. On the basis of
the chiller models developed in this thesis, the design and analysis of an individual absorption chiller
can be expanded to overall BCHP systems. An integrated design, control, and operation strategy can
be developed to maximize the overall efficiency while lowering the capital cost and later the
associated operation and maintenance fees. An annual simulation using TRNSYS tools with refined
building information can be conducted for several simplified system configurations associated with the
four heat sources for an absorption chiller and a cost model to comprehensively analyze the effects of
building occupancy and weather variations on system overall efficiencies and economic benefits.
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5.2.2.1 Chiller Performance Tables for Building Simulation Tools
TRNSYS building simulation tools can, in principle, directly integrate the inputs and outputs of
detailed EES models for system and plant equipment, but this method is not very convenient because
the iteration of building load calculations requires an iterated solution of the mathematical calculation.
Computation times are increased greatly by using this method, particularly when the equipment model
becomes more and more complex. As an alternative, chiller performance tables based on input
operating conditions and output performance from solutions of a computational model are more
appropriate for use in overall system simulation. The performance tables generated from this thesis can
greatly enrich the limited component library of TRNSYS simulation tools.
5.2.2.2 Cost Model
The purpose of a comprehensive cost model is to support the decision-making process in designing an
absorption chiller-based BCHP system. As the basis of economical analysis, a cost model can be
developed to forecast the cost of chilled water of different system configurations. The model will
make use of capital costs, operation costs (grid electricity and steam, natural gas, hot water, and
exhaust gas), interest rates, expected return-on-investment, system efficiency, and maintenance cost to
predict the system economical performance.
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Appendix 1A
97

Appendix 2A
102

Appendix 2B
118

Appendix 3A
130

Appendix 4A
150

Acronyms
ACEEE: American Council for Energy Efficient Economy
ALC: Automated Logic Co.
BAPP: building as power plant
BAS: building automation system
BCHP: building cooling heating and power
CHP: combined heating and power
CHW: chilled water
CHWS: chilled-water supply
CHWR: chilled-water return
COCHW: cost of chilled water
COP: coefficient of performance
CMU: Carnegie Mellon University
CW: cooling water
CWS: cooling-water supply
CWR: cooling-water return
EES: engineering equation solver
EIA: energy information administration.
ESS: energy supply system
GUI: graphic user interface
HVAC&R: heating ventilating air conditioning and refrigerating
HTRG: high-temperature regenerator
HTHX: high-temperature heat exchanger
HRHX: heat recovery heat exchanger
IW: the intelligent workplace
ICPC: integrated compound parabolic collectors
LiBr: lithium bromide
LPG: liquid pressurized gas
LTRG: low-temperature regenerator
LTHX: low-temperature heat exchanger
SOFC: solid oxide fuel cell
TRNSYS: transient systems simulation program
WCS: web control server
194