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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.

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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

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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.

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

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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

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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

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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

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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

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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.

1

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


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

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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 energyconsumption(99.74 Quads)

Electricity(14.2 Quads)

Power generation(40.77 Quads)

Transportation(27.79 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 productionwaste heat(24.5 -26.5 Quads)

Residential sector waste heat(1.38-2.08 Quads)

Commercial sectorwaste heat(0.8-1.2 Quads)

Industrial sectorwaste heat(2.12-4.24 Quads)

National totalwaste heat(28.8-34.02 Quads)

Industrial sector(21.18 Quads)

Residential sector(6.92 Quads)

Commercial sector(4.02 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.

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Figure 1-2: Conceptual Diagram for System Integration in Buildings

System Integration

Resource conservation:Energy, water, material, and so forth

Distributed generation: engine generator, gas turbine, and fuel cell

Renewables: solar, wind, bio-gas, day-lighting, natural ventilation, passive/active heating/cooling

Source: Volker Hartkopf [5]

1.1.1 CHP Systems

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 large-

scale 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

4

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

5

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 oC; integrated compound parabolic collectors, ICPC’s, 140 to 160 oC; flat plate collectors, 65 to

90 oC.

Table 1-1: Power generation equipment rejected heat temperature ranges

No. Power Generation Equipment, Waste Stream Temperature (oF) Temperature (oC) 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 oC. 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 (oC) 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)

6

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.

7

Figure 1-4: Basic vapor compression chiller cycle

Condenser

Evaporator

from chilled water

expansionRefrigerant

valve

Heat rejected

Work

Compressor

Heat absorbed

T

P

to cooling water

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


valve

expansionSolution

valve

from chilled waterHeat absorbed

pumpSolution

T

P

Condenser

Heat rejectedto cooling water

Evaporator Absorber


Regenerator

Heat input

8

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 oC 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.

9

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.

10

• 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

Low-temp.regenerator

High-temp.heat exchanger

Low-temp.heat exchanger

Refrigerantcombiner

pumpRecirculation

Solutioncombiner

splitterSolution

T

P


valve

expansionSolution

valve

pumpSolution

Condenser

LTRGHeat to


Evaporator Absorber


Condenser


Regenerator

Heat input

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, high-

pressure regenerator; and the partially concentrated sorbent solution then flows to the lower-pressure,

lower-temperature regenerator, as shown in Figure 1-7.

11

Figure 1-7: Typical two-stage series flow absorption chiller configuration

Low-temp.regenerator

High-temp.heat exchanger

Low-temp.heat exchanger

Refrigerantcombiner

pumpRecirculation

T

P


valve

expansionSolution

valve

PumpSolution

Condenser

LTRGHeat to


Evaporator Absorber


Condenser


Regenerator

Heat input

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:

12

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 ft2 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

13

• 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

14

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

15

results 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

16

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

17

• 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

18

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

19

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.

20

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

21

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: chilled-

water 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.

22

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-sorbent-

flow 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: Absorption chiller installed in the IW 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.

23

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

24

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

25

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

Name Quantity Unit Cooling capacity 16 kW Chilled-water return temperature 14 oC Chilled-water supply temperature 7 oC Chilled-water flow rate 2 m3/h

Chi

lled

wat

er

Chilled-water pump head 8 mH2O Rated steam pressure, absolute 0.7 mPa Steam pressure limit, absolute 0.9 mPa

Stea

m

Maximum steam consumption 24 kg/h Power voltage 220 V Power frequency 60 Hz

Pow

er

Maximum power consumption 1 kW Water-LiBr sorbent solution mass 65 Kg Water-LiBr sorbent concentration 55 %

Solu

tion

26

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 oC. The vaporization absorbs heat

from the chilled water flowing through the evaporator coil, cooling this flow from 14 oC to 7 oC.

Figure 2-4: Configuration of the lower vessel

27

At a constant flow rate of 2 m3/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 m3/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 oC. 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.

28

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.

Figure 2-5: Configuration of the upper vessel 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.

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.

29

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 oC.

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 oC. 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

30

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.

31

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 oC. The speed of the cooling tower air fan is varied to

maintain the cooling-water supply to the chiller at 30 oC.

Figure 2-6: Configuration of cooling tower

Air inlet

City water

PVC Plates

PVC mist

Spray

CTF

Water tank

CWBPV

To chiller

Air outlet

From chiller

CWPDrain

CWDV

CTWS

CWDD

collector

nozzles

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 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 oC. The speed of the cooling-tower air fan is varied to maintain the

cooling-water supply to the chiller at 30 oC.

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

32

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.

33

Figure 2-7: Simplified flow diagram of the chiller test system

IV5

IV4

IV3P2T20F1

BFP

P7 T25

IV2P3 T21

WS BFT

CTW

ESB

SS

TLHXAbsorption

chiller

F2T22P4

HW

R

BH

WSV

HW

S

ALC

Additional

P5T23

CR

sensors tochiller

Variable cooling loadSteam supply system

Cooling loads Steam supplyAbsorption chiller

2.2.1.1 Steam Supply System

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

34

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

35

80 oC 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 m3/h and a temperature of 14 oC. The chiller cools the chilled water to

7 oC. 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.

36

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

37

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

38

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

39

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

T3

Drain

Steam

CTWS

B1 T1CHWR

City water

RP

CHWP

B2

T6

L2

SP

T12

T9 T2CHWS

Con

trolle

rT10

Condensate T8

T13

T11

ME-

LGR

25

T7 P1

T14

SV

L3

CWP

L4

Air

L5

T5

T18T17

T15

T16

L1F6

Hot Humid Air

T19

T32

RPH

RBPSV

CWBPV

CWF

CWDD

DV

T33

To a

uxila

ry sy

stem

HTRG

Condenser

LTRG

AbsorberEvaporator

HTHX LTHX

HRHX

CTF

CHSV

SF

40

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.

41

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

42

Table 2-4: Instrumentation of the chiller test systems

Label Sensor location Medium Range Manufacturer Accuracy

On-site Calibration

T1 Chilled-water return Water (-15 oC) to 110 oC ± 0.1% T2 Chilled-water supply Water (-15 oC) to 110 oC ± 0.1% T3 Cooling-water supply Water (-15 oC) to 110 oC ± 0.1% T5 High-temperature regenerator Solution (-15 oC) to 210 oC ± 0.1% T6 Ambient Air (-15 oC) to 110 oC ± 0.1% T7 Steam supply Steam (-15 oC) to 210 oC ± 0.1% T8 Condensate return Water (-15 oC) to 210 oC ± 0.1%

Tem

p.

T9 Chilled-water return 2 Water (-15 oC) to 110 oC ± 0.1% L1 HTRG solution level probe Solution 4 pins L3 LTRG upper-limit level probe Solution 1 pin L4 Auto vacuum device level

probe Solution 1 pin

Lev

el

L5 Cooling-water level probe Water 1 pin B1 Chilled-water flow detector Water on/off

Flow

B2 Chilled-water flow detector Water on/off D1 Solution pump frequency,

amps, and voltage Electricity

Abs

orpt

ion

Chi

ller

(man

ufac

ture

r)

Pow

er

T11 Condensate after HTRG Surface 0-400 oF ± 0.1% 0.5 oF T12 Solution in Absorber Surface 0-400 oF ± 0.1% 0.5 oF T13 Solution entering HRHX Surface 0-400 oF ± 0.1% 2 oF T14 Solution leaving HRHX Surface 0-400 oF ± 0.1% 1.5 oF T15 Cooling water after absorber Surface 0-400 oF ± 0.1% 1.5 oF T16 Cooling water after condenser Surface 0-400 oF ± 0.1% 1.0 oF T17 Low-temperature regenerator

(LTRG) Surface 0-400 oF ± 0.1% 0.3 oF

T18 Refrigerant after condenser Surface 0-400 oF ± 0.1% 1.5 oF T19 Refrigerant from evaporator Surface 0-400 oF ± 0.1% 2.0 oF T32 Cooling water after cooling

tower Surface 0-400 oF ± 0.1% 0.0 oF

Tem

p.

T33 HTRG temperature Surface 0-400 oF ± 0.1% 1.5 oF F6 Cooling-water flow Water 0 to 30 gpm ± 1% -15%

Abs

orpt

ion

Chi

ller

(AL

C)

Mis

c.

E1 Electric power of absorption chiller

Electricity 0-2400 amps ± 1% 0.0%

T22 Steam-supply temperature Steam (50 oF) to 250 oF ± 0.1% 0.0 oF T23 Condensate-return

temperature Water (50 oF) to 250 oF ± 0.1% 0.0 oF

Tem

p.

T25 Feed-water temperature Water (50 oF) to 250 oF ± 0.1% 0.0 oF P4 Steam supply pressure Steam 0 to 40 gpm ± 0.13% P5 Condensate return pressure Water 0 to 100 psi ± 0.13%

Pres

.

P7 Feed-water pressure Water 0 to 50 psi ± 0.13% F2 Steam flow Steam 0 to 75 lb/h ± 0.5% 0-10%

Stea

m S

yste

m

Mis

c.

E2 Electric power of steam boiler Electricity 0-2400 amps ± 1% 0.0% T20 Chilled-water supply Water (-10 oF) to 110 oF ± 0.1% 0.6 oF T21 Chilled-water return Water (-10 oF) to 110 oF ± 0.1% -0.4 oF

Tem

p.

T31 Ambient temperature Air (-58 oF) to122 oF ± 0.1% P2 Chilled-water inlet Water 0 to 100 psi ± 0.13%

Pres

.

P3 Chilled-water outlet Water 0 to 50 psi ± 0.13% F1 Chilled water Water 0 to 20 gpm ± 1% 0%

Coo

ling

Loa

d Sy

stem

Flow

43

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

44

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 oC throughout the test program. While this setting remained constant, the measured value of chilled-

water outlet temperature varied ±2 oC 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.

45

Table 2-5: Input and primary output of the test program

Inputs Primary Outputs Calculated performance

CHW

return T CHW flow

CW supply T

CW flow

Steam pressure

Steam Flow

CHW supply COP

Cooling load

oC kg/s oC kg/s kPa kg/s oC kW Design

condition 13.9 0.5616 30.78 1.546 600 0.00727 7 1.03 16.63 CHW

return T 8–14 Design Design Design Design 0.00382–0.00771 6.25–6.9 0.95–1.03 9–18

CHW flow Design

0.531–0.864 Design Design Design

0.00705–0.00771

6.42–8.97 0.93–1.02 16.56–17.96

CW supply T Design Design 27.5–36 Design Design

0.00609–0.00764

6.13–9.26 1.11–0.91 19.22–13.29

CW flow Design Design Design 0.784–1.547 Design

0.00618–0.00709 7.52–9 0.81–0.98 11.73–15.23

Steam pressure Design Design Design Design 360–700

0.00556–0.00748

6.34–10.27 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

46

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

( )CHWSCHWRpCHWcooling TTCmQ −××= & (Equation 2-1)

Where CHWm& is the flow rate of the chilled water

CHWRT is the temperature of the chilled water entering the chiller

CHWST is the temperature of the chilled water leaving the chiller

pC is the specific heat of water.

The COP of the chiller is conventionally

heat

coolingthermal Q

QCOP = (Equation 2-2)

Where, ( )condensatesteamsteamheat hhmQ −×= & = the heat delivered to the chiller by steam

condensation; this quantity can also be estimated from the electrical power

consumption in the steam boiler.

steamm& is the flow rate of the steam supply

steamh and condensateh 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

energy

coolingoverall Q

QCOP = (Equation 2.3)

powerheatenergy EQQ += (Equation 2-3)

powerpowerpower IVE *=

47

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 ( thermalCOP ), 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

48

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

0

5

10

15

20

25

30

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

Time

Stea

m su

pply

flow

rate

(kg/

h),

Coo

ling

load

(kW

)

0

0.5

1

1.5

2

2.5

3

3.5

4

Cof

ficie

nt o

f Per

form

ance

(CO

P)

Steam supply flow rate Cooling load COP

COPthermal

Steam flow rate (F1)

Cooling load (Qcooling)

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.

49

The cooling water fluctuates some because of the operation of the cooling-tower fan. The cooling-

water supply temperature is consistently maintained at 30 oC within 1± oC 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

0

20

40

60

80

100

120

140

160

180

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

Time

Tem

pera

ture

(o C)

CHWR CHWS SSCWS CW after ABS CW after CONDLTRG HTRG Refrigerant after COND

Steam supply

Chilled water supply

Cooling water supply and return

HTRG sorbent

Chilled water return

LTRG sorbent

Refrigerant after Condenser

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 oC, respectively.

50

Table 2-6: Measurement data of the chiller under design condition

Label Sensor location Medium Unit Before calib. After calib. T1 Chilled-water return Water oC - 13.83

T2 Chilled-water supply Water oC - 6.65

T3 Cooling-water supply Water oC - 30.01 T5 High-temperature regenerator Solution oC - 154.87 T6 Ambient Air oC - 24.26 T7 Steam supply Steam oC - 158.56 T8 Condensate return Water oC - 99.00

Tem

p.

T9 Chilled-water return 2 Water oC - 6.58

L1 HTRG solution level probe Solution L3 LTRG upper-limit level probe Solution L4 Auto vacuum device level probe Solution L

evel

L5 Cooling-water level probe Water B1 Chilled-water flow detector Water

Flow

B2 Chilled-water flow detector Water E1 Solution pump frequency, amps, and voltage Electricity

Abs

orpt

ion

chill

er (m

anuf

actu

rer)

Ele

.

T11 Condensate after HTRG Surface oC 151.23 157.65 T12 Solution in absorber Surface oC 36.7 36.63 T13 Solution entering HRHX Surface oC 74.99 75.02 T14 Solution leaving HRHX Surface oC 90.54 90.37 T15 Cooling water after absorber Surface oC 37.46 37.7 T16 Cooling water after condenser Surface oC 40.06 40 T17 Low-temperature regenerator (LTRG) Surface oC 92.71 93.21 T18 Refrigerant after condenser Surface oC 43.29 44.08 T19 Refrigerant from evaporator Surface oC 127.78 129.24 T32 Cooling water after cooling tower Surface oC 30.78 30.74

Tem

p.

T33 HTRG temperature Surface oC 150.54 153.44 F6 Cooling-water flow Water kg/s 1.55 1.45

Abs

orpt

ion

chill

er (A

LC

)

Flow

T22 Steam-supply temperature Steam oC 163.56 164.31

T23 Condensate-return temperature Water oC 99.43 99.3

Tem

p.

T25 Feedwater temperature Water oC 70.09 59.1 P4 Steam supply pressure Steam psi 698.04 713.87 P5 Condensate return pressure Water psi 101.75 100.04

Pres

.

P7 Feedwater pressure Water psi F2 Steam flow Steam kg/s 0.00727 0.00734

Stea

m sy

stem

Flow

T20 Chilled-water supply Water oC 13.9 13.88

T21 Chilled-water return Water oC 6.83 6.39

Tem

p.

T31 Ambient temperature Air oC 28.49 23.2 P2 Chilled-water inlet Water psi 298.58 154.23

Pres

.

P3 Chilled-water outlet Water psi 372.32 233.33 F1 Chilled water Water kg/s 0.56 0.56

Coo

ling

load

syst

em

Flow

51

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

Name Unit Specification Before calibration After calibration Cooling capacity kW 16 16.62 17.62 Chilled-water return temperature oC 14 13.9 13.88 Chilled-water supply temperature oC 7 6.83 6.39 Chilled-water flow rate m3/h 2 2.02 2.02

Chi

lled

wat

er

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

Stea

m

Power voltage V 220 220 220 Power frequency Hz 60 60 60

Pow

er

Power consumption* kW 1 0.856 0.823 COP (Thermal) 0.98 1.038 COP (overall) 0.93 0.99

Perf

orm

.

* 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 oC to 8 oC 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

52

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 natural-

gas-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

Measurement values for chiller inputs Measurement values for chiller outputs

Chilled- water flow

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 Qcooling COP

No. m3/h kg/s oC oC oC oC kg/h oC 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

53

Figure 2-16: Chiller performance under various load conditions

0.4

0.5

0.6

0.7

0.8

0.9

1.0

1.1

1.2

1.3

0 2 4 6 8 10 12 14 16 18 20

Actual cooling load (kW)

Coe

ffic

ient

of P

erfo

rman

ce (C

OP)

.

Thermal COP (measurement) Overall COP (measurement)

Thermal COP(measurement)

Overall COP(measurement)

Figure 2-17: Chiller power consumption under various load conditions

0.0

0.4

0.8

1.2

1.6

2.0

0% 10% 20% 30% 40% 50% 60% 70% 80% 90% 100% 110%

Actual load / design load

Chi

ller

pow

er c

onsu

mpt

ion

(kW

) .

Steam chiller Natural gas chillerNatural gas & hot water chiller

Measured power consumption for natural gas and hot water

Rated power consumption for natural gas driven chiller

Measured power consumption for steam driven chiller

54

Figure 2-18: Comparison of chiller performance

0

4

8

12

16

20

0 4 8 12 16 20Actual cooling load (kW)

Act

ual h

eat i

nput

(kW

) .

Steam-driven chiller Natural gas-driven chiller

Steam-driven chiller

Natural-gas-driven chiller

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.

55

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.

56

Figure 3-1: Simplified flow diagram for chiller model

1

2

3

19

4

5

6

7

8

91

10

11

14

18

20 21

22

23

24

5152

42 43

32 31 4142

53

16

17

12

13

25

Heat to LTRG

LTRG (outside)

High-temp. heat exchanger

LTRG

Condenser

HTRG

LTHX

Absorber

Evaporator

HRHX

Heat to

BPHX

Refrigerantcombiner

spray nozzleRefrigerant

RPspray nozzleSolution

SP

SolutioncombinerSplitter

Solution

5254

9

92

46

45

44 41

43

Air

from

Cooling tower

Chilled water Cooling water

SteamTo HRHX

to

Condenser

Absorber

CHWP

CWP

(inside)

water

Steamtrap

Chiller

Sorbent solutionRefrigerantState points: pressure, temperature,

Equilibrium states: thermal and vapor

Absorber

47City water

HTHX

High temp. regenerator

Low-temp. regenerator

Low-temp. heat exchanger

composition, flow of streamsentering/leaving chiller components

liquid equilibrium between streamsleaving a chiller component Cooling

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.

57

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 (oC). The inclined line on the left of the plot represents vapor pressure-

temperature 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

58

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

57.39

%

T1 T4 T5 T6T7 T21

T11

T41T42

T43

T32

T31

62.81

%

61.68 %

T91

57.39

%

T22T20

T51

REFRIGERANT TEMPERATURE, C

SATU

ATI

ON

PR

ESSU

RE

(P),

kPa

SOLUTION TEMPERATURE, CEqulibrium Chart for A queous Lithium Bromide Solutions

10 20 30 40 50 60 70 80 90 100 110 120 130 140 150 160 170 180

10

20

30

40

50

60

70

80

90

100

110

120

030%

40%

50%

60%

70%

1

2

3

45

10

20

30

40

50

100

150

200

05

o

o

T52

Evaporator

Pinch point

Pinch point

Absorber

LTRG

Pinch point

Pinch point

HTRG

Condenser

Pinch point

Pinch point for LTHXPinch point for HTHX

for absorber

evaporator

Pm

Ph

Pl1

4

56

20

21

227

91

T54

T53

Condenserfor LTRG

for HTRG

HTHX

LTHX

T18

HRHX

RefrigerantWater-LiBr sorbent solutionSteam, cooling water, and chilled water

59

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

00

20

0 5 10 15 20 25 30

40

60

80

100

120

140

160

3231

414243

18

1

911

5

5123

24

5221

20

6

22

1

20

7

1

4

5

53

3514

40 45 50 kW

EvaporatorAbsorberCondenserLTRGHTRG

LTHX HRHX

HTHXCo

55 60 65 70 75 80 85

3.4 Calculation Procedure

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

60

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:

∑∑ = outin mm &&. (Equation 3 - 1)

where, m& = mass flow rate.

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

outoutSolution

inin xmxm && (Equation 3 - 2)

where, x = the weight concentration of the water-LiBr sorbent solution.

3.4.2 Energy Balance

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:

0=+− ∑ jjshaft hmWQ &&& (Equation 3 - 3)

where Q& is the quantity of heat transfer to or from the system; shaftW& 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.

61

For instance, the cycle involving with absorb process, the energy balance using equation 3-3 can be

written in the following equation:

dilutediluteconcconcwaterwater hmhmhmQ &&& =++ (Equation 3 - 4)

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:

( )jjj PPTTsteamwaterenthalpyh === ,,/ (Equation 3 - 5)

Phase equilibrium for saturated liquid and vapor in equilibrium:

( )jj PPsteamwaterSATTT == ,/_ (Equation 3 - 6)

( )jj TTsteamwaterSATPP == ,/_ (Equation 3 - 7)

( )jjj PPhhsteamwaterqualityq === ,,/ (Equation 3 - 8)

For the cooling-tower model, moist air properties can be evaluated by the following relations in EES:

)RHR ,TT,PP,irEnthalpy(Ah jjjH2Oj ==== (Equation 3 - 9)

)RHR,PP,TT,rWetbulb(AiT jjjH2Owb-j ==== (Equation 3 - 10)

)ww,PP,TT,Relhum(Air jjjH2Oj ====RH (Equation 3 - 11)

3.4.4 Heat Transfer Models

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.

62

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.

LMTDTQUA

∆= (Equation 3 - 12)

In equation (3-12), Q is the quantity of heat transferred in the components. LMTDT∆ is the log-mean

temperature difference; it is expressed below as:

( ) ( )

incoldouthot

outcoldinhot

incoldouthotoutcoldinhotLMTD

TTTT

TTTTT

,,

,,

,,,,

ln−−

−−−=∆ (Equation 3 - 13)

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, actualQ , to the maximum potential heat transfer, maxQ , below:

( )( )

( )( )incoldinhot

outcoldinhot

incoldinhot

incoldoutcoldactual

TTTT

orTTTT

QQ

,,

,,

,,

,,

max −

−

−

−==ε (Equation 3 - 14)

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

63

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 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

Steam

HTRG Copper Spiral tube Grooved LiBr Boiling Steam Condensation

Hot water HTRG Copper Straight tube Grooved LiBr Boiling Hot water Convection

HTRG Steel Comb. chamber Grooved LiBr Boiling Combustion gases Radiation Natural gas

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, ih and oh :

( )outout

inout

inin

overall

AhkLdd

Ah

UA 12

ln11

++=

π

(Equation 3 - 15)

The subscripts in and out refer to the inside and outside of the tube. inA refers to the inside surface and

outA 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

inh and outh 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.

64

Table 3-3: Heat and mass transfer correlations used in the performance model

Component Reference Process Equation Comment

Chun and Seban

Water film on tube surface ( ) ( ) 65.04.03

31

3

2

PrRe108.3 filmevpfilmevpfilmevp kgvh −−

−− ×=

⋅

For vertical tube Evaporator

Dittus and Boelter

Chilled water in tube 3.08.0

30

30 PrRe023.0 chwevpchwevpchwevp

kdh

−−− =

Vliet et al. LiBr solution film on tube surface

31

291

919191

46.0 5.1Re30.0

−

−−

Γ=

gkh filmabsfilmabs ρ

µ Equation summarized for the ¾- inch (19 mm) smooth tube bank with Octhyl Alcohol surfactant

Dittus and Boelter

Cooling water in tube

4.08.0

40

40 PrRe023.0 cwabscwabscwabs

kdh

−−− =

Vliet et al. Vapor absorption rate on tube surface fg

absfilmabsvaporabs h

qm&

& 46.0Re30.0 −− = Equation summarized for the ¾- inch (19 mm) smooth tube bank with Octhyl Alcohol surfactant

Absorber

Nagaoka et al. Vapor absorption coefficient on tube surface

0.405 + 29.571 = k 91abs Γ Equation summarized for the ¾- inch (19 mm) smooth tube bank with Octhyl Alcohol surfactant

Kern D. Q. Condensation film on tube surface

3111

31

311

2

Re51.111 −− =

gkv

h filmcond

Condenser

Dittus and Boelter

Cooling water in tube 4.08.0

49

49 PrRe023.0 cwcondcwcondcwcond

kdh

−−− =

Jakob and Hawkins

Nucleate boiling on tube surface ( )

4.0

0

3121511042

−=− p

pTTh hnucleatehtrg

Steam as heating medium HTRG

Kern D. Q. Condensation film in tube 31

51

31

351

2

Re51.151 −− =

gkv

h filmhtrg

Jakob and Hawkins


4.0

0

316241042

−=− p

pTTh mnucleateltrg

LTRG

Kern D. Q. Condensation film in tube

31

24

24

31

324

2 451.124

−

−

Γ=

µgkv

h filmltrg

Cooling Tower

( ) ( )

( ) ( )( )ctct hhshhshhshhs

NTU

δδ −−−−−−−

=

45444643

45444643

4546

ln

)h-(h See paper

Jakob and Hawkins


4.0

0

3121511042

−=− p

pTTh hnucleatehtrg

HTRG

Dittus and Boelter

Hot water in tube 3.08.0

55

55 PrRe023.0 hwhtrghwhtrghwhtrg

kdh

−−− = Hot water as heating medium

Jakob and Hawkins


4.0

0

3121511042

−=− p

pTTh hnucleatehtrg

Hottel, et al. Combustion process in chamber

( ) ( ) 44wwgggg

cc

radiation TTTTA

Q σασε −= Natural gas as heating medium

HTRG

Hausen Exhaust gas in tube

( )( )[ ] 32PrRe04.01

PrRe0668.066.3d

dgashtrg

LdLd

kdh

++=− Laminar flow

Jakob and Hawkins


4.0

0

31211042

−=− p

pTTh hgnucleatehtrg

Exhaust as heating medium HTRG

J. P. Holman Exhaust gas in tube 4.08.0

55

55 PrRe023.0 exgashtrgexgashtrgexgashtrg

kdh

−−− = Turbulent flow

65

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, hP , 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, mP , 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, lP , 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

66

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

67

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.

68

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.

69

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

70

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:

1

1

2

−

∆

=∑

nXXn

n

n

σ , (Equation 4 - 1)

71

Table 4-1: Measured values and model calculations for 100% and 55% of design load conditions

Stream name Label 100% 55%

Sensor Model Unit Sensor Model

Absolute Deviation

Relative Deviation Weight Sensor Model

Absolute Deviation

Relative Deviation Weight

Condensate after HTRG * T11 T52 oC 157.5 157.5 0.0 0.0% 10% 133.7 133.7 0.0 0.0% 10%

Solution in absorber * T12 T1 oC 36.6 36.1 0.5 1.5% 8% 33.8 33.4 0.4 1.1% 8%

Solution entering HRHX * T13 T4 oC 75.1 75.1 0.0 0.0% 10% 66.1 66.1 0.0 0.0% 10%

Solution leaving HRHX * T14 T5 oC 90.5 90.5 0.0 0.0% 10% 81.3 81.4 -0.1 -0.1% 10%

Cooling-water after absorber * T15 T42 oC 37.7 36.1 1.6 4.3% 6% 33.8 33.8 0.0 0.0% 6%

Cooling-water after condenser * T16 T43 oC 40.0 38.2 1.8 4.6% 6% 34.9 34.9 0.0 0.1% 6%

Solution leaving LTRG * T17 T6 oC 93.3 92.7 0.6 0.6% 10% 79.6 79.6 0.0 0.1% 10%

Refrigerant after condenser * T18 T11 oC 44.1 44.1 0.0 0.0% 20% 37.2 37.2 0.0 0.0% 20%

Solution entering HTRG * T19 T20 oC 129.3 128.8 0.5 0.4% 10% 113.8 113.5 0.3 0.3% 10%

Chilled-water return T20 T31 oC 13.9 13.9 0.0 -0.1% 0% 10.5 10.6 0.0 -0.1% 0%

Chilled-water supply T21 T32 oC 6.4 6.4 0.0 0.8% 0% 7.1 7.1 0.0 0.5% 0%

Steam input T22 T51 oC 164.3 164.0 0.3 0.2% 0% 140.2 140.0 0.2 0.2% 0% Condensate return * T23 T54 oC 99.3 100.0 -0.7 -0.7% 2% 69.0 70.2 -1.2 -1.7% 2%

Cooling-water supply T32 T41 oC 30.7 31.5 -0.8 -2.5% 0% 30.4 31.5 -1.1 -3.6% 0%

Tem

pera

ture

Solution leaving HTRG * T33 T21 oC 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% Flow

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% Pres

sure

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

Pow

er

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%

Cal

cula

tion

Weighted deviation )'(σ 1.67% 0.48%

* The 11 measurements inside chiller to check with model calculations

72

where σ is an overall deviation; nX∆ is the deviation between the measured value of a chiller

operating condition value 'nX and the model calculated value nX :

'nnn XXX −=∆ , (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:

∑

∆=

n

n

nn X

X1

2

*' εσ (Equation 4 - 3)

The summation of weights of all the measurement conditions has been set at 1.

∑ =n

n1

1ε (Equation 4 - 4)

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.

73

Figure 4-2: Absorption cycle at design load condition

57.39

%

46.22

36.1 92.790.6575.1T1 T4 T5 T6T7 T21

T1144.11

T4131.5

T4236.11

T4338.2

6.35T32

T3113.9

62.81 %

156.2

99.97

61.68 %

45.75T91

57.39 %

48.1T22

128.8T20

T51164


SATU

ATI

ON

PR

ESSU

RE

(P),

kPa


10 20 30 40 50 60 70 80 90 100 110 120 130 140 150 160 170 180

10

20

30

40

50

60

70

80

90

100

110

120

030%

40%

50%

60%

70%

1

2

3

45

10

20

30

40

50

100

150

200

05

o

o

T52157.5

Pm, 9.155 kPa

Ph, 85.34 kPa

Pl, 0.7665 kPa1

4

56

20

21

227

91

109.6

T54

T53

3.154T18

95.26

m =0.0073 kg/s51

m =0.04437 kg/s

21

m =0.04857 kg/s

21

m =0.0042 kg/s24

m =0.04765 k

g/s

4

m =0.09622 kg/s

1 m =0.04434 kg/s

6

m =0.003315 kg/s10

m =0.007513 kg/s18

Model solutionMeasurement

+ 5 Co

RefrigerantWater-LiBr sorbent solutionSteam, cooling water, and chilled water

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 oC. 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 oC, where the model solution gives a higher value than the

measurement. The cooling-water temperatures after the absorber (T42) and the condenser (T43) show

74

deviations of about 1.6 oC and 1.8 oC, 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± oC.

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

75

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

76

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


SATU

ATI

ON

PR

ESSU

RE

(P),

kPa


10 20 30 40 50 60 70 80 90 100 110 120 130 140 150 160 170 180

10

20

30

40

50

60

70

80

90

100

110

120

030%

40%

50%

60%

70%

1

2

3

45

10

20

30

40

50

100

150

200

05

o

o

T22T1T91 T7

T21T6 T5 T20

T52 T51

T54

T43

T41T42

T31T32

T4

T11 56.89

%m =

0.04958 kg/s

59.29

%m =

0.04756 kg/s

59.47

%m =

0.03162 kg/s

56.89

%m =

0.03305 kg/s

m =0.02012 kg/s

20

46

23

m =0.01436 kg/s10

m =0.04701 kg/s18

21

m =0.03848 kg/s51

RefrigerantWater-LiBr sorbent solution

Model solutionMeasurement

Steam, cooling water, and chilled water

Pm, 6.352 kPa

Ph, 54.4 kPa

Pl, 0.6871 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.

77

Figure 4-4: Absorption cycle variations with load changes


SATU

ATI

ON

PR

ESSU

RE

(P),

kPa


10 20 30 40 50 60 70 80 90 100 110 120 130 140 150 160 170 180

10

20

30

40

50

60

70

80

90

100

110

120

030%

40%

50%

60%

70%

1

2

3

45

10

20

30

40

50

100

150

200

05

o

o

Point 21

Point 20

Point 91

Point 1

Point 6

Point 5

Point 4

Point 22

Point 51

Point 31

Point 32

Point 43

Point 41

Point 7

Point 54

Model solution for 100%, 83%, 67%, and 34% of design load

Water-LiBr sorbent solutionSteam, cooling water, and chilled water

Chilled/cooling water supply temperature set point

4.2.2 Performance Curve

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.

78

Figure 4-5: Chiller performance curve under various load conditions

0.40

0.50

0.60

0.70

0.80

0.90

1.00

1.10

1.20

1.30

0 2 4 6 8 10 12 14 16 18 20


Coe

ffic

ient

of P

erfo

rman

ce (C

OP)

.

Thermal COP (measurement) Overall COP (measurement)Thermal COP (model)

Thermal COP(model)

Thermal COP(measurement)

Overall COP(measurement)

Figure 4-6: Heat transfer load on each component under various load conditions

0

4

8

12

16

20

24

28

32

0 2 4 6 8 10 12 14 16 18 20


Hea

t tra

nsfe

r on

chi

ller

com

pone

nt (k

W) .

Cooling tower Absorber Evaporator HTRGLTRG Condenser HTHX LTHXHRHX BPHX

Cooling tower

Absorber

Evaporator

HTRG

LTRG

Condenser

HTHX

LTHX

HRHXBPHX

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.

79

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

0

0.001

0.002

0.003

0.004

0.005

0.006

0.007

0.008

0 2 4 6 8 10 12 14 16 18 20


Stea

m fl

ow r

ate

(kg/

s)

Steam flow calculated from power measurementMeasured steam flowMeasured steam flow (corrected)

Steam flow calculated from power

Measured steam flow

Measured steam flow (corrected by the desnsity)

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

80

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

0

0.02

0.04

0.06

0.08

0.1

0.12

0 2 4 6 8 10 12 14 16 18 20


Sorb

ent s

olut

ion

flow

, kg/

s .

Sorbent solution from absorber Sorbent solution to LTRGSorbent solution to HTRG

Sorbent solution from the absorber

Sorbent solution to the HTRG

Sorbent solution to the LTRG

Figure 4-9: Sorbent solution split ratio under various load conditions

0

0.2

0.4

0.6

0.8

1

0 2 4 6 8 10 12 14 16 18 20


Dilu

te so

rben

t sol

utio

n di

stri

butio

n ra

tio, R

.

Sorbent solution distribution ratio

Split ratio, R

81

Figure 4-10 shows that the refrigerant flow vaporized in the evaporator is proportional to the cooling-

load 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

0

0.002

0.004

0.006

0.008

0.01

0 2 4 6 8 10 12 14 16 18 20


Ref

rige

rant

flow

(kg/

s)

Referigerant from LTRG Refrigerant from HTRGRefrigerant vaporizaed in evaporator

Refrigerant vaporized in the evaporator

Refrigerant produced from the HTRG

Refrigerant produced from the LTRG

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 oC at

design conditions; this temperature drops to 0.5 oC 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.

82

Figure 4-11: Refrigerant vaporization temperature under various load conditions

0

0.5

1

1.5

2

2.5

3

3.5

4

0 2 4 6 8 10 12 14 16 18 20Actual cooling load (kW)

Vap

oriz

atio

n te

mpe

ratu

re, T

18, (

o C )

.

Vaporization temperature in the evaporator

T18

Figure 4-12: Sorbent solution composition changes under various load conditions

54

56

58

60

62

64

0 2 4 6 8 10 12 14 16 18 20


Sorb

ent s

olut

ion

conc

entr

atio

n (%

) .

Sorbent form absorber Sorbent from LTRGSorbent from HTRG

Sorbent from HTRG

Sorbent from LTRG

Sorbent from Absorber

4.2.5 Composition Variations

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

83

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, 024 =q . 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

0

0.05

0.1

0.15

0.2

0.25

0.3

0.35

0.4


Vap

or q

ualit

y (q

24)

.

Vapor quality of refrigerant entering condenser

Vapor quality of refrigerant entering Condenser

84

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

0

0.5

1

1.5

2

2.5

3

3.5

4


UA

, kW

/o C

Absorber Evaporator HTRG LTRG Condenser

UA for Evaporator

UA for Absorber

UA for LTRG

UA for HTRG

UA for Condenser

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

85

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

0

0.2

0.4

0.6

0.8

1

1.2

1.4

1.6

0 2 4 6 8 10 12 14 16 18 20Sorbent solution flow (kg/s)

Surf

ace

area

, m2

Evaporator surface area LTRG surface area

Contact area of Evaporator

Contact area of LTRG

4.2.8 Deviation Variations

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

86

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

0%

2%

4%

6%

8%

10%

12%

14%

16%


Ove

rall

and

wei

ghte

d de

viat

ions

.

Overall deviationr Weighted deviation

Overall deviation

Weighted deviation

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.

87

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


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.

88

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

89

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.

90

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 oC, and the temperature difference between

the hot water inlet and the outlet is 10-20 oC. 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 oC. 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.

91

Figure 5-1: Simplified HTRG configurations for natural-gas-driven absorption chiller

400

200

200 200

ChamberCombustion

Fuel ejector

ConvectorExhaust gas

25 ExitExhaust gas

5.2.1.3 Exhaust Gas Absorption Chiller

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-gas-

driven 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 oC from an engine or a

reciprocated gas turbine, 755 oC 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 oC.

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.

92

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.

93

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97

Appendix 1A

102

Appendix 2A

118

Appendix 2B

130

Appendix 3A

150

Appendix 4A

194

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

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