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

Transcription

1 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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4 Copyright Declaration I hereby declare that I am the sole author of this thesis. I authorize Carnegie Mellon University, Pittsburgh, Pennsylvania to lend this thesis to other institutions or individuals for the purpose of scholarly research. I authorize Carnegie Mellon University, Pittsburgh, Pennsylvania to reproduce this thesis by photo copying or by other means, in total or in part, at the request of other institutions or individuals for the purpose of scholarly research. Copyright 2006 by Hongxi Yin i

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

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

7 Table of Contents Copyright 2006 by Hongxi Yin...i Acknowledgment...ii List of Figures...viii List of Tables...x Abstract...xi 1 Introduction Background and Motivation CHP Systems BCHP Systems Heat Utilization Overview of Absorption Chiller Technology Absorption Cycle Analysis Absorption Refrigeration Working Fluids Absorption Refrigeration Operating Conditions Absorption Chiller Cycle Modifications Research Objectives Research Approach The Planning and Installation of Experimental Equipment The Test Program and Experimental Data The Development of Computational Performance Model The Analysis of the Experimental Data Current Absorption Chiller Modeling Studies Absorption Chiller Modeling Approaches The Insufficiencies of Current Absorption Chiller Modeling Studies The Comprehensive Performance Model and its Applications The Chiller Model Description Applications of the Chiller Performance Design Model Preliminary Design Computations...18 iv

8 Detailed Design and Performance Computations Data analysis Chapter Overview Chiller Test System and Program Absorption Chiller System Descriptions Evaporator and Chilled-Water Pump Absorber and Solution Pump High-Temperature Regenerator Low-Temperature Regenerator Condenser Heat Recovery Devices Cooling Tower Vacuum System Absorption Chiller Test Systems System Description Steam Supply System Variable Cooling Load System Instrumentation, Control, and Data Acquisition System Structure of Instrumentation Control System Data Acquisition and Display Instrumentation for the Chiller Instrumentation for the Auxiliary Systems Instrumentation Calibration Controls for the Chiller Chiller Performance and Test Program Chiller Testing Conduct of the Testing Program Chiller Performance Chiller Performance Calculations Chiller Performance under Design Condition Chiller Performance at Reduced Capacity Condition...51 v

9 2.5 Further Information from Chiller Testing Chiller Design and Performance Model Flow Diagram Dűhring Chart Representation T-Q Diagram Calculation Procedure Mass Balance Energy Balance Thermodynamic Property and Equilibrium Relations Heat Transfer Models Overall Heat Transfer Coefficient Model Mass Transfer Models Model Assumptions Model Steps Model-based Experimental Data Analysis Analytical Method Statistical Analysis Procedure Absorption Cycle at Design Condition Overall Deviation Model Analysis Analysis of Cooling-Load Variation Performance Curve Flow Rate Variations Temperature Variations Composition Variations Vapor Quality Variations Heat Transfer Area Variations Deviation Variations Analysis of Other Test Data...86 vi

10 5 Contributions and Areas of Future Research Contributions Areas of Future Research Extended Chiller Model for Multi-Heat Resources Hot Water Absorption Chiller Natural Gas Absorption Chiller Exhaust Gas Absorption Chiller System Integration and Application Chiller Performance Tables for Building Simulation Tools Cost Model...92 References Appendix 1A Appendix 2A Appendix 2B Appendix 3A Appendix 4A Acronyms vii

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

12 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 ix

13 List of Tables Table 1-1: Power generation equipment rejected heat temperature ranges...5 Table 1-2: Water-LiBr absorption chiller thermal energy types and temperature ranges...5 Table 2-1: Component names and corresponding abbreviations...23 Table 2-2: Specifications of the absorption chiller...25 Table 2-4: Control points of the chiller...41 Table 2-3: Instrumentations of the chiller test systems...42 Table 2-5: Input and primary output of the test program...45 Table 2-6: Measurement data of the chiller under design condition...50 Table 2-7: Comparison of chiller performance under design conditions...51 Table 2-8: Primary measurement for chiller input and output...52 Table 3-1: Chiller model state point descriptions...57 Table 3-2: Physical features of heat and mass transfer components...63 Table 3-3: Heat and mass transfer correlations used in the performance model...64 Table 4-1: Measured values and model calculations for 100% and 55% of design load conditions...71 Table 5-1: Heat transfer features of the HTRG of different heating media...90 x

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

15 1 Introduction In the United States, residential and commercial buildings more than 107 million households (2001) [1] and 71.7 billion square feet of commercial floor space (2003) [2] account for more than one-third of the total energy consumption of the country. Significant energy efficiency improvements in heating, ventilation, air conditioning and refrigeration (HVAC&R) systems for residential and light commercial buildings might be achieved by the application of microscale heat-driven absorption chillers for space and ventilation air cooling. Absorption chillers are key components in a building cooling heating and power (BCHP) system to cool space in buildings. They can be driven directly by the thermal energy and heat recovered from various sources, including power generation equipment and solar receiving devices. The combination of heat recovery equipment and heat-driven absorption chillers provides significantly increased overall energy efficiency. Most of today s heating and cooling technologies for buildings, however, are not designed to make use of rejected heat. Performance modeling studies of heat-driven absorption chillers are accordingly limited, contributing to the difficulty of preparing and applying building simulation programs for BCHP system design and performance analysis. This thesis contributes important knowledge and methods for designing and integrating absorption chillers in BCHP systems that reduce energy consumption, decrease operational costs, and improve environmental benefits in residential and light commercial buildings. The gap between experiment and simulation is closed in this thesis because of the availability of a unique microscale absorption chiller and an associated experimental setup. By developing and applying a numerical performance model, a refined understanding of a particular chiller and its operation can provide improved design and modeling tools for heat-driven absorption chillers in general. The approach developed in this thesis will allow developers to simulate the interaction of the BCHP components as a system along with its interactions with: power and other energy supply systems electricity grids indoor air conditions various load profiles 1

16 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 Power generation (40.77 Quads) Electricity (14.2 Quads) Power production waste heat ( Quads) National total energy consumption (99.74 Quads) Residential sector (6.92 Quads) Commercial sector (4.02 Quads) Industrial sector (21.18 Quads) Transportation (27.79 Quads) HVAC, lighting, and others ( Quads) HVAC, lighting, and others ( Quads) Manufacturing processes ( Quads) Residential sector waste heat ( Quads) Commercial sector waste heat ( Quads) Industrial sector waste heat ( Quads) National total waste heat ( Quads) Rejected heat from power generation can be used for building operations. Renewable energy sources (such as solar thermal energy to drive absorption chillers and boilers) combined with advanced distributed electric energy generation can also be used in buildings. Figure 1-2 illustrates the system integration concepts that Volker Hartkopf put forward for the first time [5], for the opportunities of simultaneously achieving energy conservation, using renewable resource, and deploying distributed electricity generation technologies. The building of the future is conceived as a power plant (BAPP) that would generate more energy on site than is brought to it in the form of non-renewable resources. The surplus of energy (power, heating, and cooling) could export to the utility grids or neighboring buildings. 2

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

18 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 Heat Utilization Table 1-1 illustrates the temperature range of rejected thermal energy from typical power generators and heat recovery units. Among them, a solid oxide fuel cell (SOFC) gives the highest exhaust gas 4

19 temperature for heat recovery and utilization. The hot water temperature from solar collectors varies with the type of collector. Solar collectors with parabolic trough reflectors can generate hot water up to 180 o C; integrated compound parabolic collectors, ICPC s, 140 to 160 o C; flat plate collectors, 65 to 90 o C. Table 1-1: Power generation equipment rejected heat temperature ranges No. Power Generation Equipment, Waste Stream Temperature ( o F) Temperature ( o C) 1 Solid Oxide Fuel Cell Exhaust Reciprocal Engine Exhaust Molten Carbonate Fuel Cell Exhaust Gas Turbine Exhaust Microturbine Exhaust HRSG Exhaust Reciprocal Engine Jacket Water Phosphoric Acid Fuel Cell Solar Thermal Collector Table 1-2 shows typical temperature ranges for the heating medium to drive a water-lithium bromide (LiBr) absorption chiller [6]. A single-stage hot-water-driven chiller can use heat at a temperature as low as 75 o C. Tables 1-1 and 1-2 show that an absorption chiller can be found to use heat from a wide range of sources. Because of its higher thermal efficiency, this study focuses on a two-stage absorption chiller and its appropriate sources of rejected heat. Table 1-2: Water-LiBr absorption chiller thermal energy types and temperature ranges No. Heat-driven Absorption Chiller Type Pressure (kpa) Temperature ( o C) 1 Direct-fired fossil fuel (natural gas, oil, LPG etc.) - 1,000 1,800 2 Double-stage exhaust gas Single-stage exhaust gas Double-stage steam 400 1, Single-stage steam Double-stage hot water 350 1, Single-stage hot water 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 (NH 3 -H 2 O) 5

20 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, 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 Absorption Cycle Analysis A chiller produces chilled water by removing heat from it and transferring this heat to a vaporizing refrigerant. The process is illustrated in Figure 1-4 for a conventional vapor compression chiller and in Figure 1-5 for an absorption chiller. In both, the refrigerant liquid flows into an evaporator, evaporates at a reduced pressure and temperature, and absorbs heat from chilled water flowing in a tube through the evaporator. In the vapor compression process, the refrigerant vapor is compressed and condensed at a high-pressure and temperature, transferring heat to cooling water or to the surroundings in a condenser. The high-pressure condensed refrigerant is then returned through the expansion valve to a low-pressure evaporator, once again to absorb heat from the chilled-water flow. 6

21 Figure 1-4: Basic vapor compression chiller cycle Heat rejected to cooling water Condenser Refrigerant expansion valve Work P Compressor Evaporator Heat absorbed from chilled water T 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 Heat rejected to cooling water Heat input Condenser Regenerator Refrigerant expansion valve Solution pump P Solution expansion valve Evaporator Absorber Heat absorbed from chilled water T Heat rejected to cooling water 7

22 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 Absorption Refrigeration Working Fluids An absorption chiller requires two working fluids, a refrigerant and a sorbent solution of the refrigerant. In a water-libr absorption chiller, water is the refrigerant; and water-libr solution, the sorbent. In the absorption chiller cycle the water refrigerant undergoes a phase change in the condenser and evaporator; and the sorbent solution, a change in concentration in the absorber and evaporator. Water is an excellent refrigerant; it has high latent heat. Its cooling effect, however, is limited to temperatures above 0 o C because of freezing. The sorbent, LiBr, is nonvolatile, so a vapor phase in the absorption chiller is always H 2 O. The sorbent solution, water-libr, has a low H 2 O vapor pressure at the temperature of the absorber and high H 2 O vapor pressure at the temperature of the regenerator, facilitating design and operation of the chiller. The advantage of the water-libr pair includes its stability, safety, and high volatility ratio. It has no associated environmental hazard, ozone depletion, or global warming potential. 8

23 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 Absorption Chiller Cycle Modifications Several modifications can be made in the basic absorption chiller cycle to reduce the heat required to operate the chiller and to reduce the extent of heat transfer surface incorporated in the machine. Countercurrent heat interchange can be arranged between the two sorbent solution flows connecting the low-temperature absorber and the high-temperature regenerator. This interchange can significantly reduce the heat quantities involved in the operation of both; less heat will need to be supplied to the regenerator, and less heat will need to be removed form the absorber. 9

24 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 Heat input Condenser Heat to LTRG Regenerator High-temp. heat exchanger Heat rejected to cooling water P Refrigerant combiner Refrigerant expansion valve Recirculation pump Evaporator Condenser Solution splitter Solution pump Absorber Low-temp. regenerator Low-temp. heat exchanger Solution combiner Solution expansionvalve Heat absorbed from chilled water T Heat rejected to cooling water The revised flow diagrams illustrating these absorption chiller flow diagrams are shown in Figures 1-6 and 1-7. The flow of the sorbent solution from the absorber to the two regenerators can be either parallel or in series. In a parallel flow arrangement, the dilute solution from the absorber is pumped to both the high-temperature and the lower-temperature regenerators in parallel, as shown in Figure 1-6. Concentrated solutions from both regenerators are recombined and returned to the absorber. In a series flow arrangement, the solution from the absorber is first pumped to the high-temperature, highpressure regenerator; and the partially concentrated sorbent solution then flows to the lower-pressure, lower-temperature regenerator, as shown in Figure

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

26 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: The Planning and Installation of Experimental Equipment A microscale BCHP energy supply system (ESS) has been designed for the IW, a 6,500 ft 2 office environment at CMU, to provide power and space cooling heating, and ventilation. As the first stage in realizing this overall system, a 16kW steam-driven water-libr absorption chiller was installed in the south section of the IW. This chiller 12

27 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 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 The Development of Computational Performance Model On the basis of scientific and engineering principles and the specific configurations of the chiller, a detailed computational performance model was constructed to evaluate the chiller performance under various operating conditions. This model was developed for the chiller to further refine the 13

28 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 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 Absorption Chiller Modeling Approaches In the past decades, computer models have been developed to investigate the performance of various water-libr absorption chiller cycles. Among these models, some [8, 9] are system specific for particular machines, flow configurations, and working materials. Others [10, 11, 12] are generic to handle various potential absorption cycles with one modularized model. The system specific models are performance models aimed at simulating a specific design and investigating its performance under various operation conditions; the generic models are aimed at exploring novel absorption cycles and evaluating their performance under various boundary conditions. The advantage of system specific or performance models is that the model simulates the configuration of absorption chiller systems in detail. Thermodynamic cycle, heat, and mass transfer characteristics can be investigated on the basis of the physical details of the chiller. In these studies the simulation 14

29 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 The Insufficiencies of Current Absorption Chiller Modeling Studies First, the existing simulation models of water-libr absorption chillers focus on relatively large-scale installations for commercial buildings or for district energy centers. None of the studies consider microscale absorption chillers with a cooling capacity less than 17 kw for residential or light commercial applications. There are, theoretically, no distinctions between the large-scale and the microscale absorption chillers in terms of scientific and engineering principles, but the design criteria and operating conditions for microscale absorption chillers are different from those for the large capacity chillers. For instance, microscale absorption chillers for residential application must provide a more compact design and include a heat rejection unit, such as a cooling tower. Second, at present, nearly all performance models of absorption chillers have been numerical simulations without significant experimental validation under design and off-design conditions. It has been difficult to install a commercial absorption chiller in a university laboratory because of their large capacity. The requirements for operation and test of commercial chillers and their limited 15

30 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 The Chiller Model Description In the model, the absorption chiller is composed of the following components: an evaporator: a countercurrent two-phase coiled tube heat exchanger an absorber: a countercurrent two-phase coiled tube mass and heat exchanger two regenerators: one high temperature, one intermediate temperature: well mixed, two-phase boiling coiled tube heat exchangers a condenser: a countercurrent heat exchanger two plate heat interchangers: countercurrent single-phase heat exchangers two tube and shell heat recovery exchangers: countercurrent single-phase exchangers three pumps: a sorbent pump, a refrigerant pump, and a chilled-water pump associated spray nozzles, trap, valves, and pipe fittings The cooling tower associated with this chiller includes the following components: a countercurrent plate column two-phase mass and heat exchanger 16

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

32 inlet and outlet temperatures and flow, the cooling-water supply temperature and flow, the steam supply pressure and flow 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 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 o 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. the operating pressure of the condenser (with its corresponding pressure) is that of the intermediate temperature regenerator. heat transfer in the countercurrent interchangers and heat recovery exchangers is maximized by equal stream temperatures at one end of the exchanger. Preliminary design computations have proved useful in exploring the effects of various chiller configurations, component characteristics, and external operating conditions on the heating and cooling requirements, internal conditions, and power requirements of a chiller. A preliminary design model was programmed to estimate the heat/mass transfer areas of the chiller components; this was a first step in constructing a comprehensive performance model. If the design 18

33 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 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. 19

34 1.6.3 Data analysis The outputs of performance model were also compared with test data under various conditions by changing the operation parameters. Based on the performance model, the accuracy and reliability of the experimental data were assessed, and the model assumptions were validated. Measured chiller external and internal operating conditions can be used to compare those calculated from the chiller model with the results when the model is supplied with the external operating conditions and component areas. These comparisons can be used to evaluate the accuracy of measurements and to consider the validity of the model including the assumptions on which the model is based. Such comparisons and conclusions based on those comparisons are discussed in detail in chapter 3. Several measurements that differ significantly from model predicted values have been analyzed, and the procedures for correcting these measurements have been proposed and applied. On the basis of the model developed in this thesis, the validated model can then be extended to incorporate the following heat sources: hot water from solar thermal or heat recovery equipment natural gas exhaust gas from gas turbine, engine generator, and fuel cells The validated models, as a tool, can be integrated with the IW model to evaluate overall BCHP system performance incorporating with a cost model. 1.7 Chapter Overview This thesis contains five chapters followed by references, and appendixes, and list of abbreviations. Chapter 1, Introduction introduces the background and motivation of this dissertation and summarizes the research objectives of this chapter. The emerging features of the modern absorption chiller industry are summarized in appendix 1A. Chapter 2, Chiller Test System and Performance introduces the chiller and examines the overall experimental system setups. It presents detailed information concerning the instrumentation and control for the chiller and its auxiliary systems. The chiller testing program, measured experimental 20

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

36 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 System Descriptions The absorption chiller installed in the IW is a steam-driven, two-stage, water-libr, parallel-sorbentflow series-cooling-water flow chiller with a cooling tower. This chiller, provided by Broad Co., has a 16kW rated cooling capacity. It is the smallest absorption chiller available in the existing market and the only steam-driven absorption chiller of such capacity in the world. Figure 2-1 shows the absorption chiller installed on a platform adjacent to the IW. The chilled-water supply and return, steam supply, condensate return, power, and city water lines connect with the chiller at the bottom left. Figure 2-2 is a schematic flow diagram recreated from the manufacturer s brochure for a commercial natural-gas direct-fired chiller; this flow diagram shows all the heat and mass transfer components, pumps, and pipe fittings. It also indicates the design values for temperatures throughout the chiller. The measurement and control features of the chiller will be discussed in conjunction with a detailed process and instrumentation (P&I) diagram in the section that follows. The components and parts indicated in Figure 2-2 are listed in Table 2-1. Figure 2-1: Absorption chiller installed in the IW 22

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

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

39 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 Solution Power Steam Chilled water Name Quantity Unit Cooling capacity 16 kw Chilled-water return temperature 14 o C Chilled-water supply temperature 7 Chilled-water flow rate 2 m 3 /h Chilled-water pump head 8 mh 2 O Rated steam pressure, absolute 0.7 mpa Steam pressure limit, absolute 0.9 mpa Maximum steam consumption 24 kg/h Power voltage 220 V Power frequency 60 Hz Maximum power consumption 1 kw Water-LiBr sorbent solution mass 65 Kg Water-LiBr sorbent concentration 55 % o C 25

40 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 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 kpa, and low temperature, about 3-4 o C. The vaporization absorbs heat from the chilled water flowing through the evaporator coil, cooling this flow from 14 o C to 7 o C. Figure 2-4: Configuration of the lower vessel 26

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

42 The design of the HTRG differs depending both on the heating medium, gas, or liquid, and on its temperature. Many forms of thermal energy can be used in the HTRG to drive a two-stage absorption chiller, such as steam, hot water, exhaust gas, natural gas, oil, and liquid pressurized gas. In this section, only the steam-driven HTRG is discussed; other kinds of heat sources - natural gas, hot water, and exhaust gas - are discussed in the sections that follow. The water-libr sorbent, reconcentrated in the regenerators, returns to the absorber through flow restrictions that assist in maintaining appropriate liquid levels to submerge the heat transfer coils in the regenerators. The solution pump frequency is adjusted to maintain a constant level in the HTRG. Both the HTRG and the LTRG use water vapor as a heat resource; they have similar functions and structure. The heat transfer process includes condensation inside the tubes and boiling on the outer surface of these tubes. Figure 2-5: Configuration of the upper vessel 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. 28

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

44 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 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 Cooling Tower A cooling tower is widely used to dissipate reject heat from a water-cooled air-conditioning system to the surroundings. This Broad absorption chiller has a built-in cooling tower, as shown in Figures 2-2 and 2-6. Its compact design facilitates chiller installation and operation. The cooling water in the chiller flows in series through the absorber, the condenser, and then through the cooling tower. This arrangement provides for a minimum operating temperature in the absorber that is required to achieve a low chilled-water temperature; the high flow in both the absorber and the condenser provides for high heat transfer coefficients in these components. The recirculating cooling water flows down vertical plates in countercurrent contact with upward-flowing ambient air. Evaporation of a small portion of the water flowing downward through the cooling tower reduces its temperature; makeup water added to the cooling tower replaces that evaporated. The air temperature is also reduced, but the humidity increases markedly. Thus, the cooling tower functions as a two-phase countercurrent heat and mass exchanger. 30

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

46 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 System Description A system was set up to test the Broad BCT 16 absorption chiller and to evaluate its performance under a wide range of external and internal operating conditions. This test system, shown in Figure 2-7, comprises the following equipment in addition to the chiller: a steam supply a variable cooling load an instrumentation, control, and data acquisition system In Figure 2-7, the absorption chiller is in the middle. It is connected with the steam supply system on the left and the variable cooling load system on the right. The necessary control and instrumentation for the overall system has been installed to operate this test system and to process the data it provides. The measurement data are used both to monitor the system status and to calculate chiller performance. 32

47 Figure 2-7: Simplified flow diagram of the chiller test system ALC P4 T22 SS Additional sensors to chiller F2 P3 T21 IV5 IV2 CTW WS BFT P7 BFP T25 ESB Absorption chiller BHWSV HWR TLHX Steam supply system CR T23 P5 F1 T20 P2 HWS IV4 IV3 Variable cooling load Absorption chiller Cooling loads Steam supply 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 33

48 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 Variable Cooling Load System Systematic testing of the chiller requires a load that can be adjusted independently and maintained constant during a test run. This load is provided by a shell-and-tube heat exchanger fed with water at 34

49 80 o C to the shell from the building hot water grid. The flow of chilled water from the chiller outlet to the tubes of the load exchanger is controlled by a valve to achieve a desired flow set point. The flow of hot water to the exchanger is also controlled by a valve (BHWSV) to maintain a desired set point temperature for the chilled water at the inlet to the chiller. Under design conditions, the chilled water flows in the chiller at a rate at 2 m 3 /h and a temperature of 14 o C. The chiller cools the chilled water to 7 o C. Figure 2-8 shows photographs of the overall system. The photo on the left indicates the ESB, BFT, and some chilled-water supply and return pipes. The photo on the top right shows the absorption chiller installed on the deck adjacent to the IW. The picture on the bottom right shows the variable cooling load system 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 Structure of Instrumentation Control System The structure of the overall control and instrumentation system is illustrated in Figure 2-9. The left side is the internal control system from the chiller manufacturer, and the right side is the ALC control system for the steam supply and variable cooling load systems. The additional sensors used to monitor the chiller internal conditions were also implemented through the ALC control system. The ALC control system is one of the basic platforms for building an integrated BCHP system. It can not only perform regular control and data acquisition functions, but it can also perform more complicated tasks such as system diagnostics and optimization when the overall system becomes more complicated. With this automation control system, the overall chiller test system can be started up, shut down, and adjusted automatically or manually through a computer. The operating conditions can be displayed on a graphic interface, and the measured data can be collected for further analysis. 35

50 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 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 The monitoring interface is a small and independent software package with data collection functions. It is a good tool for monitoring the status of the chiller instantly and displaying the information graphically. Through the interface, the operator 36

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

52 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 Instrumentation for the Chiller Figure 2-12 is the process and instrumentation (P&I) diagram for the chiller. This figure illustrates chiller components, piping, and the measurement and control points. The chiller components and configurations have been described in Figure 2-2, Table 2-1, and section 2.1. This section discussed the instrumentation and control of the chiller. The absorption chiller has its own controls and instrumentation from the manufacturer; additional sensors are installed for the chiller by the ALC during chiller installation to study its internal operation conditions. In Figure 2-12, the sensors from 38

53 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 Hot Humid Air ME-LGR 25 Steam T7 SF T10 Controller P1 T14 L3 LTRG HTRG Condenser T15 T17 T18 L1 T16 F6 SV CHSV Condensate T8 T11 T5 CWBPV T33 T13 T19 HTHX LTHX CTF To auxilary system CHWS T9 T2 HRHX L2 Evaporator Absorber DV T6 RBPSV L4 L5 CWDD RP RPH CWF CHWR B1 T1 B2 SP T3 T32 Drain CHWP T12 CWP City water CTWS Air 39

54 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 Instrumentation for the Auxiliary Systems The steam supply system and the variable cooling-load system were discussed in sections and 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 Instrumentation Calibration These sensors provided by the ALC were calibrated on site by the following methods: the temperature sensor readings were calibrated in ice water and boiling water. the condensate from the chiller was collected in a barrel; the weight of the condensate was measured every 15 minutes for 2 hours. The weight of condensate was compared with the measured values of the steam flow meter. the chilled-water flow meter piping configuration was sent back to the flow meter manufacturer for calibration; the suggested deviation has been applied to the chilled-water flow measurement. 40

55 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 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 The features of these control components are discussed in appendix 2.A. Table 2-3: Control points of the chiller Abbrev. Name Signal Category CHWP Chilled-water pump Digital On/off CWP Cooling-water pump Digital On/off SP Solution pump Analog Quantity control RP Refrigerant pump Digital Quantity control CTF Cooling-tower fan Analog Temperature control SV Steam valve Analog Operation time control RBPSV Refrigerant by-pass valve Digital On/off CTS City-water switch Digital Quantity control CWDD Cooling-water drain device Digital On/off CWDV Cooling-water detergent valve Digital On/off CWBPV Cooling-water by-pass valve Analog Temperature control CTF Cooling-tower fan Analog Temperature control RPH Refrigerant-pump heater Digital Temperature control 41

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

57 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 Chiller Testing Chiller Test An individual chiller test was conducted by setting the six operating conditions that are the primary input to the test system, all external to the chiller: the pressure of the saturated steam supply the flow rate, inlet, and outlet temperature of the chilled water the flow rate and inlet temperature of the cooling water. Ordinarily, the chiller cooling- water pump maintains a constant flow; and the air fan maintains a constant supply temperature by varying its speed in response to the cooling load and the ambient air conditions. To test the 43

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

59 Table 2-5: Input and primary output of the test program CHW return T CHW flow Inputs Primary Outputs Calculated performance CW CW Steam Steam CHW Cooling supply T flow pressure Flow supply COP load o C kg/s o C kg/s kpa kg/s o C kw Design condition CHW return T 8 14 Design Design Design Design CHW flow Design Design Design Design CW supply T Design Design Design Design CW flow Design Design Design Design Steam pressure Design Design Design Design Conduct of the Testing Program In the testing program, various procedures were used to adjust the six operating conditions. The chilled-water return temperature was varied by adjusting the hot water supply flow to the variable load. The chilled-water flow was varied by adjusting a ball valve in the chilled-water supply pipe, but when the chilled-water flow was reduced below the design flow rate for the chiller, a water cut warning was reported, and the chiller automatically executed its shutdown procedure. Only design and higher chilled-water flows have been tested. The cooling-water supply temperature was varied by adjusting the air fan speed. The cooling-water flow rate was varied by blocking a portion of the cooling-water filter located at the bottom of the cooling tower. The steam supply pressure was varied by adjusting the pressure set point in the boiler control system. Fewer tests have been performed at operating conditions that result in loads below the design values. When the cooling load is below 50% of the design load, it has proved difficult to obtain stable data for analysis because of on-off cycling of the steam valve. 2.4 Chiller Performance The chiller performance has been calculated on the basis of the resulting measurements. The chiller performance has been compared with the chiller specification data Broad provided. Results show that 45

60 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 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 Q cooling CHW p ( T T ) = m& C (Equation 2-1) CHWR CHWS Where m& CHW is the flow rate of the chilled water T CHWR is the temperature of the chilled water entering the chiller T is the temperature of the chilled water leaving the chiller CHWS C is the specific heat of water. p The COP of the chiller is conventionally Qcooling COP thermal = (Equation 2-2) Q heat Where, Q m ( h h ) heat = & = the heat delivered to the chiller by steam steam steam condensate condensation; this quantity can also be estimated from the electrical power consumption in the steam boiler. m& steam is the flow rate of the steam supply h steam and h condensate are the enthalpies of the steam supply and the condensate, respectively. An overall COP can be defined that includes both the thermal and the electrical energy supplied to the chiller Qcooling COP overall = (Equation 2.3) Q energy heat energy Q Q + E = (Equation 2-3) E = V * I power power power power 46

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