Experimental Study of Zeotropic Refrigerant Mixture HFC-407C As a Replacement for HCFC-22 in Refrigeration and Air Conditioning Systems

Experimental study of Zeotropic refrigerant mixture HFC-407C as a replacement for HCFC-22 in Refrigeration and air conditioning systems

Changiz M. Tolouee

B. Sc. (Mech. Eng.) and M. Sc. (Mech. Eng.)

A thesis submitted for the degree of

DOCTOR OF PHILOSOPHY

School of Engineering and Science Swinburne University of Technology

2006 1

DECLARATION

This thesis contains no material which has been accepted for the award of any other degree or diploma, except where due reference is made in the text of the thesis. To the best of my knowledge, this thesis contains no material previously published or written by another person except where due reference is made in the text of the thesis.

Signed …………………………….

2 ABSTRACT

HCFC-22 is the world’s most widely used refrigerant. It serves in both residential and commercial applications, from small window units to large water chillers, and everything in between. Its particular combination of efficiency, capacity and pressure has made it a popular choice for equipment designers. Nevertheless, it does have some ODP, so international law set forth in the Montreal Protocol and its Copenhagen and Vienna amendments have put HCFC-22 on a phase out schedule. In developed countries, production of HCFC-22 will end no later than the year 2030.

Zeotropic blend HFC-407C has been established as a drop-in alternative for HCFC-22 in the industry due to their zero Ozone Depletion Potential (ODP) and similarities in thermodynamic properties and performance. However, when a system is charged with a zeotropic mixture, it raises concerns about temperature glide at two-phase state, differential oil solubility and internal composition shift.

Not enough research has been done to cover all aspects of alternative refrigerants applications in the systems. This research intended to explore behavior of this alternative refrigerants compare to HCFC-22 and challenges facing the industry in design, operation service and maintenance of these equipments.

The purpose of this research is to investigate behavior of R407C refrigerant in chiller systems. This includes performance and efficiency variations when it replaces R22 in an existing system as well as challenges involved maintaining the system charged with R407C. It is a common practice in the industry these days to evacuate and completely recharge when part of the new refrigerant blend was leaked from the system. This has proved to be extremely costly exercise with grave environmental ramifications.

This research is intended to address challenges faced in the real world and practical terms.

Theoretical and experimental approaches used as a methodology in this work. The system mathematically modeled to predict detailed system performance and effect of the leak at various conditions. To make this feasible and accurate enough, two separate approaches made, first system performance for pure R22 and R407C, and second

3 system subjected to range of leak fractions. The earlier model was relatively straight forward when compared to the latter. Modeling a system charged with R407C ternary mixture and subjected to range of leaks posed enormous challenges.

A sophisticated experimental test apparatus was also designed and built. Comprehensive and detailed tests at various conditions were conducted with special attention on instrumental accuracy and correct methodology.

The first part has been successfully modeled and predicted all the factors and performance with excellent accuracy when compared to the test results. In these approaches pure refrigerants R22 and R407C were used and simulated the system behavior at range of conditions.

However, the second part was the most challenging ever. Comprehensive leak process simulations produced trends of R32/R125/R134a composition change as function of rate of leak. Starting from this point, equations have been created to represent the composition change as function of percentage of the leak. The system thermodynamic cycle was also modeled to calculate capacity, power input and COP at the range of the conditions. Despite many affecting parameters and complexity of the model, the mathematical model successfully predicted the test outcome with a very reasonable accuracy, averaging around 3% with some times reaching to 5 to 6%.

On the experimental stage the system charged with the new HFC-407C was deliberately subjected to refrigerant leak at various leak stages. The aim was to objectively determine to what extend the gas leak can be still acceptable without going through the expensive complete gas charge. The effect of leak was tested and verified at 10% steps, from 10% up to 50% mass fraction for the total charge.

It has been observed that at the leaks beyond 30%, the adverse effect on the capacity becomes more significant, from 8 to about 15% decrease. While the power input decreased at slower pace, from 3% up to about 8% depending on the test conditions. This translated to COP decrease ranging from 4 to about 7%. This capacity loss and efficiency decrease are significant figures which suggests that the system, here chiller, can not be allowed to degrade the performance to that extend and still continue operating.

ACKNOWLEDGMENTS

I would like to express my gratitude to my supervisor Prof. Yos Morsi, for his encouragement, guidance and advice throughout the course of this research project. His supervision and contribution to this research work has been very valuable. Prof. Yos

Morsi’s initiative in establishing link between the Swinburne University of Technology and the refrigeration and air-conditioning industry was a key factor in continuation of this research work.

I would like to thank Dr Wei Yang for his advice, guidance and full support in bringing together all necessary facilities from literature to building test rig. His continual support and dedication throughout the course of this work proved invaluable.

The early stage of this work goes back to my research work in the University of New

South Wales where I was guided by Professor Masud Behnia and Professor Eddie

Leonardi to whom I would like to forward my appreciation for their contribution.

Thanks to Mr. Giovanni Giofre for his assistance during construction and commissioning of the test apparatus.

Thanks to all workshop staff for the construction and maintenance of the experimental apparatus.

TABLE OF CONTENTS

LIST OF FIGURES ...... 8

LIST OF TABLES ...... 11

NOMENCLATURE...... 12

1 INTRODUCTION ...... 14 1.1 PREFACE ...... 14 1.2 WORLDWIDE OZONE DEPELETION LEGISLATION ...... 14 1.2.1 Kyoto Protocol ...... 18 1.2.2 HCFC-22 (R22) Substitutes...... 18 1.2.3 Properties of HCFC-22 (R22) Substitutes ...... 19 1.2.4 Retrofitting Existing systems ...... 20 1.2.5 Refrigerant Solutions for Today’s Environmental Challenges ...... 21 1.2.6 HCFC-22 Phase out and Recycling ...... 23 1.3 PURPOSE OF THIS RESEARCH ...... 23

2 LITERATURE REVIEW ...... 25

3 THEORY ...... 36 3.1 INTRODUCTION ...... 36 3.2 AZEOTROPIC BLEND...... 37 3.3 ZEOTROPIC BLENDS ...... 39 3.4 THERMODYNAMIC PROPERTIES DETERMINATION ...... 42 3.4.1 Redlich-Kwong-Soave (RKS) (1980) Equation of State ...... 42 3.4.2 Vapor Density ...... 44 3.4.3 Liquid Density...... 45 3.4.4 Enthalpy ...... 45 3.4.5 Entropy...... 46 3.5 OTHER EQUATIONS OF STATE...... 47 3.5.1 Pure Refrigerants...... 47 3.5.2 Mixed Refrigerants ...... 48 3.6 CYCLE COMPONENTS ...... 50 3.6.1 Compressor ...... 50 3.6.2 Evaporator...... 52 3.6.3 Condenser...... 52 3.6.4 Expansion Valve ...... 53 3.6.5 Suction and discharge Lines ...... 53 3.7 Determination of the Polytropic Exponent ...... 53 3.8 CONSTANTS ...... 54 3.9 RESULTS ...... 56

4 CHILLER TECHNOLOGY...... 57

6 4.1 INTRODUCTION ...... 57 4.2 SYSTEM FUNDAMENTALS ...... 57 4.3 COMPRESSORS ...... 60

5 EXPERIMENTAL APPARATUS...... 75 5.1 THE TEST APPARATUS ...... 75 5.1.1 Compressor ...... 75 5.1.2 Condenser...... 75 5.1.3 Evaporator...... 76 5.2 INSTRUMENTATION AND DATA LOGGING...... 81 5.2.1 Thermocouples...... 81 5.2.2 Pressure Transducers...... 82 5.2.3 Data logging and calibration of the sensors ...... 82 5.3 TEST PROCEDURE ...... 83

6 TEST RESULTS...... 86 6.1 ANALYSIS OF THE COLLECTED DATA...... 86

7 SIMULATION AND MODELING RESULTS ...... 104 7.1 R22 and Pure R407C...... 104 7.2 Modeling of the system charged with R407C and then subjected to leak ....109 7.2.1 Calculating the new mixture properties after the leak ...... 110 7.2.2 Calculating the system performance including capacity, power input and COP 111

8 CONCLUSIONS AND RECOMMENDATIONS ...... 119 8.1 GENERAL CONCLUSIONS ...... 119 8.2 RECOMMENDATIONS FOR FUTURE WORK ...... 121

REFERENCES...... 123

7 LIST OF FIGURES

Figure 1-1 Ozone distribution over earth’s atmosphere, as on August 1, 1998...... 15 Figure 1-2 European Union HCFC production phase out schedule...... 16 Figure 1-3 Mean Condensing Pressure comparison of the alternative refrigerants to HCFC-22 at 50 C CT...... 21 Figure 2-1 Various mixture combinations of the Chlorine free R125/R32/R134a gases26 Figure 2-2 Schematic diagram of vapor and liquid leak models ...... 33 Figure 3-1 Ideal binary blends at constant temperature...... 36 Figure 3-2 Ideal binary blend at constant pressure ...... 37 Figure 3-3 Azeotropic Blend with Minimum Boiling point variety, Temperature vs. concentration ...... 38 Figure 3-4 Azeotropic binary blend with minimum boiling point, pressure vs. concentration ...... 38 Figure 3-5 Zeoptropic blend vapor compression cycle using non-isothermal phase change...... 39 Figure 3-6 Graph illustration of a vapor compression cycle using zeotropic blend with specific composition...... 40 Figure 3-7 Ternary blend at constant pressure...... 41 Figure 3-8 Compression process diagram...... 52 Figure 3-9 Various compression processes represented on a P – V diagram ...... 54 Figure 4-1 A typical water-cooled chiller system schematic with shell & tube condenser and evaporator, and with screw or semi-hermetic reciprocating compressor 58 Figure 4-2 A typical Two stage centrifugal chiller (water-cooled) with economizer.....59 .Figure 4-3 Compression process of a typical reciprocating compressor ...... 61 Figure 4-4 Capacity variation as a function of SST of a typical reciprocating compressor ...... 61 Figure 4-6 A typical dual rotor – male and female - screw compressor assembly ...... 63 Figure 4-7 A typical dual rotor – male and female - screw compression concept...... 63 Figure 4-8 Suction, trap, compression and discharge process of the scroll concept...... 64 Figure 4-9 Rotation geometry of motor shaft, journal bearing and driven scroll ...... 65 Figure 4-10 Capacity variation of positive displacement compressors versus SST at three constant SDTs ...... 66 Figure 4-11 A typical single stage centrifugal compressor cross section ...... 67 Figure 4-12 Velocity increase by impeller and conversion of velocity into static pressure in a centrifugal compression process ...... 68 Figure 4-13 Performance of a typical centrifugal compressor over a range of inlet vane positions ...... 69 Figure 4-14 A cut-away section view of the new totally oil-free compressor...... 70 Figure 4-15 Cross section view of two stage centrifugal gas compression assembly ....71 Figure 4-16 Shaft and impellers assembled with magnetic bearings...... 72 Figure 4-17 Cross section schematic of the front radial, rear radial and axial bearings with sensor rings...... 72 Figure 4-18 Operating map of the new totally oil free and variable speed centrifugal compressor ...... 73 Figure 5-1 Heat transfer coefficient as function of boiling regimes and quality inside a tube, Cengel 1998...... 77 Figure 5-2 Schematic of the experimental apparatus...... 78

8 Figure 5-3 Evaporator piping and sensor arrangement isometric schematic ...... 78 Figure 5-4 Temperature sensor (Thermocouple) location schematic ...... 79 Figure 5-5 Evaporator Thermocouple tip installation schematic...... 79 Figure 5-6 General view of the test rig ...... 80 Figure 5-7 Data acquisition and instrumentation set up of the experimental apparatus .81 Figure 6-1 Evaporator pressure data acquisition for pure R22 ...... 86 Figure 6-2 Evaporator temperature data acquisition vs. time for pure R22...... 87 Figure 6-3 Evaporator temperature data acquisition for pure R407C...... 87 Figure 6-4 Evaporator temperature data acquisition for 10% leaked R407C ...... 88 Figure 6-5 Evaporator temperature data acquisition for 30% leaked R407C ...... 88 Figure 6-6 Pressure data acquisition for pure R407C ...... 89 Figure 6-7 Condensing pressure data acquisition for pure R407C ...... 89 Figure 6-8 Condensing, average evaporating and suction pressure data acquisition for 10% leaked R407C...... 90 Figure 6-9 Average evaporating and suction pressure data acquisition for 10% leaked R407C...... 90 Figure 6-10 Average evaporating and suction pressure data acquisition for 30% leaked R407C...... 91 Figure 6-11 Condensing pressure data acquisition for 30% leaked R407C ...... 91 Figure 6-12 Average evaporating and suction pressure data acquisition for 40% leaked R407C...... 92 Figure 6-13 Condensing pressure data acquisition for pure 30% leaked R407C ...... 92 Figure 6-14 Refrigerant temperature along the evaporator at CT 40 C for pure R22....93 Figure 6-15 Refrigerant temperature along the evaporator for pure R22 ...... 93 Figure 6-16 Suction and average evaporating pressures vs. CHWLT at CT 40 and 50 C for R22...... 94 Figure 6-17 Suction and average evaporating pressure vs. CHWLT at 40 C and 50 C CT for 10% leaked R407C ...... 95 Figure 6-18 Suction and average evaporating pressure vs. CHWLT at 40 and 50 C CT for 30% leaked and topped up...... 96 Figure 6-19 Suction and average evaporating pressure vs. CHWLT at CT 40 and 50 C for 50% leaked and topped up...... 97 Figure 6-20 Average evaporating pressure variation vs. pure R22 and percentage leaked R407C at various CHWLT for 40 C CT ...... 97 Figure 6-21 Average evaporating pressure variation vs. pure R22 and percentage leaked R407C at various CHWLT for 50 C CT ...... 98 Figure 6-22 cooling capacity variation vs. pure R22 and percentage leaked R407C at various CHWLT for 40 C CT ...... 99 Figure 6-23 cooling capacity variation vs. pure R22 and percentage leaked R407C at various CHWLT for 50 C CT ...... 99 Figure 6-24 Power input variation vs. pure R22 and percentage leaked R407C at various CHWLT for 40 C CT ...... 100 Figure 6-25 Cooling capacity variation vs. pure R22 and percentage leaked R407C at various CHWL for 50 C CT...... 101 Figure 6-26 Coefficient of Performance (COP) variation vs. pure R22 and percentage leaked R407C at various CHWLT for 40 C CT...... 102 Figure 6-27 Coefficient of Performance (COP) variation vs. pure R22 and percentage leaked R407C at various CHWLT for 50 C CT...... 103 Figure 7-1 Genetron software calculation results screens for R22 ...... 105 Figure 7-2 Genetron software calculation results screens for R407C ...... 106

9 Figure 7-3 Capacity and power input comparisons, Genetron vs. test results, R22 at 40 C CT ...... 107 Figure 7-4 Capacity and power input comparisons, Genetron vs. test results, R22 at 50 C CT ...... 107 Figure 7-5 Capacity and power input comparisons, Genetron vs. test results, R407C at 40 C CT ...... 108 Figure 7-6 Capacity and power input comparisons, Genetron vs. test results, R22 at 40 C CT ...... 108 Figure 7-7 Mass fraction change of the ternary blend R134a/R32/R125 and polynomial curve fitted models ...... 110 Figure 7-8 Comparison of the capacities from the test results and the modeling at various CHWLTs and various leak percentages, 40 C CT ...... 112 Figure 7-9 Comparison of the capacities from the test results and the modeling at various CHWLTs and various leak percentages, 50 C CT ...... 113 Figure 7-10 Power input comparison from the test results and the modeling at various CHWLT temperatures and various leak percentages, 40 C CT ...... 114 Figure 7-11 Power input comparison from the test results and the modeling at various CHWLTs and various leak percentages, 50 C CT ...... 115 Figure 7-12 COP comparison from the test results and the modeling at various CHWLTs and various leak percentages, 40 C CT ...... 116 Figure 7-13 COP comparison from the test results and the modeling at various CHWLTs and various leak percentages, 50 C condensing ...... 117

10 LIST OF TABLES

Table 1–1 Montreal Protocol Production Caps...... 16 Table 1–2 European Union ban schedule on use of HCFC-22 in the new equipment....17 Table 1–3 The United States phase out schedule for the HCFCs ...... 17 Table 1–4 Characteristics of the HCFC22 and potential substitutes...... 20 Table 3–1 m and n constants...... 43

Table 3–2 Calculated kij values...... 44 Table 3–3 Constants for vapor pressure equations ...... 55 Table 3–4 Constants for ideal gas heat capacity equation ...... 55 Table 4–1 Chiller Technology alternatives – Vapor Compression cycle ...... 60 Table 5–1 Details of the conditions of series of the tests conducted upon...... 85 Table 7–1 The calculated new compositions after the system subjected to various leak percentages without adding...... 111 Table 7–2 The calculated new compositions after the system subjected to various leak percentages and topping up the system ...... 111

11 NOMENCLATURE

Avg EP Average Evaporating Pressure c Compressor clearance fraction CFC Chloro Floro Carbon Const Constant value COP Coefficient of Performance Cp Specific heat at constant pressure Cv Specific heat at constant volume CP Condensing Pressure CT Condensing Temperature DX Direct Expansion EP Evaporating Pressure ET Evaporating Temperature GWP Global Warming Potential GTD Gliding Temperature Difference h Enthalpy HCFC Hydro Chloro Floro Carbon HFC Hydro Floro Carbon HVACR Heating, Ventilation, Air Conditioning and Refrigeration ID Internal Diameter IGV Inlet Guide Vanes LCHWT Leaving Chilled Water Temperature LLSL-HX Liquid Line Suction Line Heat Exchanger N Number of cylinders n Polytropic Exponent ODP Ozone Depletion Potential P Pressure Pa Pressure component a Pb Pressure component b Pc Critical pressure PD Piston displacement Pt Total pressure

12 Q Heat transfer rate RPM Compressor speed (rev/min) S Entropy SP Suction Pressure T Temperature Tc Critical temperature TEWI Total Environment Warming Impact UV Ultra Violet v Specific volume W Compressor power X Mole fraction of liquid in mixture

X i 134aR Mass fraction of R134a in liquid after the leak

X i R125 Mass fraction of R125 in liquid after the leak

X i R32 Mass fraction of R32 in liquid after the leak

XNi 134aR New mass fraction of R134a in the liquid after leak and topping up

XNi R125 New mass fraction of R125 in the liquid after leak and topping up

XNi R32 New mass fraction of R32 in the liquid after leak and topping up

LRi Leak mass fraction

Y Mole fraction of vapor in mixture Z Compressibility factor

1 INTRODUCTION

1.1 PREFACE

The HVACR industry faces two major environmental challenges today: stratospheric ozone depletion and global climate change.

Stratospheric Ozone Depletion is believed to be caused by the release of certain manmade ozone-depleting chemicals into the atmosphere. A compromised ozone layer results in increased ultraviolet (UV) radiation reaching the earth’s surface, which can have wide ranging health effects. Legislation has been enacted worldwide through the Montreal Protocol to phase out the production of these chemicals. This phase out is in progress, and the scientific evidence indicates repair to the ozone layer is underway.

Global climate change is believed to be caused by buildup of greenhouse gases in the atmosphere. The primary greenhouse gas is carbon dioxide (CO2), created by fossil- fuel-burning power plants. These gases trap the earth’s heat, causing global warming. Legislation is not yet in place, but any policy changes or legislation will be enacted through the Kyoto Protocol. CFC and HCFC refrigerants used by the HVACR industry are suspected ozone-depleting substances. CFC, HCFC and HFC refrigerants are considered greenhouse gases. In addition, HVACR equipment is a major power consumer. Therefore, the industry is part of these environmental challenges.

1.2 WORLDWIDE OZONE DEPELETION LEGISLATION

Worldwide legislation has been enacted through the United Nations environmental Program (UNEP) to reduce stratospheric ozone depletion. The Montreal Protocol was approved in 1987 to control production of the suspected ozone-depleting substances, among them chlorofluorocarbons (CFCs) and hydro chlorofluorocarbons (HCFCs), commonly used as refrigerants in the HVACR industry. The Montreal Protocol has a provision to conduct and review future scientific, technical and economic assessments, and adjust the legislation accordingly. Further evidence in the early 1990s did suggest a phase out of ozone-depleting substances was in order. Amendments were made to the

14 Protocol in London (1990), in Copenhagen (1992) and in Vienna (1995). No changes have been made since 1995.

Figure 1-1 Ozone distribution over earth’s atmosphere, as on August 1, 1998

CFCs, which have the highest Ozone Depletion Potential (ODP), were phased out of production on January 1, 1996. Today, the only available CFC refrigerants for replacement in existing equipment are those that were stockpiled prior to 1996.

Montreal Protocol initially scheduled HCFCs including HCFC-22 (R22) for total phase out by 2030. HCFCs Production caps begin January 1, 1996, based on 2.8% of CFCs used in 1989, weighted by ozone depletion potential (ODP); plus

ODP-weighted 1989 HCFC consumption, thus on January 1:

15 Table 1–1 Montreal Protocol Production Caps Ozone Depleting Developed Countries Developing Countries Substances Hydro chlorofluorocarbons Freeze: beginning of 1996 Freeze: 2016 at 2015 base (HCFCs) 35% reduction by 2004 level 65% reduction by 2010 90% reduction by 2015 Total phase out by 2040 99.5% reduction by 2020 Total phase out by 2030

Production and import of HCFCs in developed countries are currently capped at the 1989 baseline values. There will be a 35 percent reduction in the cap in 2004, a 65 percent reduction in 2010, a 90 percent reduction in 2015, a 99.5 percent reduction in 2020 and a complete phase-out in 2030

Developing countries have an additional 10 years from the dates for developed countries. There is also a monetary fund established to help them make the transition to alternatives.

The recent Montreal Protocol meeting in Beijing resulted in a proposed amendment to limit the production of HCFCs. It will require developed countries to freeze the production of HCFCs in 2004 at 1989 levels and developing countries to do so in 2016 with a 2015 baseline. Production of 15% above baseline will be permitted to meet the "basic domestic needs" of developing countries. The Beijing amendment will enter into force after it has been ratified by 20 governments.

Figure 1-2 European Union HCFC production phase out schedule

16 The European Union has accelerated the reduction schedule for HCFCs with an eventual phase-out in 2015.

Table 1–2 European Union ban schedule on use of HCFC-22 in the new equipment 1/1/2001 HCFC-22 Ban on using in all new refrigeration equipment with a few exceptions

1/1/2002 HCFC-22 Start of the ban on the new refrigeration and air-conditioning equipment with cooling capacity less than 100 kW

1/1/2004 HCFC-22 Start of the ban on the new Reversible Heat Pumps

1/1/2020 HCFC-22 Start of the ban on the use of virgin HCFC-22 refrigerant for service purposes

1/1/2015 HCFC-22 Total phase out

The United States of America implemented the Montreal Protocol agreement as part of the Clean Air Act Amendments of 1990.

Table 1–3 The United States phase out schedule for the HCFCs

1/1/2003 HCFC-141b Ban on production and consumption

1/1/2010 HCFC-22/142b Freeze on production and consumption Ban on virgin unless used as a feedstock or refrigerant in appliance manufactured prior to 1 January 2010.

1/1/2015 All other HCFCs Freeze on production and consumption. Ban on all other virgin HCFCs used as a feedstock or refrigerant in appliance manufactured prior to 1 January 2020.

1/1/2020 HCFC-22/HCFC-142b Ban on production and consumption

1/1/2030 All other HCFCs Ban on production, consumption

Note: Consumption is defined as Production plus Imports minus Exports.

17 The United States of America adopted much relaxed policy by imposing freeze on portion of HCFCs from 1st January 2010, 14 years later than the Montreal Protocol time table. While, Australia continues to be a world leader in the phase out of ozone depleting substances, in many cases well ahead of the Protocol requirements. Australia's approach has been based on a highly co-operative partnership between industry, the community, and all levels of government

The countries of the European Community have adopted even stricter measures and new equipment banned from using HCFC-22.

After the phase-out dates, supplies of HCFC-22 should still be plentiful thanks to conservation and recycling techniques developed for CFCs. For instance, CFC-11 hasn’t been produced since January 1996, but because of recycling, it is still widely available. Recycled HCFC-22 should be even more plentiful, because it is more widely used than CFC-11. It is important to remember that the phase out affects only production, not use.

1.2.1 Kyoto Protocol

The Kyoto protocol offers new incentives for technological creativity and adoption of kind of solutions that make economic sense irrespective of climate change. Because activities and products with zero or low emissions will gain competitive advantage, the energy, transport, industrial, housing and agricultural sectors will gradually move toward more climate-friendly technologies and practices. The protocol targets 5.2% average greenhouse effect reduction and 6 greenhouse gases in 38 countries with commitment period of 2008-2012 and demonstrated progress to goal by 2005 as well as banking and carry over of reductions.

1.2.2 HCFC-22 (R22) Substitutes

HCFC-22, introduced about 60 years ago, is the world’s most widely used refrigerant. It serves in both residential and commercial applications, from small window units to large water chillers, and everything in between. Its particular combination of efficiency, capacity and pressure has made it a popular choice for equipment designers. Recently, extensive use of HCFC-22 has made it possible to reduce the use of CFC refrigerants, because its Ozone Depletion Potential (ODP) is as much as 95% lower than CFCs. 18 Nevertheless, it does have some ODP, so international law set forth in the Montreal Protocol and its Copenhagen and Vienna amendments have put HCFC-22 on a phase- out schedule. In developed countries, production of HCFC-22 will end no later than the year 2030. In intervening years, production is reduced in a series of specified steps. Detailed phase-out schedules vary from country to country

The evaluation of the Chlorine free substitutes for HCFC-22 indicates that the possibilities of direct, comparable single substance refrigerants are limited. Apart from Ammonia and hydrocarbons, with their special application criteria, as a single substance only the refrigerants HFC-32, HFC-125 and HFC-134a remain as direct potential substances for HCFC-22.This is due to their specific characteristics such as zero ODP, flammability and thermodynamic properties. A refrigerant blend containing these substances could match characteristics of HCFC-22.

Obviously, new substitutes should possess what HCFC-22 lacks from an environmental standpoint, while retaining its good thermodynamic characteristics. Environmentally, the ideal candidate must have an ODP of zero and a low Global Warming Potential (GWP). Thermodynamic similarity is important when retrofitting existing chillers; otherwise, capacity and efficiency may be lost. In new chiller designs, thermodynamic similarity is less of an issue because the equipment can be specifically designed for the replacement refrigerant’s particular characteristics.

Currently, there are four leading replacement candidates: HFC-407C, HFC-404A, HFC- 134a, and HFC-410A.

1.2.3 Properties of HCFC-22 (R22) Substitutes

All four replacement candidates are Huffs; their molecules do not include chlorine, so their ODP is zero. As far as GWP, however, the numbers do vary, and according to Dupont, the GWP ratings are illustrated in Table 1–4.

19 Table 1–4 Characteristics of the HCFC22 and potential substitutes Refrigerants Global ODP* GTD** Critical Composition Warming K Temperature Potential C (GWP)

HFC-134a 1300 0 0 101 Single Substance HFC-407C 1600 0 7.4 87 R32/R125/R134a (Blend) HFC-410A 1890 0 <0.2 72 R32/R125 (Blend) HFC-404A 3750 0 0.7 73 R143a/R125/R134a (Blend HCFC-22 1700 0.05 0 96 Single Substance Notes: (*) Ozone Depletion Potential (ODP) with reference to CFC-11(R-11) = 1.0 (**) Gliding Temperature Difference (GTD), difference between bubble and dew points at 100 kPa absolute pressure. This value varies depending on the pressure.

While there is a variance in these numbers, it is not significant. According to the TEWI (Total Equivalent Warming Impact) formula being proposed by GWP researchers, there are two ways in which refrigeration system can affect global warming: refrigerant leaks and electric-energy consumption. According to refrigerant manufacturers, all HCFC-22 substitutes will be available for the long-term. All are non-flammable and have low toxicities like HCFC-22. However, their capacities, efficiencies, and pressures vary, which affects their application.

1.2.4 Retrofitting Existing systems

To be successful in a retrofit situation, a replacement refrigerant needs to have capacity, efficiency and pressure characteristics comparable to HCFC-22. These thermodynamic properties will ensure satisfactory equipment capacity and efficiency.

20 3500

3000

2500 HCFC-22 HFC-134a 2000 HFC-407C kPa 1500 HFC-404A 1000 HFC-410A

500

0 1 Refrigerant

Figure 1-3 Mean Condensing Pressure comparison of the alternative refrigerants to HCFC-22 at 50 C CT.

HFC-407C is very similar to HCFC-22 in all thermodynamic respects, which makes it a good retrofit candidate. Consequently, HFC-407C is being used for retrofits in some areas now. The capacity of HFC-134a is much lower than HCFC-22. The fact that HFC- 134a requires about 35% more volumetric capacity on the compressor usually results in substantial de-rate, when used as a drop-in replacement for HCFC-22. So it is not a good retrofit candidate. The pressures of HFC-404A and HFC-410A are much higher than HCFC-22, so they are not good retrofit candidates either.

1.2.5 Refrigerant Solutions for Today’s Environmental Challenges

Table 1–4 shows several alternative refrigerants for the systems that use positive displacement compressors. There are enough differences between each of them that one refrigerant would not be the likely choice for all chillers. Each manufacturer will determine which refrigerant makes the most sense for its product line replacement for HCFC-22.

Progress is being made, and new systems using HFC-407C, HFC-410A and HFC-134a are now entering the market to replace HCFC-22. The incremental phase out of HCFCs was a necessary step to make a smooth transition from CFCs to more environmentally friendly refrigerants. The fact that HFC-134a requires about 35% more volumetric capacity on a compressor usually results in substantial de-rate, when used as a drop-in replacement for HCFC-22.

21 HFC-407C is a high-pressure refrigerant, similar to HCFC-22 with ODP of zero. As such, it is a long-term environmental solution to ozone depletion and is not scheduled for production phase-out. HFC-407C is a blend refrigerant with a low toxicity level. In selecting a refrigerant to replace HCFC-22, HFC-407C was found to be an ideal alternative for air-cooled DX chillers. Its operating characteristics are so similar to HCFC-22 that it has been used as a drop-in alternative in machines originally designed for HCFC-22. This allows the continued use of the proven design and components of air-cooled chillers. Additionally, a unique feature of HFC-407C – the property known as “temperature glide” – presents intriguing implications for DX chiller design. This temperature glide phenomenon, when matched with DX cooler design modifications, can improve energy efficiency. HFC-407C is a blend of three refrigerants: HFC-32, HFC-125 and HFC-134a. This composition exhibits the characteristics of a zeotropic blend, meaning that the resulting mixture does not act as a single substance. At a given pressure, it evaporates over a range of temperatures, rather than at a single temperature. Thus, the term, temperature glides.

HFC-407C has a gliding temperature difference (GTD) of approximately 6°K, which can be leveraged to give opportunities for greater operating efficiency. If the DX evaporator is designed as a counter-flow heat exchanger, refrigerant and water enter at opposite ends, and the leaving refrigerant temperature can be greater than the leaving chilled water temperature. The higher leaving refrigerant temperature means the compressor does less work, resulting in lower power consumption.

There has been more focus on R407C as a possible R22 substitute. Preliminary theoretical analyses based on its thermodynamic properties, clearly shows R407C as the best possible drop in substitute for R22. R407C is a zeotropic mixture of R32/R125/R134a (23/25/52% in weight) with a gliding temperature difference in evaporation and condensation of average about 6°K. R407C can be considered as a long term substitute of R22 because it is an HFC.

HFC-410A is a high-pressure refrigerant, having approximately 50% higher pressure than HCFC-22. As a result, it can provide significant capacity gain to a compressor designed to handle the pressure. HFC-410A is a leading candidate for unitary residential and commercial equipment. HFC-410A is a blend refrigerant consisting of HFC-125 and HFC-32 (50/50%). It has a low temperature glide of 0.5 °C, small enough to have

22 little or no effect on heat exchanger performance. HFC-410A has an ODP of zero. As such, it is a long-term environmental solution to ozone depletion and is not scheduled for production phase-out. However, due to its high pressure, the existing systems need to be redesigned, manufactured and certified for the new pressure level. This in turn poses enormous challenges in terms of engineering and manufacturing costs while try to stay competitive in a price driven market. Therefore, HFC-410A cannot be considered as a drop-in alternative for the existing R22 systems.

1.2.6 HCFC-22 Phase out and Recycling

International law set forth in the Montreal Protocol and its Copenhagen, Vienna and Kyoto amendments have put HCFC-22 on a phase out schedule. In developed countries, production of HCFC-22 will end no later than the year 2030. In the years leading up to the proposed year, production is reduced in a series of specified steps. Detailed phase out schedule vary from country to country.

In Australia, HCFC-22 is frozen since 1996 and has become a controlled substance. While in the United States, production of this gas will be frozen on at baseline level on January 1, 2010, and the production of virgin refrigerant will be banned unless it is used as a feed stock for other refrigerants, or equipment manufactured prior to this date.

The countries of the European Community have more restricted measures by banning all new equipment with HCFC-22.

In the mean time, during the scheduled cut in production and phase out, supplies of HCFC-22 should still be available due to conservation and recycling techniques developed during the last few years.

In Australia, Industry Codes of Practice is designed to minimize escape of HCFC-22 gas into the atmosphere. It is expected that, due to the cost involved, the recycled HCFC-22 will become far more expensive than the alternative gas like HFC-407C.

1.3 PURPOSE OF THIS RESEARCH

The purpose of this research is to investigate behavior of R407C refrigerant in chiller systems. This includes performance and efficiency variations when it replaces R22 in an existing system as well as challenges involved maintaining the system charged with 23 R407C. Various researches have been conducted on this refrigerant, which were touched base earlier ranging from performance comparisons in evaporator, condenser and system to leak simulations and effect of the leak in more general terms.

However, this research is intended to address challenges faced in the real world and practical terms. First, to evaluate precisely the performance of a given R22 chiller when charged with R407C at different conditions that are relevant to real design and operating conditions. Second, to investigate chiller systems that subjected to substantial gas leaks. These cases are not uncommon in the industry that, due to composition change of the remaining refrigerant, requires replacement of entire gas.

Majority of chillers are large machines with high capacity and are installed in plant rooms of central cooling systems. They contain large amount of refrigerant, which can range from about 50 kg up to 1000 kg. Thus expensive exercise as chillers contain large amount of refrigerant, while, it is desirable to top up the leaked gas to the required amount. In reality, majority of the leaks are slow and in vapor form, therefore, can be considered as isothermal vapor leaking process. It is important to investigate to what extend composition change would affect the performance of the system after the leak as well as after the remaining refrigerant topped up with an original refrigerant.

2 LITERATURE REVIEW Popular interest in the use of refrigerant blends started in the late 1950’s.The emphasis was placed on energy savings through the reduction of irreversibility in the heat exchanger and on capacity variation during operation through the control of the fluid composition. Cooper (1982) performed tests to determine the heating and cooling performance of standard air conditioned unit using the refrigerant mixture of R13B1/R152a to replace R22. An improvement of 27% in COP was obtained for heating at very low outdoor temperatures, while an improvement of only 4% could be achieved for cooling.

Kruse (1998) studied chlorine free azeotropic and zeotropic mixtures to understand concerns about the internal composition shift, differential oil solubility and the leakage. When leakage occurs in a system, the remaining mixture has higher concentration of the higher boiling point component than the original blend. In order to evaluate this effect, ternary zeotropic blend of R32/R125/R134a and azeotropic blend of R143a/R125/R134a were studied. While the separation effect by leakage with the azeotropic mixture is negligible and below 1% until 40% leaked out, a change of composition between 5 to 10% could be noticed with the zeotropic mixture.

Figure 2-1 Various mixture combinations of the Chlorine free R125/R32/R134a gases

Due to the composition shift during evaporation process, Zeotropic mixtures are not suitable for flooded evaporators,. Composition shift of up to 3% have been calculated with dry expansion evaporators using a simulation program. Therefore, more investigations are required on application of zeotropic mixtures in the dry expansion systems.

Differential solubility of the individual component in the oil is another concern with the refrigerant mixtures. This effect applies to both azeotropic and zeotropic refrigerant mixtures depending on their individual behavior with the oil which influenced by temperature and pressure in the system, especially in the oil receiver. Concentration shifts by oil solubility of around 5 to 7% for the low boiling components have been measured in a water chiller charged with ternary mixture R32/R125/R134a.

Tolouee et al (1991) examined blend of R22/R142b as a drop in replacement for R12. A computer simulation program using refrigerant blends had been developed and the predictions compared with the test results. The computer model predicted the experimental values with a maximum deviation of five percent. It became evident that the saturation pressure of a mixture with 60% mass fraction of R22 closely follows that of R12.

26 Gabrielii et al (1998) compared R22 and R407C in terms of their thermodynamic and energetic performances. The experimental tests have been carried out in a vapor- compression refrigerating plant. The final conclusion reached on the basis of the experimental results is that the efficiency of R407C is significantly lower than that pertaining to R22. Therefore, a plant working with R407C requires higher electric power consumption in order to provide the same refrigerating load. Besides the direct costs, environmental impact is the indirect cost, since more fuel must be burned and higher amounts of carbon dioxide discharged in the atmosphere. The efficiency of R407C must be increased substantially prior to any practical application. To achieve this, practical improvements that suggested by a detailed analysis on each device of the plant need to be applied. In most cases the thermodynamic design of the plant is perhaps more focused towards R22 characteristics.

They suggested that the analysis should start from the heat exchangers of the vapor compression system. In the present works attention has been focused on the evaporator. The mean heat transfer coefficient of R407C has been evaluated and compared with R22. By determining these coefficients extra costs due to under design or over design of evaporators, boilers and other two-phase process equipment can be avoided..

Goto et al (2001) studied condensation and evaporation heat transfer of R410A inside internally grooved horizontal tubes. Spiral groove tube of 8.01 mm o.d. and 7.30 mm mean ID, and a herring-born groove tube of 8.00 mm o.d. and 7.24 mm mean i.d. were used in the experiments. Condensation and evaporation heat transfer coefficient, and pressure drop were measured with the system was charged with both R410A and R22. Test results indicated that local heat transfer coefficients of the herring-bone grooved tube are about twice as large as those of spiral one for condensation and are slightly larger than those of spiral one for the evaporation. The measured local pressure drop in both condensation and evaporation is well correlated with the empirical equation proposed by the authors. The conclusion is that the herring-bone grooved tube is more effective in enhancing evaporation and condensation heat transfer than the conventional spiral groove tube. It is possibly that the enhancement of heat transfer during the evaporation and the condensation is made by the mixing of liquid film of refrigerant and increasing the effective heat transfer surface area on heat transfer, where the thin liquid

27 film forms. However, the liquid film flow mechanism inside these tubes is not explained.

The local evaporation and condensation heat transfer coefficient calculated using basic heat transfer correlations and experimental data. Refrigerant side local equilibrium quality x is calculated by means of local heat balance and local enthalpies as:

xin=(hin-hLin)/(hVin-hLin), xout=(hout-hLout)/(hVout-hLout) 2.1

The equation predicted the pressure drop inside a grooved tube within ±20% for both the evaporation and the condensation heat transfer of R22 and R410A.

An experimental apparatus made up of a heat pump refrigerant-loop and two water- loops was used. The refrigerant-loop consisted of a compressor, an oil separator, a four- way valve, a condenser, a mass flow meter, an expansion valve, an evaporator and an accumulator. A constant temperature water tank, a pump, a flow control valve, and a flow meter used on each water loop to supply the heating/cooling water to the condenser and the evaporator, respectively. The test heat exchanger consisted of four 2-m long tube in tube type test sections with the total working length of 8 m. The refrigerant flows inside an inner tube and the water flows between the inner and outer tubes at a counter flow regime.

The accuracy of the measurements were predicted to be within ±0.15 K for ‘T’ and K- type thermocouples, within ±0.05 K for Pt resistance thermometers, within ±0.1% for the refrigerant mass flow meter, within ±0.3% for the water flow meter, within ±3.9 kPa for the absolute pressure transducers and within ±0.3 kPa for the differential pressure transducers, respectively. The accuracy of the measured heat transfer coefficient and frictional pressure drop PF were estimated within about ±40% and ±5%, respectively.

Aprea et al (2000) carried out experimental evaluation on evaporative heat transfer coefficient of R22 and R407C in a vapor compression plant. The mean heat transfer coefficients of a coaxial counter-flow heat evaporator (20 mm ID) in a vapor compression system were measured. A typical small size refrigeration system working conditions used as test parameters with the capacity ranging from 1.9 kW to 9.1 kW and the mass flux varied from 30 to 140 kg/m2 s. The results showed that the R22 heat

28 transfer coefficient is always greater than that of R407C. The over-predicted R407C coefficients are in line with the most of the theoretical predictions by credited literature.

(a) An overview of heat transfer correlations

A large number of correlations have been proposed to predict the heat transfer coefficient in forced convective boiling in horizontal and vertical smooth tube for pure substances and mixture. The prediction of heat transfer coefficient remains essentially empirical due its complex hydrodynamics and heat transfer processes involved. The first general correlation for saturated flow boiling of pure substances is the well-known Chen (1997) correlation. At the start of boiling, both nucleate boiling and liquid convection may become effective heat transfer mechanisms. While the quality is low, the vapor void fraction is relatively low and the nucleate boiling is much stronger than the forced convective effect. Downstream the flow vaporization occurs, the void fraction rapidly increases. Consequently, the flow must accelerate, which tends to enhance the convective transport of heat from the heated wall to the evaporating fluid.

(b) Test apparatus and test conditions

A refrigeration system consisted of a semi-hermetic vapor compression compressor, a coaxial counter-flow condenser connected to a liquid receiver, a coaxial counter-flow evaporator and a manual expansion device. A Coriolis-effect flow meter with accuracy of ±0.2% is fitted at the outlet of the liquid receiver to measure refrigerant flow rate. All the probes are calibrated in the proper operating range in a thermostatic bath equipped by a precision platinum resistance thermometer; the estimated uncertainty is ±0.1 K. Absolute pressure throughout the refrigerant loop is measured by piezoelectric pressure transducers with a range of 0 to 700 or 0 to 3000 kPa and accuracy of ±0.5% of the full scale value. Volumetric water flow rate is measured using a turbine flow meter with an uncertainty of ±0.25%.

A typical working condition for small-scale refrigeration systems is used to determine heat transfer coefficients of the evaporator. The evaporating pressure ranging from 360 to 420 kPa, quality after expansion of 15-20%., superheat 1-2 K, heat flux 1.9-9.11 kW/m2, refrigerant mass flux of 30-140 kg/m2 s and glycol solution mass flow rate: 0.16-0.30 kg/s.

29 Test results showed the R22 heat transfer coefficient is always greater than R407C for any given refrigerant mass flux, while, the difference percentage decreases with increasing refrigerant mass flux.

Experimental heat transfer coefficient for R22 and R407C were compared to the heat transfer coefficient evaluated with the correlations of Shah (1982), Kandlikar (1989) and Chen (1966) with about ±20% deviation over 80% of the experimental points. Similar results obtained with the comparison of the heat transfer coefficients predicted with the correlations of Gungor et al (1986), Yoshida et al (1994). and Kattan et al (1998), with the experimental values.

In conclusion, it is shown that the heat transfer coefficient of R22 is always higher than that of R407C. The difference ranges from 54 to 24% which decreases with increasing the refrigerant mass flux.

The experimental data have been compared with predictions by available correlations. Predictions by these correlations were compared with the experimental data. Theoretical calculations results were under predicted for R22 and over predicted for R407C in most of the cases. The Kandlikar (1989) correlation seems to be the best-fitting correlation for R22 experimental data, while, the Gungor et al (1986) correlations seem to be the best-fitting correlations for R407C.

Johansson et al (2000) have developed a method to estimate the circulated composition in refrigeration and heat pump systems using zeotropic refrigerant mixtures. The method has been tested and evaluated on a well-equipped experimental rig and tests show that it is possible to estimate the composition of the circulated refrigerant mixture to within 2%, by measuring only two temperatures and pressures.

When a system is charged with zeotropic refrigerant mixtures, the circulated composition may change from the nominal to a different composition. The composition may change in the system due to a few reasons such as leakage one phase or accumulation of refrigerant in liquid or vapor phase in certain cycle components. The shifts in composition would affect thermodynamic characteristic of the system by means of capacity, efficiency, pressures, temperatures, etc. It is necessary to know the circulated mixture properties at all stages of the cycle in order to evaluate the system properly.

30 The molar fraction is used in the leakage model for computational reasons. While the nominal mass fraction composition of R407C is R134a=0.52, R32=0.23 and R125=0.25, its nominal mole fraction composition becomes R134a=0.4393, R32=0.3811 and R125=0.1796. A container with the volume V is presumably charged with nz moles of nominal composition of R407C at the temperature T, until a vapor mole quality Q has been obtained. At the equilibrium state ny, i mol of the i-th component is in vapor phase and nx, i mol is in liquid phase. The fraction of the i-th component in the bulk, vapor and liquid is zi, yi and xi, respectively.

Bulk: zR125=1-zR134a-zR32 2.2 Vapor: yR125=1-yR134a-yR32 2.3 Liquid: xR125=1-xR134a-xR32 2.4

Two major leak types of isothermal and adiabatic leakage have been studied. A small leak from a system simulates the isothermal and a fast leak simulates the adiabatic.

The temperature, T, of the refrigerant is kept constant in the isothermal model; the system pressure slowly decreases with vapor leakage, and slowly increases in the liquid leak case. The bulk is slowly enriched with the two more volatile components, R32 and R125, in the liquid leak case, while it is slowly enriched with the less volatile component, R134a.

Both vapor and liquid leakage simulations have been performed at three different temperatures -10, 0 and +10°C.

In the adiabatic leakage case no heat transfer is allowed through the container wall, the temperature and pressure of the system drops rapidly. In other words, the internal energy at any stage is the sum of the internal energy of the liquid and vapor phase before the leak, minus the enthalpy of the leaked out refrigerant.

Refrigerant leakages have been simulated using purpose built vessel for isothermal and adiabatic leak. The composition of the bulk, liquid and vapor changes, when refrigerant vapor or liquid leaks out from a vessel containing two phase equilibrium of a zeotropic refrigerant mixture. The test results shows that as the vapor leak continues, during an isothermal vapor leak, share of R134a in the bulk increases, while, it decreases during an isothermal liquid leak at 10°C, the differences in initial composition between bulk

31 and vapor phase are smaller at higher temperatures, and larger at lower temperatures. It was also observed that the composition change of the vapor phase is much higher than liquid phase in an adiabatic vapor leak.

The author has also simulated two consecutive leaks. A bottle charged with nominal composition of R407C allowed leaking vapor out to another initially evacuated bottle and accumulated there. In the second bottle, the accumulated composition is richer with R32 and R125. When the refrigerant in the second vessel subjected to a new isothermal leak, the accumulated bulk composition follows near vicinity of the same normal leak curve.

It has concluded that compositions actually fit themselves more or less on a curve in any leak process. In the case of R407C, this curve can be explained by ternary function or mole fraction of the first component (R134a) as function of mole fraction of the last one (R32). The latter was curve fitted to the following polynomial function, provided that the original composition was the nominal composition of R407C.

Polynomial correlations also can be generated for the other ternary zeotropic refrigerant mixtures, as they show similar leak behavior to R407C.

Circulated composition in a refrigeration system was estimated using in-situ method by means of iterative solution of multiple equations. So, three equations are needed to find mole fraction of each component in the circulated R407C as it composed of three components, R134a, R32 and R125. The first equation is based on the principle of adiabatic expansion that means enthalpy of the refrigerant is constant before and after the expansion Equation 2.6. The second equation is the conservation of mass theory Equation 2-7, and the third equation is polynomial equation generated earlier i.e. Equation 2.5, with the assumption of composition shift, in case, due to leakage or other reasons. h1s(p1s,t1s,R134a,R32,R125)=h2s(p2s,t2s,R134a,R32,R125) 2.6

R125=1-R134a-R32 2.7

The estimated circulated composition values using this method with absolute deviation of approximately 2% compare to values measured by gas chromatography.

32 Kim et al (1995) simulated leak processes of zeotropic refrigerant mixtures. Two ideal leak scenarios isothermal and adiabatic leak processes were considered in this study. The isothermal leak process represents an ideal case of a very slow leak, and the adiabatic leak process represents an ideal case of a very fast leak from a system or a refrigerant storage bottle. Two refrigerant mixtures R-32/134a (30/70) and R- 32/125/134a (30/10/60) have been selected for calculations.

The modeled system is consisted of a cylinder with a small hole through to let the refrigerant leak through in liquid or vapor forms, as shown in the schematic diagrams (a) and (b) of Figure 2-2.

∆ = ∆ .ynn iv

T, P, V, Z T, P, V, Z

V y n i nV yi

VAPOR VAPOR

LIQUID LIQUID xi xi

ni ni

∆ = ∆ .Xnn ii

(a) Vapor leak (b) Liquid leak

Figure 2-2 Schematic diagram of vapor and liquid leak models zi is defined as the overall mole fraction of the i-th component in the cylinder, which is sum of the mole fractions of the component i in both liquid and vapor phases. While, xi is the mole fraction of the i-th component in the liquid phase and yi is the mole fraction of the i-th component in vapor phase.

33 Initially, the cylinder is assumed full of the saturated liquid of mole fraction xi, thus xi = zi, the overall mole fraction. During the isothermal leak process, Pressure, and both liquid and vapor mole fractions are interactive. Liquid and vapor mole fractions are determined from the equation of state, once temperature and pressure are given.

Once all the liquid is evaporated and the system is full of saturated vapor leakage from this point would result in pressure decrease and no further composition shift.

During isothermal leak of R-32/134a (30/70) and R-32/125/134a (30/10/60) mixtures, vapor and liquid mass fractions of the more volatile component (YR-32 and XR-32) decreases. While, in adiabatic leak case, liquid mass fraction of the more volatile component (R-32) decreases and the vapor mass fraction of this component increases. In fact, a refrigerant charging process is close to the adiabatic leak process, and during this charging process, the mass fraction change for a zeotropic mixture is negligible.

Man-Hoe Kim (2002) evaluated Performance of R-22 and four alternative refrigerants R-134a, R-32/134a(30/70%), R-407C, and R-410A. Testing at operating conditions typical for a residential air conditioner with pure cross-flow condenser and counter-flow evaporator heat pump carried out the experimental evaluation. Counter- flow evaporator and cross flow condenser with water/ethylene and water used as heat transfer fluid.

Test results at 1800 RPM indicated similar capacities for R-22, R-407C and R-32/134a, while the capacity of R-410A is 40% greater than the R-22, and the capacity of R-134a is 32% lower than R22. As for COP, R22 and R407C have almost the same; R32/R134a is 4.7% higher and R134a only 2% improvement compare to R22, while R410A was 7% worse.

However, at 1000 RPM, COP of R410A is 22% higher and R134a much lower than R22 and R407C. The higher RPM resulted in negative effect on the R134a efficiency (COP) due to the higher friction losses and excessive pressure drop across the heat exchanger as a consequence of high specific volume of the refrigerant. The condenser temperature profiles showed steeper temperature slope in zeotropic refrigerant than R22. This is due to the temperature glide associated with zeotropic mixture’s composition shift and influence of pressure drop during a phase change.

34 The temperature profile of R407C showed flatter slope than that of R22 in the evaporator, which indicates that phase change pressure drop offset temperature glide during the evaporation process.

The two zeotropic mixtures R-407C and R32/R134a had a discharge pressure similar to that of R-22, and the R-134a pressure was lower, while R410A had the highest discharge pressure. All three zeoptropic refrigerants R410A, R407C and R32/R134a had mass flow rate of within 10% of the mass flow of R22, but mass flow rate of T134a was much lower.

This research concludes that zeotropic mixtures, R-407C and R-32/134a (30/70), have the closest performance characteristics to R-22, with R-32/134a having a slightly better COP. The major thermodynamic properties such as evaporation and condensing pressures did not deviate significantly from the R-22 values. The binary near-azeotrope, R-410A, displayed a 44% higher capacity than R-22 when tested at the same compressor rpm. However, at a reduced compressor speed at which R-410A capacity matched that of R-22, the COP of R-410A was 22% better than the COP of R-22. It has to be realized that this COP improvement resulted from significantly lower pressure losses especially in the evaporator, suction line, and at compressor valves and from reduced friction losses in the compressor running at a lower speed.

Use of the LLSL-HX, in other words economizer, improved COPs of all fluids tested with exception of R-410A which was not included in these tests. Presence of two phase refrigerant on the suction side did not help to improve COP for R-22 and R-134a. However, for the zeotropic mixtures, R-407C and R-32/134a, a COP improvement was measured even at small superheat values leaving the LLSL-HX. Although the COP increases using LLSL-HX was predicted in theory, but has not been quantified. The author therefore does not recommend any further study on that direction. Effect of oil miscibility has also been mentioned as another factor in the COP increase as using the same oil with different refrigerants would have advantaged some and disadvantaged other refrigerants heat transfer. That is why it is difficult to objectively evaluate the performance potential of different fluids; even if the tests are performed using the same laboratory apparatus. The use of the same oil may penalize the refrigerants those are not miscible with it.

35 3 THEORY

3.1 INTRODUCTION

In a binary solution, the fluid properties ideally show a proportional relationship with the concentration of the mixture. According to Raoult’s Law, partial vapor pressure of each compound in a liquid solution is proportional to the concentration of an ideal mixture. Total pressure will be equal to sum of the partial pressures, when Dalton’s Law is satisfied at the low total pressure, see Figure 3-1 and 3-2. However, the composition of vapor in equilibrium with the liquid will be richer in the more volatile component than the composition in the liquid phase. When the deviation from Raoult’s Law becomes very significant, the total vapor pressure curve may pass through a maximum or minimum, Figures 3-3 and 3-4.

T = Constant

0 20406080100

Figure 3-1 Ideal binary blends at constant temperature

36 0 20406080100

Figure 3-2 Ideal binary blend at constant pressure

3.2 AZEOTROPIC BLEND

In this case, the composition of vapor in equilibrium must be the same as liquid at that specific concentration. This blend is referred as azeotropic composition, which behave just like pure refrigerants. The 500 series azeotropic mixed refrigerants are commonly used in the industry for quite long time. Their vapor pressure is higher than of either component. For example, azeotropic blend of R502 is composed of R22 and R115 and has slightly higher pressure than R22; it is used in low temperature applications instead of R22 where high discharge pressure may have adverse effect on the system.

37 0 20406080100

Figure 3-3 Azeotropic Blend with Minimum Boiling point variety, Temperature vs. concentration

T = Constant

0 20406080100

Figure 3-4 Azeotropic binary blend with minimum boiling point, pressure vs. concentration

38 3.3 ZEOTROPIC BLENDS

Mixtures that their total pressure curve at any concentration does not pass through a maximum value are defined as zeotropic blends. The composition difference between liquid and vapor cause non-isothermal condensation and evaporation. This variation which is obvious on dew point and bubble-point curves is referred as gliding temperature difference (GTD). Ignoring pressure drop effects, pure refrigerants phase change process takes place at a constant saturated temperature corresponding to the condenser or evaporator temperatures. There is no constant condensing and evaporating temperatures with the zeotropic blends, but rather a range of gliding temperature.

Figure 3-5 Zeoptropic blend vapor compression cycle using non-isothermal phase change

A graphic illustration is used to explain thermodynamic cycle of a basic refrigeration system charged with a fixed composition of a zeotropic blend, Figure 3-5. An azeotropic blend behaves like a single substance with theoretically constant evaporation and condensing temperatures at given pressures. Zeotropic blends however experience a temperature glide during the evaporation and condensing cycle, because the phase changes don’t occur isothermally. This is due to the fact that each of the compounds tends to keep its dominant thermodynamic characteristic while acting as a blend. The Figure shows a vapor compression cycle using a zeotropic blend where, evaporating temperature at the inlet t01 is lower than the outlet t02. The difference between these two temperatures is called Gliding Temperature Difference (GTD). A similar phenomenon occurs during condensing, where the condensing temperature at the

39 beginning, saturated vapor line tx1, is higher than at saturated liquid line tx2. Therefore, GTD evap = t02 – t01 and GTD cond = tx1 – tx2.

Vapor-liquid equilibrium properties represented by two upper and lower envelopes at two pressure levels corresponding to the evaporator P1 and the condenser P2.The cycle beginning at the compressor; the zeotropic blend is compressed and hot gas temperature is raised above the corresponding saturated condensing temperature. During condensation process, the total vapor fraction is reduced and the vapor-phase composition shifts to the right along the dew-point line reflecting the ongoing depletion of component A, which is the higher boiling point fluid. Along the bubble point line, concentration of the component A is initially high, in the liquid phase. However, as the process proceeds and the liquid fraction increases, the composition shifts to the right until all of vapor is condensed, where the liquid composition becomes the same as the initial charge. The temperature declines during entire phase-changing process and further cooling results in sub-cooled liquid.

Reverse of the condensation process takes place in the evaporation process, with the more volatile component B evaporating at a higher rate initially

Figure 3-6 Graph illustration of a vapor compression cycle using zeotropic blend with specific composition

40 The non-isothermal evaporation and condensation can be shown on temperature - entropy and pressure-enthalpy diagrams, see Figure 3-6. These represent steady state operation at a fixed composition of the zeotropic blend. If the composition change occurs, then the fluid properties change and would result in a modified diagram.

Ternary blends pressure and temperature behavior can be illustrated in a three dimensional curves, Figure 3-7. Axis AC, BC and AB represent mass fraction of each binary when mass fraction of the third is assumed zero. It can be seen that on the AB axis the binary is Azeotropic, while the other two binaries demonstrate Zeotropic (Non- Azeotropic) behavior. However, the resulting ternary blend would be Zeotropic blend.

Figure 3-7 Ternary blend at constant pressure

41 3.4 THERMODYNAMIC PROPERTIES DETERMINATION

In this section thermodynamic properties of pure refrigerants are theoretically analyzed and extended to determination of binary as well as ternary and high order blend properties. To do this, analyzing Equation of State is the major and the most critical part of this theoretical study.

Depending on suitability, the program uses an Equations of States for a specific family of refrigerants. The Equations of State that used in this prediction software are discussed in this section in details. Pure and mixed refrigerants properties are calculated using relevant Equation of State to get the best accuracy possible. This method makes it flexible in order to get the best and most accurate results.

3.4.1 Redlich-Kwong-Soave (RKS) (1980) Equation of State

For each pure component, it is necessary to have the critical temperature and pressure, molecular weight, ideal gas heat capacity coefficients, and some pressure –temperature data. It is thus possible, using the RKS equation of state and appropriate thermodynamic relations to calculate a variety of thermodynamic properties (enthalpy, entropy and fugacity) of a blend, from a very limited set of experimental pure and mixture data. The original Redlich-Kwong equation of state was proposed in 1949,

RT a P = − ()− bv ()− bvv ) 3.1

Where P is pressure, T is temperature, v is molar volume and, a and b are composition and temperature dependent parameters. Soave (1980) modified this equation by including an expanded ‘a’ parameter for pure components to include a temperature dependent function with two constants m and n. This equation is known as the Redlich- Kwong-Soave (1980) equation of state (RKS).

= α aa cii Thus, for each pure component i, 3.2

Where

42 ⎛ T ⎞⎛ T ⎞ 3.3 α ⎜ −+= ⎟ + ci i ⎜11 ⎟⎜ nm ii ⎟ ⎝ Tci ⎠⎝ T ⎠ 2 2 TR ci aci = 42748.0 3.5 Pci

RTci bb ci == 08664.0 Pci 3.6

The constants m and n have been determined from experimental data for each pure component, using a least square curve fitting technique.

Calculated m and n values for a number of refrigerants are listed in

Table 3–1 m and n constants

Refrigerant m n R11 0.6417 0.179 R12 0.5927 0.1937 R22 0.6737 0.1944 R23 1.15 -0.1156 R114 0.6420 0.2577 R142b 0.7502 0.1601 R152a 0.9331 0.677 R500 0.668 0.1792 R502 -0.1165 0.6696 R13B1 0.6209 0.1684 R13 0.5756 0.1986

Furthermore, to determine mixture properties, Soave’s (1980) method includes an experimentally determined interaction constant kij in the mixing rule for a.

For mixture of components i and j,

2/1 3.7 ij ()1−= ij ()aiajka 3.8 = ∑∑ i axxa ijj ij

43 ⎛ + bb ⎞ = ⎜ ji ⎟ ∑∑ i xxb j ⎜ ⎟ 3.9 ij ⎝ 2 ⎠

Where, xi and x j are the qualities of each component.

The interaction constants kij are determined, for each binary mixture, from experimental bubble point pressures using a least square curve fitting technique.

Calculated kij values for a number of mixtures are listed in table 3-2

Table 3–2 Calculated kij values

Refrigerant mixture kij Average error (%) R22/R12 0.0351 0.5 R152a/R13B1 0.0833 0.26 R12/R114 0.0027 1.1 R13/R12 0.03 1.1 R22/R114 0.060 3. 19 R22/R142b -0.018 3.24

3.4.2 Vapor Density

Edmister (1968) has shown that the R-K-S equation of state can be rewritten as a cubic polynomial of the compressibility factor Z. For certain values of temperature and pressure this is found to have three real roots. The largest root represents the vapor density; the smallest root is the liquid density, whilst the third has no physical significance.

This polynomial takes the form;

23 ( 2 ) ABZBBAZZ =−−−+− 0 Where 3.10

Pv aP bP Z A == ,, B = RT R T 22 RT 3.11

44 Although this equation is satisfactory for calculating the vapor density, it is generally recognized that it does not result the desired accuracy for calculating liquid density.

3.4.3 Liquid Density

The following correlation is recommended by Downing (1974) for the calculation of the liquid density,

3/1 3/2 4/3 ⎛ T ⎞ ⎛ T ⎞ ⎛ T ⎞ ⎛ T ⎞ ρ ⎜ −+= ⎟ ⎜ −+ ⎟ ⎜ −+ ⎟ ⎜ −+ ⎟ BA fff ⎜1 ⎟ C f ⎜1 ⎟ D f ⎜1 ⎟ E f ⎜1 ⎟ ⎝ Tc ⎠ ⎝ Tc ⎠ ⎝ Tc ⎠ ⎝ Tc ⎠ 2/1 ⎛ T ⎞ ⎛ T ⎞ ⎜ −+ ⎟ ⎜ −+ ⎟ F f ⎜1 ⎟ G f ⎜1 ⎟ ⎝ Tc ⎠ ⎝ Tc ⎠ 3.12

Where ρ f is the liquid density (lb/ft3), T and Tc are the temperature and critical temperature (R) respectively.

The constants used in the liquid density equation, for a number of refrigerants, are listed in Table 3–1 m and n constants

3.4.4 Enthalpy

The vapor and liquid enthalpies, H, at the state (P, T) are evaluated using,

T ⎛ ∆H ⎞ ⎛ ∆H ⎞ += − 3.13 ∫ p dTCH ⎜ ⎟ RT ⎜ ⎟ RTD TD ⎝ RT ⎠T ⎝ RT ⎠ D

In which the isothermal change in enthalpy is computed using,

∆H − 'Taa ⎛ Z ⎞ = ln⎜ ⎟ Z −+ 1 3.14 RT bRT ⎝ + BZ ⎠

Where

45 2/1 ()1− ⎛ aakxx ⎞ a ' = ⎜ cjciijji ⎟ ()' + αααα ' 3.15 ∑∑ ⎜ αα ⎟ ijji 2 ⎝ ji ⎠

And

Tn m α ' −−= icii i 2 3.16 T Tci

CP is the mixture ideal gas heat capacity at constant pressure, TD is the reference datum temperature (chosen to be –40 F/C) and the subscript D denotes the datum temperature. The reference enthalpy is at 0.0 BTU/h/lb liquid enthalpy at –40 F/C.

3.4.5 Entropy

The vapor or liquid entropies at the (P, T) are evaluated using,

T C ⎧⎛ ∆S ⎞ ⎛ ∆S ⎞ ⎫ S = ∫ P + RdT ⎨⎜ ⎟ − ⎜ ⎟ ⎬ 3.17 TD T ⎩⎝ R ⎠T ⎝ R ⎠ D ⎭

in which,

∆S ⎛ − BZ ⎞ a′ ⎛ B ⎞ = ⎜ ⎟ + ⎜1ln + ⎟ R ⎝ P ⎠ bR ⎝ Z ⎠ 3.18

For each pure component, at least three experimental vapor pressure data points must be specified. These data points are input to a regression program which determines the characteristic constants m and n for each pure substance. For each binary mixture, at least two experimental pressures versus composition data point must be specified. These data are input to regression program which determines the characteristic interaction constant kij for each pair. The constants m and n for each component and the constant kij for each pair are input to a set of subroutines for computing properties over the entire range of composition.

46 Computation of Enthalpy and Entropy requires pure component ideal-gas heat capacity data in the form of polynomial coefficients. Ideal gas heat capacity equation constants for several refrigerants are given in table 3-4. Coefficients for the liquid density equation must be provided by the user in the data file.

If the user inputs the condenser dew point temperature, it is necessary first to have a preliminary estimate of the pressure prior to starting an iteration to determine the actual pressure. The vapor pressure equation and constants used, for a number of refrigerants, are listed in table 3-3.

3.5 OTHER EQUATIONS OF STATE

The thermodynamic properties can also be calculated using comprehensive equations of state. Using this approach allows calculations at all conditions with thermodynamic consistency. Other methods, such as the combination of a vapor-phase model with vapor pressure and liquid density equations may not be applicable in the compressed liquid and supercritical regions.

3.5.1 Pure Refrigerants

Depending on the availability of data, there are three other models that are widely used for the thermodynamic properties of pure components. However, one of them considered to be of high accuracy for refrigerants other than R123. This high-accuracy pure-fluid equation of state is expressed in terms of reduced molar Helmholtz (1882) free energy:

A id r ti dt kk lk a == a + a = ln ∑ i ++ ∑ k exp(− γδδταταδ ) 3.19 RT i k

Where, the first two terms on the left side of the Equation 3.19, represent the ideal-gas property effect. The second summation is the real-fluid, effect a r . Two dimensionless variables expressing the temperature and density,τ = * TT and = ρρδ * , where the parameters T* and ρ * have reducing effect and are often equal to the critical parameters.

47 The parameters, αi and αk are coefficients obtained from the experimental data, and the exponentstk , ti , and dk are determined by processing a large number of data on a software program. The parameter γ is equal to 0 when lk = 0; [and it is equal to 1 whenlk ≠ 0 . This model is used for the international standard calculation of R134a. Due to its complete description of the thermodynamic properties, this model is sometimes considered as the fundamental equation

3.5.2 Mixed Refrigerants

A new model is developed by Tillner-Roth ( 1997) to calculate the thermodynamic properties of mixtures. This model is based on the Helmholtz (1882) energy that applies mixing rules to the mixture components:

n−1 n id r excess mix = ∑[]()jj j ln xxaaxa jj +++ ∑∑ p aFxx pqpqq 3.20 p=+11= pq

1 ∞ a r ∫ ()−= ρ dVRTP 3.21 RT v

T T id href Sref ⎛ Tρ ⎞ 1 1 C p aid +−−= ln1 ⎜ ⎟ + id dTC − dT 3.22 RT R ⎜ ρ ⎟ RThT p R ∫∫ T ⎝ refref ⎠ Tref Tref

Where the id subscript denominates the ideal gas and the superscript r denominates the real fluid terms for each of the pure fluids in the n component mixture The first summation in Equation. 3.20 represent the ideal solution. The lnxx jj terms are coming from the entropy of ideal gases mixture where x j is the mole fraction of component j. The double summation in the same equation refers to the “excess” deviation from the

excess ideal solution. The parameters and term(s) Fpq and its multiplier a pq are generalizing parameters, relate to the behavior of one binary pair with another, are empirical

r excess functions related to the experimental binary blend data. The a and a pq functions are determined at a reduced temperature and density τ andδ . The Parameters href and Sref are reference enthalpy and entropy at the defined conditions of Tref and ρ ref .

48 * T ρ mix τ = and δ = * with Tmix ρ

n n * 1 crit crit = ∑∑ , ppqT ()pq + TTxxkT q 3.23 p==11q 2

1 n n 1 ⎛ 11 ⎞ = xxk ⎜ + ⎟ 3.24 ρ * ∑∑ , pqv qp ⎜ crit ρρ crit ⎟ p==11q 2 ⎝ p q ⎠

Where, the k , pqT parameter is determined by taking into account the bubble point pressures that shows azeotropic behavior. The volume changes on a binary blend represented by the k , pqv parameter. Ternary and higher order blend grouped into binary pairs with k , pqT =1 and k , pqv = 1for p = q .

excess In the cases where limited vapor-liquid equilibrium data can be obtained, the a pq term can be eliminated. This term can only be determined if extensive single-phase pressure- volume-temperature and heat capacity data is obtainable. A function to determine a

excess a pq has been developed by Lemmon using data for 28 binary pairs of various type of gasses including HFC’s.

The thermodynamic criteria for the critical components for three component or higher order blends, is a complex calculation exercise. Therefore, combination of the binary critical lines used to estimate the critical parameters.

6 crit crit crit k ii jj ++= ∑ k i xxcTxTxT j 3.25 k=1

The values are constants that can be obtained from test results for each combination of binary mixtures.

For three component and higher number of component blends, the following equation is used.

49 n−1 n crit ∑∑ ijji ()ZTxx crit i=+11= hij T = n−1 n 3.26 ∑∑ xx ji i=+11=ij

crit Where, the ij ZT )( is the binary critical temperature for the pair ij)( mixture, based on the mixture composition x , and at a combining-unauthentic-composition Z .

Pure-component and binary-pair mixture boundary conditions or empirical methods used to estimate the critical volume. Given the critical temperature and volume calculated by the method outlined above, the critical pressure is determined using the mixture equation of state.

3.6 CYCLE COMPONENTS

The major components used in the vapor compression refrigeration cycle include the compressor, evaporator, condenser, expansion valve, suction line and discharge line. The steady state conservation and performance equations are similar to those found in many refrigeration and air conditioning texts.

3.6.1 Compressor

A typical reciprocating compression process is used in this modeling. Thermodynamic cycle of a reciprocating compression process is basically similar to all positive displacement compressors including screw and scroll with different volumetric efficiencies. However, the method of suction and compression is different to each of them, more details in Chapter 4.

In the calculation of the compressor power consumption the effects of clearance volume, polytrophic compression, leakage losses, pressure drops across the valves and superheating on the intake stroke, have been included.

The compressor power consumption is given by,

50 ⎡ n−1 ⎤ ⎛ n ⎞ ⎛ p ⎞ n & −= ⎜ ⎟ ⎢⎜ c ⎟ − ⎥ mW & vp bb ⎢⎜ ⎟ 1⎥ 3.27 ⎝ n −1⎠ ⎝ pb ⎠ ⎣⎢ ⎦⎥

In which n is the poly tropic Index, the subscript b and c denote the cylinder suction and discharge conditions respectively, and ,vP bb , and m are evaluated using,

ab 4 ∆−== pppp in

c 5 ∆+= ppp ex

a = TT 4

ab ∆+= TTT cyl PD m& = ηv v4 ⎛ Bre2 ⎞ ××= StrNrpmPD ⎜π ⎟ ⎝ 4 ⎠ And the volumetric efficiency is obtained from,

/1 n ⎡ ⎛ p ⎞ ⎤ v η ⎢ −+= ⎜ c ⎟ ⎥ 4 v 1 cc ⎜ ⎟ 3.33 ⎣⎢ ⎝ pb ⎠ ⎦⎥ vb

For most cases, a clearance volume of 6-7% may be considered as reasonable.

A low overall volumetric efficiency may be caused by leakage losses passed the discharge and suction valves, piston and stuffing boxes. Valve leakage may be due to un satisfactory grinding of the valve or ring to the valve plate.

Hence, the actual volumetric efficiency may be written as follows,

⎛ p ⎞ ηη −= ⎜ cc ⎟ vva vlc⎜ ⎟ 3.34 ⎝ pb ⎠

Where vlc is assigned a value of 0.01, however, for old compressors this value may be higher.

51 The heat transfer rate from the compressor to the surroundings can be determined using,

3.35

& Comp & ()45 +−= WhhmQ &

Figure 3-8 Compression process diagram

3.6.2 Evaporator

The evaporator capacity may be calculated using,

& e & ()−= hhmQ 23 3.36

In which h3 is the enthalpy at evaporator outlet that includes pressure drop in the evaporator as well as superheating, and h2 is the enthalpy at evaporator inlet.

3.6.3 Condenser

In a similar fashion the heat rejected in the condenser is evaluated using,

52 & −= c & ()hhmQ 16 3.37

In which h6 is the enthalpy at condenser inlet (including discharge line losses) and h1 is the condenser outlet enthalpy (including sub cooling and condenser pressure drop).

3.6.4 Expansion Valve

The expansion valve is considered to be adiabatic thus the exit enthalpy is equal to the inlet enthalpy.

3.6.5 Suction and discharge Lines

Heat gain in the suction line accounted for and calculated from,

& ()−= hhmQ sl & 34 3.38

Whilst, the heat loss in the discharge line that connects the compressor outlet to the condenser inlet is obtained using,

& −= dl & ()hhmQ 65 3.39

3.7 Determination of the Polytropic Exponent

In practice the compression process is neither isothermal nor adiabatic, but polytropic. Figure 3-9 illustrates the typical Process in equation 3.40 for various values of the exponent.

n = constPv . 3.40

The value of the exponent n may vary from 0 to 1.4, however for reciprocating compressor this lays within the limits of 1.0 and 1.3.

Exp. = 1 Isothermal

Exp. = n Polytropic Exp. = γ Adiabatic V

Figure 3-9 Various compression processes represented on a P – V diagram

1〈n〈γ C γ = p Cv

The numerical value for the exponent n may be determined using, − loglog PP n = 1 2 3.41 2 − loglog vv 1

In which P1, v1, P2 and v2 are the gas pressure and specific volume after intake valve (before compressor) and before exhaust valve in the compressor cylinder, indicated as the points b and c in Figure 3-8.

3.8 CONSTANTS

The vapor pressure equation which was used is given by,

B ⎛ −TF ⎞ log AP log ++++= EDTTC ⎜ ⎟ log()−TF 3.42 T ⎝ T ⎠

In which P has the units of psia, T in Rankin and the constants are presented in Table 3–3.

54 Table 3–3 Constants for vapor pressure equations

Refrigerant A B C D E F R11 42.147028 4344.3438 -12.8459 0.0040083725 0.0313605 862.07 R12 39.883817 -3436.6322 -2.4715 0.004730442 0 0 R22 29.357544 -3845.1931 -7.86103 0.002190939 -0.4457467 686.1 R23* 328.90853 -7952.7691 -144.514 0.24211502 0.000212806 0.000000094349 768.35 R114 27.071306 -5113.7021 -6.30867 0.006913003 0.7814211 0.0 R500 17.780935 -3422.6971 -3.63691 0.0005.02722 0.4629401 695.57 R502 10.644955 -3671.1538 -0.36983 -0.001746352 0.8161139 654.0

* The form of the equation used is, B log AP log 2 +++++= FTETDTTC 3 3.43 T

⎛ BTU ⎞ Ideal gas heat capacity ⎜ ⎟ is evaluated using the polynomial, ⎝ °− Rlb ⎠ In which the polynomial coefficients for several refrigerants are given in Table 3–4. f dTcTbTaC 32 ++++= P T 2 3.44

Table 3–4 Constants for ideal gas heat capacity equation

Refrigerant a b c d f 0.02691 −4 −7 −11 -380.59 R11 3.16266× 10 -.39982× 10 6.77887× 10 −3 −4 −7 −11 0 R12 9.2196× 10 3.789× 10 -2.7494× 10 7.6582× 10 0.033191 −4 −8 0 303.66 R22 2.66138× 10 -7.6813× 10 0.09306 −6 −7 −10 0 R23 -9.2254× 10 4.76601× 10 -2.9949× 10 0.019075 −4 −7 0 0 R114 3.804× 10 1.8203× 10 0.030552 −4 −7 0 0 R500 3.2694× 10 -1.10771× 10 0.0231755 −4 −7 −11 0 R502 4.19845× 10 -1.59926× 10 2.5093× 10

3.9 RESULTS

Various compositions of R22 and R142b were used to validate calculation results versus test outcomes. Comparison of the computed, using the above equations, and experimental COPs, for the entire range of mass fraction, proved that the predicted results followed the trends with the maximum deviation of about 5%. The pressure ratio, mixture vapor-liquid pressures and GTD versus composition showed that the minimum pressure ratio and maximum GTD occur at a mass fraction of 30%, which coincides with the maximum COP.

It has been demonstrated that the computer model determines steady-state performance characteristics with reasonable accuracy.

The model predicted compressor power consumption and COP for the mixed refrigerant system experimentally investigated with a maximum deviation of 5%.

Further comparisons were performed using a software program calculating thermodynamic properties of the pure and mixed refrigerants in the upcoming chapter 7. Also, a simple refrigeration cycle is used to calculate the system performance. This program uses the relevant Equations of State out lined in section 3.5 for pure and mixed refrigerants. The test results showed that they are in agreement with the calculation results with a good accuracy of somewhere between -1 to +4%.

56 4 CHILLER TECHNOLOGY

4.1 INTRODUCTION

Liquid Chillers are large capacity cooling machines that produce cold water, brine or other secondary coolants for industrial and commercial air-conditioning or refrigeration use. The system usually is factory assembled, shipped and installed as one machine. Liquid chillers are either absorption or vapor compression type. Absorption liquid chillers are not subject of discussion in this work.

A vapor compression liquid chiller consists of basic components vapor-compression compressor with drive, liquid cooler (evaporator), condenser, flow control device, electrical and control. Depending on the application and design, it may also include expansion turbine, economizer and a receiver. Some auxiliary components such as oil cooler, oil separator, oil pump and purge unit may be used.

4.2 SYSTEM FUNDAMENTALS

A liquid chiller system with a single compressor and one refrigeration circuit with a water-cooled condenser are shown in Figure 4-1. Usually chilled water enters at 12 C and leaves at 7 C. Nominal design point condenser water enters a cooling tower at about 35 C and leaves at 30 C. Air-cooled and evaporative cooled condensers are also widely used.

Figure 4-1 A typical water-cooled chiller system schematic with shell & tube condenser and evaporator, and with screw or semi-hermetic reciprocating compressor

Multiple compressors with multiple circuit chillers, specifically dual systems, are widely used in the industry. Multiple chillers are often used to add more flexibility and redundancy.

Depending on the type of condensers used, three main categories of chillers are widely used in the industry: water-cooled, evaporative cooled and air-cooled. Chillers also can be either flooded or Direct Expansion (DX) type, depending on which method of evaporation is utilized. Due to their lower condensing temperature, water-cooled and evaporative cooled chillers are more efficient than the air-cooled, and are widely used in high capacity applications. However, their higher cooling tower maintenance cost and associated legion Ella risk has made unpopular for low capacity applications. Therefore, the air-cooled chillers are more widely used in low capacity end of the market.

Figure 4-2 A typical Two stage centrifugal chiller (water-cooled) with economizer

Due to their high evaporating temperature and lower approach, the chillers with flooded evaporators are more efficient than the DX types. But they are more expensive to manufacture and specifically a lot of complexity and challenges involved with oil management. For the aforementioned reasons, they are economically feasible with the large capacities only. But this is about to change with commercialization of the new revolutionary totally oil free centrifugal compressors.

The novel invention of the modular chillers invented in Australia and installed worldwide, offers many advantages such as flexibility for manufacturer and end user, compactness and redundancy as well as part load efficiency.

The chilled liquid is supplied to the medium cooling heat exchangers such as fan coils. They are used in central comfort cooling as well as industrial process cooling systems. Their capacity ranges from about 70 to up to 7000 kW cooling capacity, but mostly fit into 200 to 1000 kW chiller range.

59 Table 4–1 Chiller Technology alternatives – Vapor Compression cycle Compressor Typical Capacity Range Drop in Refrigerant Alternatives Reciprocating 70 to 500 kW HFC-407C

Screw 150 to 1500 kW HFC-407C HFC-134a Scroll 70 to 300 kW HFC-407C HFC-134a Centrifugal Over 500 kW HFC-134a HCFC-123

4.3 COMPRESSORS

Two main energy sources are available for running chillers: electricity and natural gas -- this includes steam and waste heat.

Electric chillers include reciprocating, screw, scroll and centrifugal compressors. The first three categorized as positive displacement compressor, whereas the last one uses centrifugal force to push the gas.

a) Reciprocating

Reciprocating technology, which has been around since the beginning of commercial refrigeration, is the most noted for its rugged design and ability to be rebuilt in the field, but less so for its energy efficiency. Space requirements are minimal, and noise and vibration levels are medium to high. Their capacity roughly ranges from 70 to 500 kW and their overall efficiency is relatively unspectacular. This type of compressor increases gas pressure by means of displacing pistons.

60 Suction Compression Discharge stroke stroke stroke

Figure 4-3 Compression process of a typical reciprocating compressor

ers lindd 140 kW 6 cycylin s derer ylinlin 4 ccy 70 kW ers lind

Capacity y Capacity 2 cylin

0 kW -7°C -1°C 4°C 10°C

Saturated Suction Temperature (SST)

Figure 4-4 Capacity variation as a function of SST of a typical reciprocating compressor

61 A compressor capacity versus saturated suction temperature (SST) at a constant saturated discharge temperature (SDT), Figure 4-4, reveals that the capacity of the compressor increases as the suction temperature increases. In a given compression cycle, a greater mass flow rate of refrigerant can be compressed at higher suction pressure and therefore higher capacity of the compressor. b) Screw

As one of the rotary family, screw technology recently crossed over from the industrial refrigeration sector to the comfort cooling sector. Like centrifugals, they don't require much space in a mechanical room, but their noise profiles are significantly higher than centrifugals. Their capacity ranges and efficiencies are similar to centrifugals; however they cannot utilize variable speed drives, so they can not be as efficient. Refrigerants that can be used include R-22, R-134a, and most recently R-407C. Single and dual rotor screws are the most common types of screw compressors. They trap gas in hollow spaces of female rotor and compress it by means of male rotor or two sun wheels type pinions squeezing it towards the discharge port. This technology has bees around for decades; however, until recently, machining of the rotor profiles had always been a major challenge.

In the air-conditioning industry, helical-rotary compressors are most commonly used in water chillers ranging from 150 to 1500 kW.

62 female rotor

male rotor

housing

Figure 4-5 A typical dual rotor – male and female - screw compressor assembly

discharge meshing point port

Figure 4-6 A typical dual rotor – male and female - screw compression concept

Figure 4-6 is a graphical illustration of a dual rotor screw compression process. Continued rotation of the meshed rotor lobes drives the trapped refrigerant vapor (to the right), toward the discharge end of the compressor, ahead of the meshing point. This action progressively reduces the volume of the pockets, compressing the refrigerant. Finally, when the pocket of refrigerant reach the discharge port,

63 the compressed vapor is released and the rotors force the remaining refrigerant from the pockets.

c) Scroll

Scroll compressor technology is a major invention which increases efficiency and reliability - and reduces noise levels. Its highly complex scroll profiles of these two components optimize the hermetic technology. Scroll technology is rapidly overtaking the niche of reciprocating chillers in comfort cooling. They provide small size, low noise and vibration, and good efficiency. Available in air-cooled and water cooled configurations, scroll chiller capacity can reach approximately 300 kW, which makes them good candidates for spot cooling or make-up cooling applications. Scroll technology is pretty new compare to reciprocating and screw. The recent advances in CNC and precise machining paved the way for cost effective scroll manufacturing. The compression process in the Scroll compressor is of course based on the Scroll technology which is described below.

intake phase

compression phase

discharge phase

Figure 4-7 Suction, trap, compression and discharge process of the scroll concept

64 As shown in Figure 4-7, Scroll compression concept consists of two main components, involutes scrolls that intermesh. The top scroll which contains the gas discharge port is fixed and the bottom scroll orbits. This principle does not require either suction or discharge valves and the absence of clearance volume makes it a very high volumetric efficiency machine. The two scrolls are maintained with a fixed angular phase relation (180º) by an anti-rotation device. As the bottom scroll orbits within the fixed one crescent shape gas pockets are formed, their volumes are reduced until they concentrate at the center section of the scroll. All Suction, compression and discharge processes are performed simultaneously.

journaljournal motor bearing shaft

direction of rotation

driven scroll

Figure 4-8 Rotation geometry of motor shaft, journal bearing and driven scroll

As the orbiting motion continues, Figure 4-8, the relative movement between the orbiting scroll and the stationary scroll causes the pockets to move toward the discharge port at the center of the assembly, gradually decreasing the refrigerant volume and increasing the pressure. Two or three orbits are required to

65 accomplish the volume reduction or compression process allowing very smooth operation.

nsstt Con TT33 SD nst Con TT22 SSDD nst Coon T1T1 Capacity Capacity SDSD

Saturated Suction Temperature (SST)

Figure 4-9 Capacity variation of positive displacement compressors versus SST at three constant SDTs

All the abovementioned three types of compressors, namely reciprocating, screw and scroll fall into positive displacement category. Figure 4-9 illustrates capacity variation of positive displacement compressors as function of SST and constant SDTs, with SDT1 being higher than SDT2 and SDT3 respectively. Capacity increases as SST increase, however, it decreases with discharge pressure rise.

d) Single speed centrifugals

Electric centrifugals have moderate space requirements and noise levels. They range in capacity from 500 kW to 7000 kW (150 to 2000 Tons) for standard production units. This technology has the most flexibility in refrigerant usage including, R-134a, and R-123.

66 volute

diffuser passages

radial impellerimpeller passages

blades impellerimpeller

Figure 4-10 A typical single stage centrifugal compressor cross section

The inlet of impeller is fitted with blades that draw refrigerant vapor into radial passages that are internal to the impeller body. The rotation of the impeller causes the refrigerant vapor to accelerate within these passages, increasing its velocity and kinetic energy, Figure 4-10.

67 refrigerant enters diffuser Velocity or Kinetic Pressure Velocity or Kinetic Pressure refrigerant refrigerant enters volute enters impeller Static pressure Static pressure

Path through compression process

Figure 4-11 Velocity increase by impeller and conversion of velocity into static pressure in a centrifugal compression process

The accelerated refrigerant vapor leaves the impeller and enters the diffuser passages. As size of the diffuser passage increases, the velocity and therefore the kinetic energy of the refrigerant decreases. Thus, the refrigerant’s kinetic energy (velocity) is converted to static energy, or static pressure. The volute also becomes larger as the refrigerant travels through it. As the size of the volute increases, the kinetic energy is converted to static pressure, Figure 4-11.

68 A e rg 90 su B 75 C 75

63 51 63 36 10 14 25 vane position pressure difference pressure pressure difference (degrees) unloading line

capacity

Figure 4-12 Performance of a typical centrifugal compressor over a range of inlet vane positions

Performance of a typical centrifugal compressor over a range of inlet guide vane positions is shown in Figure 4-12. The pressure difference between the compressor suction and discharge is on the vertical axis and compressor capacity is on the horizontal axis. Most of the cases pressure ratio is used instead of pressure differential; however, it does not make difference on the graph trends. As the load on the compressor decreases from the full-load operating point A, the inlet guide vanes partially close, reducing the flow rate of refrigerant vapor and balancing the compressor capacity with the new load B. At the new part load, required heat rejection reduced, hence reducing approach and consequently condensing pressure. This results in reduced the pressure difference between the evaporator and the condenser. Continuing along the unloading line, the compressor remains within its stable operating range until it reaches C.

Surge line (dashed) is where compressor can not operate beyond it due to surge phenomenon caused by less than minimum required gas velocity to overcome discharge pressure. If compressor forced to operate beyond surge line, it would become very unstable and potentially cause damage. Inlet vanes on a centrifugal

69 compressor allow it to unload over a broad capacity range while preventing the compressor from operating in the surge region.

e) The new totally oil-free centrifugal

It is fair to say that, for the last few decades, the compressor technology has been very slow in adapting innovations and has been stuck with single speed motors and oil dependant lubricating system. This is about to change by invention of the revolutionary world’s first totally oil free, variable speed and high efficiency centrifugal compressor, which was developed in Australia and is commercialized in North America, Europe and Australia, Figure 4-13.

2 stage centrifugal Inverter speed compressor control

Pressure and temperature sensors

Synchronous brushless DC motor Inlet Guide Vanes

Motor and bearing control

Figure 4-13 A cut-away section view of the new totally oil-free compressor

Dual stage centrifugal compression is used to maximize full load efficiency, as centrifugals provide the industry’s highest efficiency. In addition, when coupled with a variable speed drive, they provide the highest part load efficiency. It has only one moving part, a shaft fitted with two stage impellers, Figures 4-13 and 4-14.

Figure 4-14 Cross section view of two stage centrifugal gas compression assembly

The new totally oil free compressor uses the latest magnetic bearing technology to levitate the shaft, only moving part, by a digitally controlled magnetic bearing system consisting of two radial and one axial bearing, Figure 4-15 and 4-16. Position sensors provide real-time repositioning of the rotor over one hindered thousand times a second. The levitated shaft rotates floated on air without friction at high speed up to 48,000 RPM.

Figure 4-15 Shaft and impellers assembled with magnetic bearings

Figure 4-16 Cross section schematic of the front radial, rear radial and axial bearings with sensor rings

It also utilizes advantages of variable speed drive as integral part of the compressor to modulate the capacity by adjusting speed to the load, hence achieving unprecedented efficiencies that never seen before.

Being oil-free offers enormous advantages in sustainable performance over the entire lifetime of the application. It simplifies design of the systems and cuts costs and complexity associated with oil management as well as the failure factor due to lack oil in the system. By eliminating oil from the system, improves heat transfer coefficient significantly, as oil coats tubes surfaces and degrades its conductivity.

Figure 4-17 Operating map of the new totally oil free and variable speed centrifugal compressor

Physics of centrifugal compression theory is fundamentally different from positive displacement. Positive displacement method is simply to compress gas by means of reducing volume hence increasing gas pressure, while centrifugal method converts kinetic energy that created by centrifugal force into high pressure. It therefore is sometimes called dynamic compression that can vary depending on flow rate, and low and high pressure changes. Each compressor map plays an important role on operating range and limits of that compressor. We briefly discussed about single speed centrifugal map earlier in this chapter.

However, as it can be seen on Figure 4-17, the new compressor has multiple numbers of map curves, each for a specific constant speed. In fact, due to infinite speed steps, numbers of curves exist in reality, but a few specific speeds selected for better understanding of the map. As it can be seen on the map, Figure 4-17, each curve represent and specific constant speed that follow pattern of upward to a maximum or

73 turning point and continues downward. Turning point is where inlet guide vanes (IGV) starts to close by modulating vanes. From that point to the right all the way to the end tail of the curve is where IGV is fully open and is efficient operating range until the tail end to the right where compressor reaches its maximum flow rate limit. This is also called chock limit which can be determined using Mach number of gas at inlet conditions. However, from that point on to the left is where operation enters surge area and IGV is partly closed depending on gas flow rate and pressure ratio. The end tail to the left of the curve is the minimum load limit and real surge point.

The compressor full map is stored in the compressor controller memory which uses it to know the position at each operating condition and therefore try to optimize as well as reposition it from the limiting boundaries. Due to this and other reasons, this compressor is also called the first intelligent compressor ever in the industry.

74 5 EXPERIMENTAL APPARATUS

5.1 THE TEST APPARATUS

The test apparatus used in this work contains the basic components found in any refrigeration system. The system is basically a single stage vapor compression refrigeration cycle with the nominal capacity of about 3.0 kW. The other major parts are air cooled condenser with air flow control, specially designed and built counter flow tube in tube evaporator and thermostatic expansion valve. Special attention were paid to the evaporator design to resemble small scale chiller as well as a near ideal counter flow heat exchanger. The evaporator was also fitted with number of thermocouples and pressure sensors, so the boiling and superheating process could be better viewed and analyzed. All the sensors then connected to corresponding special data loggers. Various components of the plant are briefly discussed below.

5.1.1 Compressor

This is a hermetic type reciprocating compressor suitable for R22 and R407C with pressure and temperature transducers are fitted at the discharge and suction ports. The compressor is a small swept volume of positive displacement type which is widely used in the industry. Performance and behavior of this compressor would resemble characteristics of the large capacity size compressors currently used in the chiller industry. Overwhelming majority of the compressors that are used in the chillers manufacturing is of positive displacement type. When a volume of gas sucked into a chamber and physically compressed to higher pressure by means of reducing volume is called positive displacement compression. Unlike centrifugal compression, positive displacement compression requires high torque and consequently high starting current draw. In some cases, especially residential air conditioning system, soft starters are fitted to reduce the starting current. The starting current sometimes also called locked rotor amps. Nevertheless, this was not an affecting factor in the intended tests, as the performance data were collected at stable steady operating conditions.

5.1.2 Condenser

The condenser used in this rig is of air cooled coil type to reject heat by transferring heat from inside tubes where condensation phase change is taking place to blowing

75 through air on the coil fins. Copper tubes expanded to aluminum fins, air cooled coil with below through axial fan. Majority of heat exchange takes place from the aluminum fins to air. To enhance the performance, internally grooved tubes were fitted which improves heat transfer by creating turbulence inside tubes. Air inlet adjustments and heaters were used to maintain the desired conditions. The condensation phase change process has always been predicted with reasonable accuracy using heat transfer correlations. The heat transfer research works have proved that internal grooved tube significantly improves condensation efficiency. For this reason, almost all of the compact and high performance condensers equipped with internally grooved tubes. Various fin shapes are proven to enhance heat transfer on the air side by introducing turbulence on the air flow stream.

5.1.3 Evaporator In the design of this test plan, special attention has been given to the evaporator. The evaporator is a tube in tube, dual pass and counter flow U shaped heat exchanger. The intension was to allow the refrigerant effect to occur within the inner tube alone, be able to observe and measure evaporation and superheating behavior all along the evaporation process. The inner tube was of special type internally rifled and externally finned with nominal 3/4 inch diameter. The internally rifled tubes are specifically designed and manufactured to improve boiling phase change. Unlike condensation phase change, evaporation process is extremely challenging to predict with reasonable accuracy. Extensive research work has been carried out on this topic to model various types of evaporators. Each and any type has its unique behavior represented in complex correlations with defined boundary conditions. Of all those, boiling inside a tube with heat transfer fluid at outer diameter have attracted good research work with reasonable modeling correlations. One example of those works describes the relationship between the heat transfer coefficient of pure substance and different flow regimes as a function of the vapor quality in Figure 5-1. Other studies of R407C boiling inside a tube showed similar supporting evidence that the maximum heat transfer coefficient obtained at 65- 80% vapor quality and then drops dramatically characterizing the “dry out” region. The only noticeable difference is that the annular flow is achieved earlier in pure substance than the mixtures. This can be explained due to the effect of composition on nucleation boiling and a change in physical properties of mixture with the composition shifts from boiling

Figure 5-1 Heat transfer coefficient as function of boiling regimes and quality inside a tube, Cengel 1998

The outer jacket tube diameter was 1 5/8” inch with total length of 2.4 m at two parallel lines. The refrigerant flows inside the inner diameter and the chilled liquid flows counter currently in the annular space which is driven by a pump, Figure 5-1. The outer jacket was properly insulated to stop heat transfer between the evaporator and ambient.

A liquid thermal reservoir tank fitted with adjustable electric heaters was used to maintain desired entering chilled liquid temperature. Baffles also installed to properly mixing the water and providing with more stable bulk temperature. Small pump is used to circulate the liquid. The chilled liquid was mixture of water and glycol, which referred to in the industry as brine. Calibrated chilled liquid side flow meters were also used.

COMPRESSOR

PUMP

Glycol EVAPORATOR Solution HEATER Reservoir CONDENSER

TX VALVE

Refrigerant Line

Figure 5-2 Schematic of the experimental apparatus

Temperature Sensors Chilled Glycol IN

To Compressor Suction

TX Valve Refrigerant Liquid From Condenser

Chilled Glycol OUT

Figure 5-3 Evaporator piping and sensor arrangement isometric schematic

The evaporator was fitted with series of temperature sensors as shown on Figures 5-2 and 5-3. Total of 22 temperature sensors used, 16 for refrigerant side which the tip of the sensors were installed at the centre of the inner tube and the rest used for measuring chilled liquid, suction, discharge and liquid temperature measurements, Figures 5-4 and 5-5. These sensors provided very useful detailed behavior information, specifically along the evaporator. They make it possible to plot and view refrigerant temperature change trend during boiling and superheating along the evaporator. Pure refrigerants theoretically boil at constant temperature at constant pressure and rise during superheating process. It is very important to view this in practice where pressure drop is

78 more than zero and rate of heat transfer is not necessarily uniform. For the mixed refrigerants, the boiling process is not isothermal. Instead, the evaporating temperature glides during boiling along the evaporator. This phenomenon is called Gliding Temperature Difference (GTD). The temperature sensor set up would provide opportunity to log and view this process therefore more detailed insight of the evaporation process of pure and mixed refrigerants as well as the leaked system.

14 12 13 com

20 pres 22 sor

5 6 7

16 condenser 18 unit

17 11 10 9 8

4 3 2 1 15 19

Glycol Solution Reservoir pump

Figure 5-4 Temperature sensor (Thermocouple) location schematic

Figure 5-5 Evaporator Thermocouple tip installation schematic

79 For convenience and accuracy sake, the pressure sensors were connected to flare end of the capillary tubes extended from the position where the pressures were intended to be measured. Due to lack of flow and velocity inside the capillary tubes, the measured pressures represent exact pressures on the evaporator.

The thermocouples had to be built specifically for this purpose by the University technical staff. Special attention was paid to tip, the sensing point, of the thermocouples to make sure that bulk temperature is measured. This part was a challenging part of building the evaporator, see Figure 5-4, as they had to be extended past through the outer jacket all the way to the middle of evaporator cross section. Measuring the temperatures on the tube wall would not be considered accurate enough for research work due to temperature gradients from middle of the tube and outside temperature effect.

Figure 5-6 General view of the test rig

Figure 5-7 Data acquisition and instrumentation set up of the experimental apparatus

5.2 INSTRUMENTATION AND DATA LOGGING

5.2.1 Thermocouples

The evaporator was fitted with sixteen K type thermocouples, all along the evaporator, to measure refrigerant temperatures. These were penetrated all the way from outer jacket through to the middle of the inner tube. Another 4 thermocouples were fitted to measure chilled water side temperatures, and additional sensors were fitted for 81 measuring discharge, liquid and the suction gas temperatures. These thermocouples had to be fitted to specially built fittings to seal from outside as well as the water ingress. Unlike pressure sensors, temperature sensors have to be fitted precisely at exact position for accurate and reliable measurement. The temperature sensors in this test rig were positioned on the very spot of the intended measurement. Tip of the thermocouples were inserted from top side of the tube into precisely middle of the evaporator tube cross section.

5.2.2 Pressure Transducers

One low and one high range pressure transducers were fitted on the suction and discharge side of the compressor, respectively. Additional two were used to alternate between a few points. All were initially calibrated using a certified reference pressure gauge. Measuring three pressures on the evaporator side made it possible to calculate pressure drop and relate it to temperature change behavior of the evaporator. For accurate and apple to apple comparison of performances at various conditions, average evaporating temperature corresponding to average pressure was used.

5.2.3 Data logging and calibration of the sensors

All the thermocouples and pressure transducers were connected to a data logger. Due to different signal generated by each category sensors, special data logger was used for each purpose, one for thermocouples and another for the pressure transducers. All signals and data were relayed back from there to a computer. All the data then were logged into the computer transferable to excel file which then used as raw data and then processed.

Before starting any test and data logging, all the pressure transducers were calibrated using certified accurate high resolution pressure gauge against the values read from the computer. Discharge pressure transducer was calibrated to high pressure range, while the other three evaporator side transducers were calibrated to lower pressure range. Calibration graphs were prepared for each one and used in correcting the measured data.

The thermocouples were also calibrated using certified accurate and high resolution thermocouple as reference.

82 Current were measured using current transducers by measuring inductance around the cable to accurately determine the current passing through. The voltages were also measured simultaneously that correspond to the measured current to take into account minor voltage fluctuations. To excel accuracy, currents and voltages were measured over a period of time and average taken to calculate power consumption.

5.3 TEST PROCEDURE

After the test rig set up was complete, the rig was thoroughly checked and commissioned. All the major components, piping, electrical, control and data logging were precisely inspected and tested for system function. The system then subjected to series of tests at various conditions, fine tuned charge, evaporator water circulation and flow meter tuned and expansion valve adjusted. Sample data were taken by the computer and format was set for better reading and test identification. Once all these initial set up, inspection, adjustment, tuning, etc completed, the research work tests started as per the following plan.

At the first stage, series of tests have been carried out using R22 refrigerant under various conditions. At the second stage the system was charged with R407C refrigerant. Series of comprehensive tests have been carried out at the identical various conditions. Details of the various test conditions are outlined in Table 5-1. Before logging and recording data for each test at the specific condition, utmost attention and effort were made to stabilize the system at the intended conditions and let it run at this condition for a period of about 20 minutes. Data would only be recorded if the system proved stable for this period of time.

First, the system was charged with pure R22, tested at the intended various indicated conditions, and logged and recorded the data. Tests were carried out the same way on pure R407C refrigerant by following the same procedures. The tests with pure R22 were carried out so they can be used for comparison purposes. The reason being, R407C is considered as the most suitable drop in alternative for R22. This stage of tests were carried out without a major hick up, except had to make sure the conditions that the system kept were identical. For each condensing temperature, leaving chilled water and sub cooling temperatures as well as the chilled water flow rate had to be kept at certain value. This was very important to resemble real world chiller behavior.

83 The second phase of the experiments was the most important, time consuming and the most challenging part of this work. The system was tested using several combination of leaked gas to simulate, as close as possible, what would happen to chiller equipment if subjected to gas leak. The phenomenon were analyzed from the wide variety of prospective such as effect on capacity, suction and discharge pressures and temperatures, power input, evaporation and condensation, etc. This time the system was tested with a refrigerant blend which its composition was changed due to leak and was not any more a pure defined refrigerant. A fixed percentage of the gas was let to leak out, in vapor form, and measured by collecting it in a bottle. The most volatile substance of the mixture tends to leak during an isothermal vapor leak. The remaining refrigerant becomes more concentrated in less volatile and low pressure substances. The situation gets worse as more percentages of the refrigerant are leaked. This would have consequences in two fronts. Capacity and efficiency drop due to shortage of gas in the system, and refrigerant characteristics change due to high concentration of the less volatile substances in the composition. In academic world, these effects could be interesting from system behavior stand point, but it means much more in the industry. Besides what interests the academic work, performance and efficiency loss, it poses system failure risk due to chilled water freeze up, lack of oil return, high discharge gas temperature, etc.

The next step of the experiments conducted in a way that simulated what the industry has been doing for many years by simply topping it up the system with pure R407C. This is a desired way by the industry; however, this can be applied when using pure refrigerants, not zeotropic mixture like R407C. The system subjected to isothermal vapor leak, at various proportions, and then compensated by adding up R407C refrigerant in liquid form. Special attention was paid to the leak process to ensure that the process was carried out as per the definition and identical at other conditions. It has been established earlier that the vast majority of leaks occur in the real world systems are of isothermal type and very rarely adiabatic. So it is very important to apply precisely the isothermal leak which means slow gas leak. The leaked gas then directed to specially prepared gas cylinder on a scale to weigh the amount using certified weights. Specially fitted valves were used to simulate slow leak process. The idea was to observe and record system behavior, at each proportion, from boiling characteristics to pressures, temperatures, capacity variations and most important of all efficiencies.

84 From there, to find out the acceptable leak proportion and top up level. This is a very important side of this research, as the solution as a result of this work has a potential to save enormous amount of time and money as well as inconveniences for the industry, specially the central equipment owner and maintenance staff.

Table 5–1 Details of the conditions of series of the tests conducted upon. Condensing Condensing Temperature Temperature 40 C 50 C Chilled Water Leaving 7 10 12 7 10 12 Temperature (CHWL) C Pure R22 * * * * * *

Pure R407C * * * * * *

R407C Leaked 10% Vapor and topped up * * * * * *

R407C Leaked 20% vapor and topped up * * * * * *

R407C Leaked 30% Vapor and topped up * * * * * *

R407C Leaked 40% Vapor and topped up * * * * * *

R407C Leaked 50% Vapor and topped up * * * * * *

In order to have an apple-to-apple comparison, it is utmost important to set correct test parameters at various conditions, especially evaporator. A small variation on evaporator conditions would result in considerable effect on the performance of the system, while, condenser side has less effect. Before each test stage, the system was properly evacuated, charged with the required refrigerant, measurement devices and data logger connected to be outlined later.

85 6 TEST RESULTS

6.1 ANALYSIS OF THE COLLECTED DATA

The data that collected through the data acquisition system included temperatures, pressures, flow rates, voltages and current. Upon stabilization of the system, series of data acquired every 10 seconds for the duration of ranging from two to five minutes. Samples of data series has been plotted to show stability of the system during the data collection for each test. Average of each series of data has been used for calculations.

Data Aquisition,40 C CT, 10 C CHWLT 500

480

460

440

Pressure (kPa) Suction Pressure (SP)

420 Average Evaporating Pressure (Avg EP) 400 12345678910 Logged every 10 seconds intervals

Figure 6-1 Evaporator pressure data acquisition for pure R22

Figure 6-1 shows very stable and constant pressures after the 3rd interval with maximum deviation of only about 2 kPa. This represents less than 0.5% fluctuations during the data logging period, which is better than expected for a thermal system. Due to the multiple affecting factors involved, it is a challenge to keep the instability at minimal level. However, even this negligible variation is absorbed by taking average and using them in the data processing.

86 Data aquisition, 40 C CT, 10 CHWLT

10 Temp 12 8 Temp 13 6 Temp 14 4 Temp 15

Temperature C 2 Temp 16 0 12345678910 Logged every 10 seconds intervals

Figure 6-2 Evaporator temperature data acquisition vs. time for pure R22

Data aquisition, 50 C CT, 10 C CHWLT

10 9 Temp1 8 Temp1 7 Temp2 6 Temp3 5 Temp4 4 Temp5 Temp6 3 Temperature C Temperature Temp7 2 Temp8 1 Temp9 0 Temp10 1 3 5 7 9 1113151719 Temp11 Logged every 10 seconds intervals

Figure 6-3 Evaporator temperature data acquisition for pure R407C

87 Data Acquisition, 50 C CT, 7 C CHWLT 7 6 Temp 1 5 Temp 2 4 Temp 3 3 Temp 4 2 Temp 5 Temperature C Temp 6 1 Temp 7 0 1234567891011 Logged Every 10 seconds intervals

Figure 6-4 Evaporator temperature data acquisition for 10% leaked R407C

Data aquisition, 40 C CT, 12 C CHWLT

10 Temp 1 Temp 2 8 Temp 3 6 Temp 4 Temp 5 4

Temperature C Temp 6 2 Temp 7

0 1234567891011121314 Logged data every 10 seconds

Figure 6-5 Evaporator temperature data acquisition for 30% leaked R407C

Figure 6-5 shows the maximum deviation of only about 0.2 C on all of the graphs

With exception of two other graphs represent very stable and constant temperatures after the 2nd interval. The other two graphs show about 0.4 C during the last four readings. This variation is further minimized by using average value of those data. Therefore, the accuracy was not compromised even with the interactive thermal system.

Data aquisition, 50 C CT, 7 C CHWLT

500

480

460 Avg Ep SP 440 Pressure kPa

420

400 123456789101112 Logged every 10 seconds intervals

Figure 6-6 Pressure data acquisition for pure R407C

Data aquisitio, 50 C CT, 7 C CHWLT

2000

1950

1900 CP Pressure kPa

1850

1800 123456789101112 Logged every 10 seconds intervals

Figure 6-7 Condensing pressure data acquisition for pure R407C

89 Data aquisition, 50 C CT, 7 C CHWLT

2000

1800

1600

1400 SP 1200 Avg EP CP 1000 Pressure kPa

800

600

400 1 2 3 4 5 6 7 8 9 10111213141516 Logged every 10 seconds

Figure 6-8 Condensing, average evaporating and suction pressure data acquisition for 10% leaked R407C

Data aquisition, 50 C CT, 7 C CHWLT

450

440

430 SP Avg EP 420 Pressure kPa

410

400 12345678910111213141516 Logged every 10 seconds intervals

Figure 6-9 Average evaporating and suction pressure data acquisition for 10% leaked R407C

Data aquisition,40 C CT, 12 C CHWLT

500

480

460 Avg EP 440 SP

Pressure kPa 420

400 123456789101112 Logged every 10 seconds

Figure 6-10 Average evaporating and suction pressure data acquisition for 30% leaked R407C

Data aquisition, 40 C CT, 12 C CHWLT

1500

1480

1460 CP 1440

Pressure kPa 1420

1400 12345678910111213 Logged every 10 seconds

Figure 6-11 Condensing pressure data acquisition for 30% leaked R407C

91 Data aquisition, 50 C CT, 7 C CHWLT

450

440

430 SP Avg EP 420 Pressure kPa

410

400 123456789101112 Logged every 10 seconds

Figure 6-12 Average evaporating and suction pressure data acquisition for 40% leaked R407C

Data aquisition, 50 C CT, 7 C CHWLT

2000

1900

1800 CP Pressure kPa

1700

1600 123456789101112 Logged every 10 seconds

Figure 6-13 Condensing pressure data acquisition for pure 30% leaked R407C

92 Refrigerant tempreature along the evaporator, at 40 C CT, R22 8

4 CHWLT 7 C Temperature (C) Temperature CHWLT 10 C CHWLT 12 C 2 0246810 Sensor position along the evaporator

Figure 6-14 Refrigerant temperature along the evaporator at CT 40 C for pure R22

Refrigerant tempreature along the evaporator at 50 C CT, R22 8

4 CHWLT 7 C Temperature (C) CHWLT 10 C CHWLT 12 C 2 0246810 Sensor position along the evaporator

Figure 6-15 Refrigerant temperature along the evaporator for pure R22

93 SP & avg EPs , R22

510 500 490 480 470 460 Pressure (kPa) SP @ 40 C CT 450 Avg EP @ 40 C CT 440 SP @ 50 C CT Avg Ep @ 50 C CT 430 7.0 8.0 9.0 10.0 11.0 12.0 CHWLT C

Figure 6-16 Suction and average evaporating pressures vs. CHWLT at CT 40 and 50 C for R22

94 SP & Avg EP, R407C 10% leaked

470

450

430

SP @ 40 C CT Avg EP @40 C CT kPa410 SP @ 50 C CT Avg EP @ 50 C CT 390

370

350 789 10 11 12 CHWLT(C)

Figure 6-17 Suction and average evaporating pressure vs. CHWLT at 40 C and 50 C CT for 10% leaked R407C

Figure 6-17 shows suction and average evaporating pressures trend and pressure drops for R407C after 10% leaked and then topped up. At 40 C CT, the pressure drop across the evaporator appears to be slightly high towards lower CHWLT of 7 C, about 25 kPa. The lower evaporating as a consequence of the lower CHWLT resulted in lower refrigerant velocity and partly oil clogging inside the evaporator tube. Although an average 10 kPa pressure drop is not unusual in direct expansion evaporators, the 25 kPa pressure drop does not pose a serious concern. However, this should be treated as warning sign which could be exaggerated with higher leak ratios.

95 SP & Avg EP, R407C 30% leaked

480

460

440 kPa SP @ 40 C CT Avg EP @ 40 C CT SP @ 50 C CT 420 Avg EP @ 50 C CT

400

380 789101112

CHWLT (C))

Figure 6-18 Suction and average evaporating pressure vs. CHWLT at 40 and 50 C CT for 30% leaked and topped up

As the leak ratio increases, signs of trend shape change for the pressure drop between average evaporation and suction pressure becoming noticeable, Figure 6-19. The pressure drop across the evaporator appears consistent with average of about 15 kPa for both 40 C and 50 C CT. The pressure drop does not vary significantly with increasing CHWLT. Increasing CHWLT means increasing ET and consequently increasing capacity. At the 30% leaked R407C, the refrigerant start to show noticeable behavior change that would have adverse effect on the system operation and performance.

96 SP and Avg Eps, R407C 50% leaked

490

470

450

430 kPa

410

SP @ 40 C CT 390 Avg EP @ 40 C CT

370 SP @ 50 C CT

Avg EP @ 50 C CT 350 7 8 9 10 11 12 CHWLT (C)

Figure 6-19 Suction and average evaporating pressure vs. CHWLT at CT 40 and 50 C for 50% leaked and topped up

Avg EP at 40 C CT

500 R407C R22 CHWLT 7 C 480 R22 R407C 10% leak CHWLT 10 C 460 R407C CHWLT 12 C R22 30% leak 440 10% leak20% leak 50% leak 10% leak 40% leak 420 20% leak30% leak 50% leak 20% leak 40% leak

Pressure (kPa) Pressure 400 30% leak40% leak 380 50% leak

360

Figure 6-20 Average evaporating pressure variation vs. pure R22 and percentage leaked R407C at various CHWLT for 40 C CT

The average evaporating pressures were plotted for range of refrigerant charges from pure R22 to R407C and all the leaked compositions. The pressure trend moves upward from R22 to R407C, and then points downward for the leaked compositions depending on the ratio of the leakage. The trend is almost consistent from pure R407C down to

97 20% leaked composition, but start to show irrational up and down behavior at higher leaked. This is an indication of the fact that the refrigerant loosing its oil miscibility and sitting inside evaporator tube, thus causing more pressure drop. This phenomenon can pose a very serious reliability problem in the existing refrigeration and air-conditioning systems

Avg EP at 50 C CT

520 R407C CHWLT 7 C R407C 500 R22 R407C CHWLT 10 C R22 10% leak 480 R22 CHWLT 12 C 10% leak 460 20% leak30% leak 10% leak 40% leak 20% leak 50% leak 440 30% leak 40% leak

Pressure (kPa) Pressure 20% leak 50% leak 40% leak 420 30% leak 50% leak

400

Figure 6-21 Average evaporating pressure variation vs. pure R22 and percentage leaked R407C at various CHWLT for 50 C CT

A similar trend is followed with the average evaporating pressure at 40 C CT. All indications point on behavior change from 20% leak onward. One of the reasons could be the refrigerant oil miscibility characteristic change which in turn can cause partial oil blockage in the evaporator tube. When a system start to show this type of operating indications, then reliability of operation is under big question mark, and something need to be done. In a real chiller system, these could lead to freeze protection pressure or temperature cut out, and the system practically inoperable. The test rig is used to carry out tests is controlled manually, and none of the pressure and temperature safeties installed in order to make it possible to carry out tests in extreme conditions.

Each pressure line corresponds to specific LCHWT, i.e. 7, 10 and 12 C. Also, the higher CT or condensing temperature indicates the higher average evaporating temperature. This is due to the fact that the higher condensing pressure has its effect even after the expansion device such as thermal expansion valves. The expansion devices cause specific pressure drop at constant enthalpy.

98 Capacity variations at 40 C CT

2900 R22 CHWLT 7 C 2800 R407C 10% leak CHWLT 10 C 2700 20% leak R22 CHWLT 12 C 2600 30% leak R407C 10% leak 2500 R22 40% leak 20% leak 50% leak 30% leak 2400 R407C 10% leak 40% leak

Capacity (W) 20% leak 50% leak 2300 30% leak 40% leak 2200 50% leak 2100 2000

Figure 6-22 cooling capacity variation vs. pure R22 and percentage leaked R407C at various CHWLT for 40 C CT

Capacity variations at 50 C CT

2700 2600 R22 CHWLT 7 C 2500 R407C CHWLT 10 C 2400 10% leak CHWLT 12 C R22 20% leak 30% leak 2300 R407C 2200 10% leak 40% leak R22 20% leak 50% leak 2100 30% leak Capacity (W) R407C 40% leak 2000 10% leak 50% leak 20% leak 1900 30% leak 40% leak 1800 50% leak 1700

Figure 6-23 cooling capacity variation vs. pure R22 and percentage leaked R407C at various CHWLT for 50 C CT

Figure 6-22 and Figure 6-23 show declining trend of the capacity from pure R22 towards R407C and continue to decrease as the leak percentage increases. The trend is similar for both 40 C and 50 C SCT. With R407C when compared with R22, the system capacity drops between 4 to 6%, depending on the pressure ratio. The capacity is higher for 40 C than the 50 C CT, and higher CHWLT result in higher capacity. These are due to two factors, nature of positive displacement compressors and phase change and enthalpy change behavior of the refrigerant; the higher CHWLT corresponds to higher

99 evaporating temperature and pressure, hence higher density, hence higher mass flow. This is vice versa with the SCT or condensing pressure due to mainly the gas phase change behavior on the P-h diagram.

The average capacity loss at 30% leak when compared to pure R407C is about 15%, however, this loss goes up to about 25% with the 50% leak. In the commercial chiller systems, a capacity loss of up to 10% may be tolerated, but losses beyond that will not be acceptable. Therefore, in addition to other factors such as freeze up protection, any leak over 20% would need serious system evaluation.

Power Input variations at 40 C CT 900 CHWLT 7 C R22 CHWLT 10 C 880 R407C 10% leak CHWLT 12 C 20% leak 30% leak R22 860 50% leak 40% leak R407C 10% leak 840 R22

Power Power Input (W) 20% leak 30% leak R407C 10% leak 820 50% leak 20% leak 40% leak 30% leak 50% leak 800 40% leak

Figure 6-24 Power input variation vs. pure R22 and percentage leaked R407C at various CHWLT for 40 C CT

Power input decreases from R22 down towards pure R407C and continuous to decline as the leak percentage increases, as it can be seen on Figure 6-24. Again we can notice kind of declining trend down to 20% leak, but unusual behavior from the 30% leak onward. This confirms the earlier interpretation that a drastic change in the refrigerant’s thermodynamic and thermo-physical property.

100 Power Input variations at 50 C CT 1040 CHWLT 7 C 1000 R22 R407C CHWLT 10 C 10% leak CHWLT 12 C 20% leak 960 30% leak 40% leak R22 R407C 50% leak 920 10% leak20% leak 30% leak R22 R407C 40% leak50% leak

Power Input (W) 880 10% leak 20% leak30% leak 50% leak 840 40% leak

800

Figure 6-25 Cooling capacity variation vs. pure R22 and percentage leaked R407C at various CHWL for 50 C CT

Power Input variations at 50 C CT 1040 CHWLT 7 C 1000 R22 R407C CHWLT 10 C 10% leak CHWLT 12 C 20% leak 960 30% leak 40% leak R22 R407C 50% leak 920 10% leak20% leak 30% leak R22 R407C 40% leak50% leak

Power Input (W) 880 10% leak 20% leak30% leak 50% leak 840 40% leak

800

Figure 6-25 shows similar trend as of 40 C SDT, Figure 6-25, however, indicates almost the similar power input values for both pure R22 and pure R407C refrigerants and slightly different trend on the 12 C CHWLT curve. It is clear that the power input curves with higher CHWLT are higher than those with lower ones. Also the higher CHWLT correspond to higher evaporating temperature or pressure and consequently higher ET and SP. Although the pressure ratio or differential is lower with the higher

101 SP, but increase in the refrigerant mass flow rate reflected in increased capacity requires higher power input.

The power input is higher in general for the 50 C CT than 40 C CT. This is due to the higher pressure differential.

COP variations at 40 C CT 3.5 CHWLT 7 C 3.3 CHWLT 10 C CHWLT 12 C R22 10% leak 3.1 R22 R407C 20% leak R22 R407C 10% leak

COP 20% leak30% leak 30% leak40% leak 2.9 R407C 10% leak 50% leak 20% leak 30% leak 50% leak 40% leak

2.7 50% leak

2.5

Figure 6-26 Coefficient of Performance (COP) variation vs. pure R22 and percentage leaked R407C at various CHWLT for 40 C CT

Overall Coefficient of Performance (COP) shows declining trend from the pure R22 to pure R407C and continuous downward as the leak increases for both 40 C and 50 C CTs as well as all the CHWLTs. The COP varies from the maximum of about 3.2 at pure R22 and 12 C LCHWT down to about 2.65 at 50% leak and 7 C CHWLT.

The tests indicate minor COP drop from pure R407C to 20% leak, however, it becomes significant at 50% leak, Figure 6-26.

102 COP variations at 50 C CT

2.8 CHWLT 7 C 2.6 R22 CHWLT 10 C R22 CHWLT 12 C R407C 10% leak 30% leak R22 20% leak 2.4 R407C COP 10% leak 20% leak 30% leak 40% leak 50% leak R407C 10% leak 40% leak 50% leak 2.2 20% leak 30% leak 40% leak 50% leak 2

Figure 6-27 Coefficient of Performance (COP) variation vs. pure R22 and percentage leaked R407C at various CHWLT for 50 C CT

A similar trend was observed at the 50 C CT for various CHWLTs, with lower overall COP when compared to the 40 C CT.

Declining COP at higher leaks is not a major factor, but it is another contributing factor to re-evaluate system beyond 30% leak. This could become more and more influencing factor as more and more nations are becoming energy efficiency and environmental conscious.

103

7 SIMULATION AND MODELING RESULTS To simulate and calculate thermodynamic properties of the refrigerants including blends, a software program has been developed by a major player in the refrigeration industry, specifically refrigerant producer and supplier, Du Pont. The software is called Genetron. The ASHRAE standard is used as a reference for the Enthalpy and Entropy, being 0 at -40 C.

7.1 R22 and Pure R407C

Depending on suitability, the program uses an Equations of States for a specific family of refrigerants. The Equations of State that used in this prediction software are discussed in section 3.5 in details. Pure and mixed refrigerants properties are calculated using relevant Equation of State to get the best accuracy possible. This method makes it flexible in order to get the best and most accurate results.

The program also is capable of performing calculations for a basic refrigeration system. It requires inputting the detailed cycle information at the various points. Namely, average evaporating, average condensing, super heat at the compressor suction, gas temperature at the evaporator outlet and liquid sub-cool before the expansion system. On the compressor, it requires swept volume rate and isentropic efficiency to perform the compression calculations. The volumetric efficiency, however, is not considered as an input variable. This factor can affect accuracy of the calculations; however, it’s influence on is considered insignificant on comparison basis. It could vary from one condition to the other, which will be discussed later in this chapter.

Only the pure refrigerants could be simulated using this software. Due to the refrigerants composition change that occurs during the process of leak and later top up, the thermodynamic properties and performance could not be simulated in this section.

104

Figure 7-1 Genetron software calculation results screens for R22

105

Figure 7-2 Genetron software calculation results screens for R407C

Performance calculations were carried out, using the Genetron software, on both R22 and R407C refrigerants at various conditions. It uses basic refrigeration cycle with a few data input options. The Compressor displacement rate, isentropic efficiency, means condensing and evaporating temperatures, and other cycle details were used as input data. The software program was able to come up with all calculated performance of the cycle and thermodynamic properties of the refrigerant at various states. It should be noted that due to the limited number of decimal points, the compressor displacement has not been shown in these screen print outs. The displacement of the compressor that was used in the testing was 0.0008 m3/s.

106 Genetron vs. Test R22, 40 C CT 3000

2500 Capacity-Genetron Capacity-Test 2000 Power-Genetron

(W) Power-Test 1500

1000

500 71012 CHWLT (C)

Figure 7-3 Capacity and power input comparisons, Genetron vs. test results, R22 at 40 C CT

Genetron vs. Test R22, 50 C CT 3000

2500

Capacity-Genetron 2000 Capacity-Test (W) 1500 Power-Genetron Power-Test

1000

500 71012 CHWLT (C)

Figure 7-4 Capacity and power input comparisons, Genetron vs. test results, R22 at 50 C CT

Figures 7-3 and 7-4 show capacity and power input comparisons for R22 at 40 C and 50 C saturated discharge temperature, between the test and simulation results at various chilled water leaving temperatures.

The simulation performance results indicate slightly higher capacity and slightly lower power than the test results. These differences can be explained due to the fact that the simulation calculations do not take into account losses associated with motor efficiency 107 and volumetric efficiency. Also, in any thermal testing system, here pilot test rig, some tolerances attributed to the instrumentation accuracy. Considering the earlier two major parameters, the variations are at acceptable level.

Genetron vs. Test

3000 pure R407C, 40 C CT

2500

2000 Capacity-Genetron

(W) Capacity-Test 1500 Power-Genetron Power-Test 1000

500 71012 CHWLT (C)

Figure 7-5 Capacity and power input comparisons, Genetron vs. test results, R407C at 40 C CT

Genetron vs. Test pure R407C, 50 C CT 3000

2500

2000 Capacity-Genetron Capacity-Test (W) 1500 Power-Genetron Power-Test 1000

500 71012 CHWT C

Figure 7-6 Capacity and power input comparisons, Genetron vs. test results, R22 at 40 C CT 108 Figures 7-5 and 7-6 shows capacity and power input comparisons for R407 C at 40 C and 50 C saturated discharge temperature, between the test and simulation results at various chilled water leaving temperatures.

Capacity and power variations appear to be wider at 7 C chilled water leaving temperature and get closer as the temperature increases. Higher the chilled water temperature means lower pressure ratio and also higher the temperature resolution which helps lesser compression chamber leak ratio hence higher volumetric efficiency and improving measurement accuracy.

Also, the average variation in the 50 C condensing temperature appears to be lower than the 40 C one. Given that the earlier has higher pressure ratio than the latter one, the compression chamber leak ratio would be expected to be higher with the earlier. Hence lower volumetric efficiency than the latter.

7.2 Modeling of the system charged with R407C and then subjected to leak

Any refrigeration system can be subjected to adiabatic leak or isothermal leak. However, the adiabatic leak process can happen only at very exceptional conditions where the leak occurs at a very fast paste with totally insulated piping and vessels. On the other hand, vast majority of the leak cases occur at isothermal process where the leak process is slow and can be assumed constant temperature. The test rig in this work has been subjected to isothermal leak process.

Comprehensive leak process simulations and experiments produced trends of R32/R125/R134a composition change as function of rate of leak. The vapor and liquid mass fraction of the most volatile refrigerant (R32) decreases during the vapor leak. As a result, the mass fraction of the least volatile component increases. Starting from this point, equations have been created to represent the composition change as function of percentage of the leak.

Using this method, the liquid composition of the leaked refrigerant has been determined. Since the system topped up for the leaked amount by liquid R407C, the new refrigerant composition needs to be determined. These new compositions have been calculated using proportional combination formulas.

109 7.2.1 Calculating the new mixture properties after the leak

Thermodynamic properties of the new refrigerant mixtures have been determined Refprop program. The new refrigerant mixtures came to existence after the system subjected to the isothermal leak at various percentages and topped up with R407C. Therefore, they can not be considered as standard refrigerant mixture that assigned by ASHRAE, hence no standard property tables.

This program, developed by the National Institute of Standards and Technology (NIST) which makes it possible to define a new refrigerant mixture and determine their properties. It is based on the most accurate pure fluid and mixture models currently available. It implements three models for the thermodynamic properties of pure fluids: equations of state explicit in Helmholtz (1882) energy, the modified Benedict-Webb- Rubin (1995) equation of state, and an extended corresponding states (ECS) model. The Helmholtz (1882) energy of the mixture components mixing rules applied. This program makes it possible to define new refrigerant mixtures using non standard composition values, and determine the thermodynamic properties.

Mass Fraction Change 0.9

3 2 0.8 y = 0.625x - 0.375x + 0.3x + 0.6 R2 = 1 0.7

0.6

0.5 X R134a X R32 0.4 3 2 X R125

Mass Fraction Mass y = 0.2083x - 0.375x - 0.0833x + 0.3 0.3 R2 = 1 Poly. (X R134a) Poly. (X R32) 3 2 0.2 y = -0.2083x + 0.125x - 0.0667x + 0.1 Poly. (X R125) R2 = 1 0.1

0 0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 Mass fraction Leak out

Figure 7-7 Mass fraction change of the ternary blend R134a/R32/R125 and polynomial curve fitted models

110 Composition change behavior of a system subjected to isothermal leak is plotted on the graph, Figure 7-7. The graph clearly indicates that during isothermal leak from the system, R32 and R125 mass fractions drops while R134a mass fraction increases inside the ternary mixture. The mass fraction decrease of R32 faster than R125. This is due to the fact that R32 is more volatile than R125 and therefore tend to boil first and increase its concentration in the vapor phase. The fitted polynomial equations matched perfectly the trend so that the trend graphs are completely covered by the graphs plotted by the equations. The calculated value of R 2 = 1 indicates a perfect match with close to zero deviation between the fitted equations and the data gathered from the tests.

Table 7–1 The calculated new compositions after the system subjected to various leak percentages without adding Leak mass fraction LRi Xi134a Xi125 Xi32 Total composition 0 0.52 0.25 0.23 1 0.1 0.546875 0.2443717 0.2181283 1.009375 0.2 0.57 0.2399936 0.2000064 1.01 0.3 0.593125 0.2356159 0.1768841 1.005625 0.4 0.62 0.2299888 0.1500112 1 0.5 0.654375 0.2218625 0.1206375 0.996875

XNi 134aR = − LRi 52.0)1( +× × XLR ii 134aR 7.1

XNi R125 LRi 25.0)1( +×−= × XLR ii R125 7.2

XNi R32 LRi 23.0)1( +×−= × XLR ii R32 7.3

Table 7–2 The calculated new compositions after the system subjected to various leak percentages and topping up the system New Composition after leak and top Up % leaked and topped up XN 134a XN 125 XN 32 Total composition 10% 0.5226875 0.24943717 0.22881283 1.0009375 20% 0.53 0.24799872 0.22400128 1.002 30% 0.5419375 0.24568477 0.21406523 1.0016875 40% 0.56 0.24199552 0.19800448 1 50% 0.5871875 0.23593125 0.17531875 0.9984375

7.2.2 Calculating the system performance including capacity, power input and COP

Using the above curve fittings, equations and tables, the Refrprop software were used to calculate thermodynamic properties of the refrigerant mixtures with changed compositions. Two stages of process applied, first subject the system to leak at various fractions (percentages), calculated the new compositions’ mass fractions. At the second

111 stage, each percentage leak topped up with the same percentage of the pure R407C refrigerant which would create new composition mass fractions. Mass fraction of the new composition then calculated using equations 7.1 to 7.3 as well as the table 7.1. Thermodynamic properties of the new composition then were calculated using the Proref software and tables were created.

To complete the modeling, the calculated thermodynamic properties were downloaded to Excel spread sheet for further system performance determinations including capacities, power inputs and COPs at various conditions. The formulas from Chapter 3 used to determine capacities, power inputs as well as COPs at the same conditions the tests were to be conducted. After calculation of the above, comparisons were made with the test results at the corresponding conditions. The following graphs show the comparison data which indicates performance prediction with a reasonably good accuracy. It should be noted that theoretical performance prediction in thermal systems, specifically refrigeration systems, with fewer than 10% accuracy is considered a good modeling result.

Capacity comparison, Test vs modeling at 40 C CT

2900

2700

2500 7 C Test 7 C Modelling (W)2300 10 C Test 10 C Modelling 2100 12 C Test 12 C Modelling 1900

1700

1500 10% leak 20% leak 30% leak 40% leak 50% leak

Figure 7-8 Comparison of the capacities from the test results and the modeling at various CHWLTs and various leak percentages, 40 C CT

112 Capacities calculated by modeling predictions and test results at 40 C condensing temperature were compared on Figure 7.8. Comparisons were also made at various chilled water temperatures. Both test and model results show similar trend as function of percentage of leak that the system subjected to. The biggest variation is seen at 20% leak (about 3% deviations) and smallest at 50% leak (less than 1% deviation). With such a small variations, this proves an excellent match between the predicted performance results and tested performance results.

Capacity comparison, Test vs modeling at 50 C CT

2500

2400

2300

2200 7 C Test 2100 7 C Modelling (W) 10 C Test 2000 10 C Modelling 1900 12 C Test 1800 12 C Modelling

1700

1600

1500 10% leak 20% leak 30% leak 40% leak 50% leak

Figure 7-9 Comparison of the capacities from the test results and the modeling at various CHWLTs and various leak percentages, 50 C CT

The capacity comparison between the modeling prediction and the test results at 50 C condensing temperature is shown on Figure 7-9. The modeling capacity change trends as function of leak percentage indicates similarity to that of the test results, however, with slightly higher deviations of up to 5%.

113 Power Input comparison, Test vs modeling at 40 C CT

900

850

800 7 C Test 7 C Modelling (W) 10 C Test 750 10 C Modelling 12 C Test 700 12 C Modelling

650

600 10% leak 20% leak 30% leak 40% leak 50% leak

Figure 7-10 Power input comparison from the test results and the modeling at various CHWLT temperatures and various leak percentages, 40 C CT

Calculating the required power to compress certain amount of gas using theoretical modeling is a challenge. There are quite a few minor factors that affect the compression process and are not easy to estimate extend of their effect on the outcome. In this case, a polytropic compression model is used to compress gas inside a chamber with suction and discharge valves. This model and fundamentals of thermodynamic properties determination are discussed in the section 3.6 and chapter 3 in details.

The model prediction of power input at 40 C condensing is shown on the Figure 7-10 with comparison to the test results. The model predicts slight power input decrease slop as the leak percentage increases. At 12 C CHWLT, the graph indicates accurate enough prediction with less than 3% deviation which increases to over 5% at 50% leak. The deviations are slightly higher at 10 and 7 C LCHWT to a bit over 6% and increased gap at 50% leak.

114 Power Input comparison, Test vs modeling at 50 C CT

1000

950

900 7 C Test 7 C Modelling (W) 10 C Test 850 10 C Modelling 12 C Test 800 12 C Modelling

750

700 10% leak 20% leak 30% leak 40% leak 50% leak

Figure 7-11 Power input comparison from the test results and the modeling at various CHWLTs and various leak percentages, 50 C CT

Power input comparison at 50% as shown on Figure 7-11, indicates relatively better prediction with average deviation of less than 4% and maximum of about 5%. Deviations between the model and tests are generally consistent with various LCHWT conditions.

Minor trend change on the test result at the 50% leak represents slight inconsistency which can be explained by variations in which the system components characteristics may be affected. This could well be behavior of the expansion valve at a very high leak condition causing unstable operation of the component and consequently pressure fluctuations. This phenomenon can also be seen at 40 C condensing as well.

Despite complexity of the power input calculations and the fact couple of effecting parameters are hard to predict accurately enough, the out come is very promising. The modeling predicted power with encouraging accuracy with average deviation of 3 to 4% and maximum of around 5% at complete range of conditions.

115

COP comparison, Test vs modeling at 40 C CT

3.5

3.3

7 C Test 3.1 7 C Modelling COP 10 C Test 10 C Modelling 2.9 12 C Test 12 C Modelling

2.7

2.5 10% leak 20% leak 30% leak 40% leak 50% leak

Figure 7-12 COP comparison from the test results and the modeling at various CHWLTs and various leak percentages, 40 C CT

The Coefficient of performance (COP) values were calculated by dividing capacity by power input and plotted on the Figures 7-12 and 7-13 at various conditions vs. the leak percentages. The model prediction indicates very small change in COP values as function of the leak percentages with a slight downward trend to 30% leak and slight upward trend toward 50% leaks. The slight increase at high leak is due to increase in concentration of R134a refrigerant in the system. R134a is a more efficient refrigerant when compared to the other components of the ternary blend which constitutes R407C.

At 40 C condensing, the average deviation between the modeling predicted values and the test results is around 3% with maximum reaches over 6% at the highest leak percentage. The jump can only be related to combination of system inconsistency, instrumentation accuracy and test method effects at high leak conditions.

116 COP comparison, Test vs modeling at 50 C CT

2.8

7 C Test 2.6 7 C Modelling COP 10 C Test 10 C Modelling 2.4 12 C Test 12 C Modelling

2.2

2 10% leak 20% leak 30% leak 40% leak 50% leak

Figure 7-13 COP comparison from the test results and the modeling at various CHWLTs and various leak percentages, 50 C condensing

The COP comparison graph at 50 C condensing, Figure 7-13, shows similar trend as of the 40 C condensing. The COP values at the earlier vary from 2.7 up to 3.1, while these values range from 2.1 up to 2.4. This is due to thermodynamic fundamentals of the vapor compression cycle which requires higher power to compress at higher pressure ratios.

On the other hand, the test results show slight drop in COP at the high leak percentages. This difference in trend may look a bit unusual at fist glance; however, this can be explained in the context of the system performance. Since COP is product of dividing capacity by power input, the variations and deviations were better explained on the related graphs.

Modeling a refrigeration system that is subjected to leaked Zeotropic refrigerant at various mass fractions and conditions is a very challenging task. Fragmented works have been done on this subject, but this is the first work that completely models a system that isothermally leaked and then topped up with the blend r407C refrigerant.

117 Despite the enormous challenge faced, in overall, the modeling predicted performance values accurately enough and the out come is very promising. The modeling predicted power with encouraging accuracy with average deviation of 3 to 4% and maximum of around 5% at complete range of conditions.

118 8 CONCLUSIONS AND RECOMMENDATIONS

8.1 GENERAL CONCLUSIONS

This research work can be categorized in two major areas.

First, comparing and analyzing system performance when converted from the HCFC-22 to HFC-407C. These types of drop in conversions are not unusual in the industry. Therefore, it is very important to understand what to expect after the conversion from all aspects such as capacity, efficiency, pressures and temperatures. The test results proved that HFC-407C is the closest possible to HCFC-22 when it comes to refrigerant conversion in those types of equipments with exception of the chillers of flooded evaporator types.

When charged with the new refrigerant, the system capacity dropped in the range of 4 to 6%, depending on the pressure ratio, and power input remained almost the same. The condensing and evaporating pressures increased slightly in the vicinity of about 5%, which has a negligible factor on the pressure vessel and piping design. However, gliding temperature difference (GTD) phenomenon in both evaporator and condenser was the major difference, meaning no longer the temperature constant phase change. This phenomenon has more prominent effect on the evaporation than condensation. It reduces heat transfer between the refrigerant and to be cooled fluid, here water. And also for this reason, chillers with flooded evaporators can not be converted to HFC- 407C. Pool boiling would end up concentration of low volatile component of the mixture in the liquid.

The tests also proved that, for a given condensing temperature or pressure, the capacity and power input increases as the evaporating temperature or pressure rises. Also for a given evaporating temperature or pressure, the capacity decreases and power input increases as the condensing temperature rises. This behavior is line with the single speed positive displacement compressors such as the reciprocating type used for purpose of this research work.

The second and most important area to cover was to analyze the effect of system refrigerant gas leak on performance, efficiency and reliability of the system. The system 119 charged with the new HFC-407C was deliberately subjected to refrigerant leak at various leak stages. It is a common practice in the industry these days to evacuate and completely recharge when part of the new refrigerant blend was leaked from the system. This has proved to be extremely costly exercise with grave environmental ramifications.

The aim was to objectively determine to what extend the gas leak can be still acceptable without going through the expensive complete gas charge. The effect of leak was tested and verified at 10% steps, from 10% up to 50% mass fraction for the total charge. At the leaks beyond 30%, the adverse effect on the capacity becomes more significant, from 8 to about 15% decrease. While the power input decreased at slower pace, from 3% up to about 8% depending on the test conditions. This translated to COP decrease ranging from 4 to about 7%. This capacity loss and efficiency decrease are significant figures which suggests that the system, here chiller, can not be allowed to degrade the performance to that extend and still continue operating.

The system was also mathematically modeled to predict detailed system performance and effect of the leak at various conditions. To make this feasible and accurate enough, two separate approaches made, first system performance for pure R22 and R407C, and second system subjected to range of leak fractions. The earlier model was relatively straight forward when compared to the latter. Modeling a system charged with R407C ternary mixture and subjected to range of leaks poses enormous challenge.

The first part has been successfully modeled and predicted all the factors and performance with excellent accuracy when compared to the test results. In these approach pure refrigerants R22 and R407C were used and simulated the system behavior at range of conditions.

However, the second part was the most challenging ever. Comprehensive leak process simulations produced trends of R32/R125/R134a composition change as function of rate of leak. Starting from this point, equations have been created to represent the composition change as function of percentage of the leak. Using this method, the liquid composition of the leaked refrigerant has been determined. Since the system topped up for the leaked amount by liquid R407C, the new refrigerant composition needed to be determined. These new compositions have been calculated using proportional combination formulas. The system thermodynamic cycle was also modeled to calculate

120 capacity, power input and COP at the range of the conditions. Despite many affecting parameters and complexity of the model, the mathematical model successfully predicted the test outcome with a very reasonable accuracy, averaging around 3% with some times reaching to 5 to 6%. Needless to say that achieving accuracy of this level should be considered as fulfilling effort of this work.

In addition to this, another major factor is the effect of large leak on the system reliability. Chiller systems are normally fitted with pressure and temperature controls set at specific range to protect the system from various malfunctions such as low pressure, high pressure and freeze up etc.

The tests indicated that up to about 20% mass fraction leak had adverse effect on the system capacity of about 5%. While at the same range, the efficiency (COP) decrease was only less than 3%.

8.2 RECOMMENDATIONS FOR FUTURE WORK

This research work focused on the effect of leak on the system performance and system reliability as well as its adverse effect on system maintenance and cost associated with tackling complexity of leak in the systems charged with zeoptropic refrigerant blend. As these systems, in the event of leak, can not be simply topped up with refrigerant, it introduces major challenges to the industry dealing with this issue. The chiller systems contain significant amount of refrigerant due to its shear size and capacity. Any complete refrigerant replacement would be a costly exercise. This work showed that, depending on the system and efficiency and safety sensitivity, leaks up to about 10-20% can be tolerated by just topping up the system charge with R407C. This is due to the fact that the efficiency and characteristic change of the system would be minimal.

The key point here is how to measure or estimate the amount of leak on a commercial system installed on site. Pressure and/or temperature on the evaporator is one way of estimating this. To this end, chilled water leaving temperature is normally set at fixed value, e.g. 7 C, which corresponds to a certain range of evaporating temperatures depending on the type the evaporator and condenser is used. Usually amount of the pressure or temperature drop relates somehow to extend of the leak out of the system. This suggests that there could be a complex relation between percentage of the leak and

121 variables like temperature and pressure. The refrigerant pressure can easily be measured on site by the technical staff, however, measuring accurate temperature depends on the system equilibrium, or in other words the refrigerant has more or less the same temperature at different parts of the system. It is recommended that the future work focuses on investigating the effect of leak amount on pressure at given temperatures in the systems charged with zeotropic refrigerants, more specifically HFC-407C. The research need to come up with proposals that would result in tangible series of graphs and correlations accurately predict the ratio of the leaked and remaining mixture as well as composition of the mixture after leak. Using these graphs and equations, the new composition of the remaining mixture could be calculated. From here, using composition mass fraction of the remaining mixture in the system, fraction or percentage of the leaked mixed refrigerant can be determined. To achieve this, reverse process of calculations need to be adopted.

In long term, these method and correlations then can be used to develop an instrument to determine state of the mixture in the system, estimating fraction of leakage. By knowing the fraction, the leaked amount can be calculated. This would give enough information together with other findings of this work, to decide whether to top up the system with R407C or recover and completely recharge the system. In overwhelming majority of the cases, they would end up only topping up the system, which in turn would save significant amount of time and money to the chiller plant owners.

122 REFERENCES

Apera C., de Roddi, F and Greco, A., “Experimental evaluation of R22 and R407C evaporative heat transfer coefficients in a vapor compression plant”, International Journal of Refrigeration, vol. 23, issue 5, August 2000.

Belghazi, M., Bontemps, A., Signe, J.C. and Marvillet, C., “Condensation heat transfer of a pure fluid and binary mixture outside a bundle of smooth horizontal tubes. Comparison of experimental results and classical model”, International Journal of Refrigeration, Vol. 24, No. 8, December 2001.

Cengel. Y, Heat Transfer – A Practical Approach, McGraw-Hill USA, 1998

Chen Jiufa and Kruse Horst, “Concentration Shift Simulation for the Mixed Refrigerants R404A, R32/R134a and R407C in an Air-Conditioning System”, HVAC&R Research, Vol. 3 No 2, April 1997

Cheng-Shu Kuo, Chi-chuan Wang, Horizontal Flow Boiling of R22 and R407C in a 9.52mm Micro-fin tube, Applied Thermal Engineering Vol.16 (1996), pp719-736

Chen, L., Taylor, A., Tolouee, C. M., Leonardi,E., Behnia, M., “A non-azeotropic mixture performance test” , Journal of Australian Refrigeration, Air-Conditioning and Heating (AIRAH), March 1991.

Connon, H.A., Drew, W.D., “Estimation and calculation of thermodynamic properties for a non-azeotropic refrigerant mixture”, International Journal of Refrigeration, Volume 6, No. 4, 1983.

Cooper, M. G., “Saturation nucleate pool boiling: a simple correlation”, International Chemical Engineering Symposium, Volume 86, 1984.

Devotta, S., Waghmare, A. V., Sawant, N. N. and Domkundwar, B. M., “Alternatives to HCFC-22 for air conditioners”, Applied thermal Engineering, Vol. 21, 2001.

123 Dowining, R.C., “Refrigerant Equations”, ASHRAE Transactions, Vol. 80, Part II, 1974.

Edmister, W. C., “Applied Hydrocarbon Thermodynamics”, Hydrocarbon Processing, Volume 47, No. 9, 1968.

Gabrielli, C. and Vamling, L., “Changes in optimal design of a dry-expansion evaporator when replacing R22 with R407C”, International Journal of refrigeration, Volume 21, Issue 7, 1998.

Gabriellii, C. and Vamling, L., “Drop-in replacement of R22 in heat pumps used for district heating – influence of equipment and property limitations”, International Journal of Refrigeration, Vol. 24, issue 7, July 2001.

Genetron 407C Product Brochure, Allied Signal Chemicals, G525-013, 1/95 5K, 1995.

Goto, M., Inoue, N. and Ishiwatari, N., “Condensation and evaporation heat transfer of R410A inside internally grooved horizontal tubes”, International Journal of refrigeration, Volume 24, Issue 7, 2001.

Gungor, K. E., and Winterton, R. H. S., “A genral correlation of flow boiling in tubes and annuli”, International Journal of Heat and Mass Transfer, Volume 29, 1986.

Heating, Ventilating and Air-Conditioning (HVAC) Systems and Equipment, ASHRAE Handbook, 2004.

Huber, M.L. and Ely, J.F.,” A predictive extended corresponding states model for pure and mixed refrigerants including an equation of state for R134a”. Int. J. Refrigeration 17: 18-31, 1994.

Huber, M.L., Friend, D.G. and Ely, J.F,.” Prediction of the thermal conductivity of refrigerants and refrigerant mixtures”, Fluid Phase Equilibria 80: 249-261, 1992.

124 Johansson, A. and P. Lundqvist, “A method to estimate the circulated composition in refrigeration and heat pump systems using zeotropic refrigerant mixtures”, International Journal of Refrigeration, July 2000.

Kandlikar, S. G., “A general correlation for saturated two phase flow boiling heat transfer inside horizontal and vertical tubes”, Journal of Heat Transfer, volume 112, 1989.

Kattan, N., Thomas, J. R. and Favrat, D., “Flow boiling in horizontal tubes: part 3 – development of a new heat transfer model based on flow pattern”, transactions of the ASME, Volume 120, 1998.

Kim M. S., Ph.D., and D. A. Didion, D. Eng., P.E., “Simulation of Isothermal and Adiabatic Leak Processes of Zeotropic Refrigerant Mixtures”, International Journal of Refrigeration, Vol. 1, issue 1, January 1995.

Kruse H., “Alternative refrigerants – Synthetic and Natural Fluids”, International Conference, Romania, Institute of refrigeration, 1996

Lemmon, E.W., “A generalized model for the prediction of the thermodynamic properties of mixtures including vapor-liquid equilibrium”, Ph.D. thesis, University of Idaho, Moscow, ID, 1996.

Lemmon, E.W. and Jacobsen, R.T., “A generalized model for the thermodynamic properties of mixtures”, Int. J. Thermophysics 20: 825-835, 1999.

McLinden, M.O., Klein, S.A. and Perkins, R.A., “An extended corresponding states model for the thermal conductivity of refrigerants and refrigerant mixtures”. Int. J. Refrigeration 23: 43, 2000.

Man-Hoe Kim, “Performance evaluation of R22 alternative mixtures in a breadboard heat pump with pure cross-flow condenser and counter-flow evaporator”, Energy, 27, 167-181, 2002.

125 Min Soo Kim, Ph.D., David A. Didion, D. Eng., PE, “Evulation of a Leak/Recharge Process of Refrigerant Mixtures”, HVAC&R Research, Vol. 1, No 3, July 1995.

Olivier Zurcher and Thomas John R., and Daniel Farvat, “In-Tube Boiling of R407C and R407C/Oil Mixtures Part II: Plain Tube Results and Predictions”, HVAC&R Research, Vol. 4, No 4, October 1998.

Reid, R.C. and Prausnitz, J.M., Poling, B.E., “The Properties of Gases”, 4th edition, Mc Graw Hill, 1987.

Report, “The Montreal Protocol on substances that deplete the ozone layer”, Department of Environment and Heritage, Australian Government, October 2004.

Samuel M. Sami and Daniel E. Desjardins, “Performance analysis of new alternatives to HCFC-22 inside air/refrigerant-enhanced surface tubing”, International Journal of Energy research, 24:759-768, 2000.

Shah, M., “Chart Correlation of Saturated Boiling Heat Transfer: Equations and Further Study”, ASHRAE transactions, Volume 88, 1982.

Shao, David W. and Granryd, Eric G., “Flow Pattern, Heat Transfer and Pressure Drop in Flow Condensation Part II: Zeotropic Refrigerant Mixtures (NARMs)”, HVAC&R Research, Volume 6, No. 2, April 2000.

Soave, G, “Rigorous and Simplified Procedures for determining the pure component Parameters in the Redlich-Kwong-Soave Equation of State”, Chemical Engineering Science, Volume 35, 1980.

Thomas, J. R., “Boiling of new refrigerants: a state of art review”, International Journal of Refrigeration, Volume 19, No. 7, 1996.

126 Tillner-Roth, R., Baehr, H.D., “An international standard equation of state for the thermodynamic properties of 1,1,1,2-tetrafluoroethane (HFC-134a) for temperatures from 170 K to 455 K at pressures up to 70 MPa”, Journal of Phys. Chem. Ref. Data 26, 1994.

Tillner-Roth, R., Yokozeki, A. , “An international standard equation of state for difluoromethane (R-32) for temperatures from the triple point at 136.4 K to 435 K at pressures up to 70 MPa”, Journal of Phys. Chem. Ref. Data 26, 1997.

Tolouee, C., M., Behnia, M. and Leonardi, E., “Computer simulation of mixed refrigerant cycles”, Journal of Australian Refrigeration, Air-Conditioning and Heating (AIRAH), May 1991.

United Nations Environment Program (UNEP), “Protecting Ozone layer”, volume 1, Refrigerants, 2001.

127