Energy and Global Warming Impacts of

HFC Refrigerants and Emerging Technologies

James R. Sand

Steven K. Fischer

Van D. Baxter

Oak Ridge National Laboratory

Oak Ridge, Tennessee

Sponsored by

Alternative Fluorocarbons Environmental Acceptability Study (AFEAS)

U.S. Department of Energy

1997

TABLE OF CONTENTS

ACKNOWLEDGMENTS

1. EXECUTIVE SUMMARY ..................................................... 1

1.1 INTRODUCTION ......................................................... 1

1.2 TECHNICAL EVALUATIONS .............................................. 3

1.2.1 Alternative and Next Generation Technologies ................................ 3

1.2.1.1 Engine Driven Systems ............................................. 4

1.2.1.2 Absorption Chillers ............................................... 4

1.2.1.3 Absorption Heat Pumps ............................................ 4

1.2.1.4 Desiccant Dehumidification ......................................... 5

1.2.1.5 Advanced Vapor Compression ...................................... 5

1.2.1.6 Evacuated Panel Insulation .......................................... 6

1.2.2 Update of the 1991 and 1994 Analyses ..................................... 6

1.2.2.1 Refrigerator-Freezers .............................................. 6

1.2.2.2 Unitary Air Conditioning Equipment ................................... 7

1.2.2.3 Supermarket Refrigeration Systems ................................... 8

1.2.2.4 Chillers ........................................................ 9

1.2.2.5 Automobile Air Conditioners ....................................... 10

1.2.3 Analysis Limitations ................................................... 10

1.3 CONCLUSIONS ........................................................ 11

1.3.1 Individual Applications ................................................. 13

1.3.1.1 Household Refrigeration ........................................... 13

1.3.1.2 Automobile Air Conditioners ....................................... 13

1.3.1.3 Chillers ....................................................... 14

1.3.1.4 Unitary Equipment ............................................... 14

1.3.1.5 Commercial Refrigeration .......................................... 14

1.4 SUMMARY ............................................................. 15

2. INTRODUCTION .......................................................... 17

3. REFRIGERATOR-FREEZERS ................................................. 21

3.1 INTRODUCTION ........................................................ 21

3.2 ENERGY EFFICIENCY ................................................... 21

3.3 REFRIGERANTS ........................................................ 22

3.3.1 HFC-134a .......................................................... 22

3.3.2 Hydrocarbon Refrigerants .............................................. 23

3.4 Blowing Agents ........................................................... 24

3.4.1 HCFC-141b and HFC Alternatives ....................................... 24

3.4.2 Cyclopentane and Other Hydrocarbons .................................... 25
3.4.3 Vacuum Panels ...................................................... 25

3.5 ALTERNATIVE TECHNOLOGIES .......................................... 26

3.5.1 Absorption Refrigeration ............................................... 26

3.5.2 Thermoelectric Refrigeration ............................................. 26

3.6 ASSUMPTIONS ......................................................... 26

3.6.1 North American Refrigerator-Freezer ...................................... 27

3.6.2 European Refrigerator/Freezer ........................................... 27

3.6.3 Japanese Refrigerator-Freezer ........................................... 28

3.7 METHODOLOGY ....................................................... 28

3.8 DISCUSSION ........................................................... 29

3.9 RESULTS AND CONCLUSIONS ........................................... 30

3.9.1 North American Refrigerator-Freezer ...................................... 30

3.9.2 European Refrigerator-Freezer ........................................... 31

3.9.3 Japanese Refrigerator-Freezer ........................................... 32

3.10 CONCLUSIONS ....................................................... 34

4. UNITARY AIR CONDITIONING .............................................. 37

4.1 INTRODUCTION ........................................................ 37

4.2 TECHNOLOGY DESCRIPTION ............................................ 37

4.2.1 Refrigerants ......................................................... 38

4.2.1.1 HFC-134a .................................................... 39

4.2.1.2 HFC Mixtures .................................................. 39

R-407C ....................................................... 39

R-410A ....................................................... 40

4.2.1.3 Propane ...................................................... 40

4.2.1.4 Ammonia ...................................................... 41

4.3 ALTERNATIVE TECHNOLOGIES .......................................... 42

4.4 ASSUMPTIONS ......................................................... 42

4.5 METHODOLOGY ....................................................... 43

4.6 RESULTS AND DISCUSSION ............................................. 45

4.7 CONCLUSIONS ........................................................ 54

5. CHILLERS ................................................................ 55

5.1 INTRODUCTION ........................................................ 55

5.2 REFRIGERANTS ........................................................ 55

5.2.1 HCFC-123 ......................................................... 55

5.2.2 HFC-134a .......................................................... 56

5.2.3 HCFC-22 .......................................................... 56

5.2.4 Other Refrigerants .................................................... 57

5.2.4.1 Ammonia Chillers ............................................... 58

5.2.4.2 Theoretical Performance .......................................... 58

5.3 ALTERNATIVE TECHNOLOGIES .......................................... 59

5.3.1 Natural Gas Engine Driven Chillers ........................................ 59
5.3.2 Absorption Chillers ................................................... 59

5.4 ASSUMPTIONS ......................................................... 61

5.4.1 Carbon Dioxide Emission Rates From Power Plants ........................... 61

5.4.2 Chiller-Specific Assumptions ............................................ 61

5.4.2.1 Chiller Capacities ................................................ 61

5.4.2.2 Refrigerant Charge ............................................... 61

5.4.2.3 Refrigerant Loss Rates – End of Life Charge Loss ....................... 62

5.4.2.4 Chiller Life ..................................................... 62

5.4.3 Chiller Energy Consumption Calculations ................................... 63

5.4.3.1 U.S. Chiller Calculations .......................................... 63

5.4.3.2 Japanese Chiller Calculation ........................................ 63

5.5 METHODOLOGY ....................................................... 64

5.6 RESULTS .............................................................. 65

5.7 DISCUSSION ........................................................... 70

5.8 CONCLUSIONS ........................................................ 71

6. COMMERCIAL REFRIGERATION SYSTEMS ................................... 73

6.1 INTRODUCTION ........................................................ 73

6.1.1 Direct Expansion Systems .............................................. 73

6.1.2 Secondary Loop Systems ............................................... 73

6.1.3 Distributed Systems ................................................... 74

6.2 ASSUMPTIONS ......................................................... 74

6.2.1 Temperature Ranges .................................................. 74

6.2.2 Refrigeration Loads ................................................... 75

6.2.3 Alternative Refrigerants ................................................ 75

6.2.4 Refrigerant Charge .................................................... 76

6.2.4.1 Direct Expansion ................................................ 76

6.2.4.2 Secondary Loop Systems ......................................... 77

6.2.4.3 Distributed Systems .............................................. 77

6.2.5 Refrigerant Emission Rates .............................................. 77

6.2.5.1 Direct Expansion Systems ......................................... 77

6.2.5.2 Secondary Loop Systems ......................................... 77

6.2.5.3 Distributed Systems .............................................. 78

6.2.6 COPs ............................................................. 78

6.2.6.1 Direct Expansion Systems ......................................... 78

6.2.6.2 Secondary Loop Systems ......................................... 79

6.2.6.3 Distributed Systems .............................................. 79

6.2.7 Compressor On-Time ................................................. 79

6.2.8 Condenser Fan Power ................................................. 80

6.2.9 System Lifetime ...................................................... 80

6.3 METHODOLOGY ....................................................... 80

6.3.1 Power Consumption ................................................... 80

6.3.1.1 Compressor ................................................... 81
6.3.1.2 Secondary Loop Pumping Power .................................... 81

6.3.1.3 Distributed System Condenser Loop Pumping Power ..................... 81

6.3.2 Direct Effect ......................................................... 82

6.4 RESULTS .............................................................. 82

6.5 CONCLUSIONS ........................................................ 83

7. AUTOMOBILE AIR CONDITIONERS ......................................... 85

7.1 INTRODUCTION ........................................................ 85

7.2 ASSUMPTIONS ......................................................... 86

7.2.1 Cycle Efficiencies ..................................................... 86

7.2.1.1 Evaporating Temperatures ......................................... 86

7.2.1.2 Approach Temperatures .......................................... 86

7.2.1.3 Subcooling And Return Gas Temperatures ............................. 87

7.2.2 Compressor Efficiency ................................................. 87

7.2.3 Clutch / Transmission / Belt Efficiency ...................................... 88

7.2.4 Auxiliary Power Consumption ........................................... 88

7.2.5 COP At Idle And Highway Conditions ..................................... 88

7.2.6 Air-Conditioner Weight ................................................ 89

7.2.7 Refrigerant Emissions And Direct Effect .................................... 89

7.3 METHODOLOGY ....................................................... 91

7.3.1 Ambient Temperature Distributions ........................................ 91

7.3.2 Compressor On-Times ................................................. 91

7.3.3 Idle and Highway Operation ............................................. 91

7.3.4 Cooling Capacity ..................................................... 92

7.3.5 TEWI Calculations .................................................... 92

7.3.5.1 Power Consumption For Air Conditioning ............................. 92

7.3.5.2 Incremental engine efficiency ....................................... 92

7.3.5.3 CO2 Emissions From Air-Conditioning Energy Use ...................... 92

7.3.5.4 CO2 Emissions From Transporting The Air Conditioner ................... 93

7.3.5.5 Direct Effect Of Refrigerant Emissions ................................ 93

7.4 RESULTS .............................................................. 93

7.5 CONCLUSIONS ........................................................ 94

8. RESIDENTIAL GAS HEATING/COOLING OPTIONS ............................. 97

8.1 INTRODUCTION ........................................................ 97

8.2 UNITARY GAS TECHNOLOGIES .......................................... 97

8.2.1 Gas Engine Driven Heat Pumps .......................................... 97

8.2.2 Gas-Fired Absorption Heat Pumps ........................................ 98

8.3.3 Desiccants .......................................................... 98

8.3 ASSUMPTIONS ......................................................... 99

8.4 METHODOLOGY ...................................................... 100

8.4.1 Central Heating and Air Conditioning ..................................... 100

8.5 RESULTS AND DISCUSSION ............................................ 100
8.6 CONCLUSIONS ....................................................... 102

9. CONCLUSIONS .......................................................... 103

9.1 GENERAL CONCLUSIONS .............................................. 103

9.2 Individual Applications ................................................. 105

9.2.1 Household Refrigeration ........................................... 105

9.2.2 Automobile Air Conditioners ........................................ 105

9.2.3 Chillers ........................................................ 105

9.2.4 Unitary Equipment ............................................... 106

9.2.5 Commercial Refrigeration .......................................... 106

10. REFERENCES ........................................................... 109

APPENDIX A: ELECTRIC POWER CO2 EMISSION RATE DATA .................... 121

A.1 REGIONAL AVERAGES AND RANGES .................................... 123

A.2 SENSITIVITY OF TEWI TO CO2 EMISSION RANGES ........................ 127

A.3 IMPACT OF FUEL PROCESSING AND TRANSPORTATION ENERGY

CONSUMPTION ON CO2 EMISSIONS ..................................... 128

A.4 IMPACT OF AVERAGE, SEASONAL, AND TIME-OF-DAY VARIATIONS IN CO2

EMISSION RATES ...................................................... 129

APPENDIX B: ATMOSPHERIC LIFETIMES AND GWPs FOR REFRIGERANTS AND

BLOWING AGENTS ....................................................... 133

APPENDIX C: SPREADSHEETS FOR REFRIGERATOR/FREEZER TEWI CALCULATIONS 139

APPENDIX D: AVERAGED HEATING AND COOLING LOADS AND HEAT PUMP

PERFORMANCE FACTORS FOR NEW SINGLE-FAMILY HOMES AND COMMERCIAL

BUILDINGS IN EUROPE .................................................... 145

APPENDIX E: UNITARY AIR CONDITIONING TEWI RESULTS .................... 151

APPENDIX F: CALCULATION OF COOLING TOWER ELECTRICAL BURDEN FOR LARGE

CHILLERS ............................................................... 161

F.1 CHILLER HEAT REJECTION ............................................. 163

F.2 ELECTRICAL AUXILIARIES FOR HEAT REJECTION ......................... 164

F.2.1 Condenser Water Pumping ............................................ 164

F.2.2 Cooling Tower Fans ................................................. 165

APPENDIX G: SPREADSHEETS FOR CHILLER TEWI CALCULATIONS .............. 167

APPENDIX H: SECONDARY HEAT TRANSFER LOOP CALCULATIONS FOR

COMMERCIAL REFRIGERATION ............................................ 175
APPENDIX I: COMPLETE RESULTS FOR COMMERCIAL REFRIGERATION .......... 179

APPENDIX J: AUTOMOBILE AIR-CONDITIONING REFRIGERANT EMISSION

SCENARIOS .............................................................. 193

APPENDIX K: AMBIENT TEMPERATURE DISTRIBUTIONS ....................... 197

APPENDIX L: TEWI FOR AUTOMOBILE AIR CONDITIONING ..................... 201

APPENDIX M: AUTOMOBILE AIR CONDITIONING (EQUATIONS) ................. 207

APPENDIX N: RESIDENTIAL GAS HEATING/COOLING TEWI RESULTS ............ 211
List of Figures

1. Factors to consider in an Integrated Approach to select an optimum solution for a given heating,

cooling, or refrigeration requirement. ........................................... 2

2. Historical and projected refrigerator-freezer energy improvements for the United States. ....... 22

3. Measured energy use for European 3-star refrigerator/freezers using CFC alternatives. ........ 23

4. TEWI results for a 510 liter (18 ft3) North American, top-mount refrigerator/freezer insulating foam

options. ............................................................... 31

5. TEWI results for a 230 liter European refrigerator/freezer versus cabinet insulating foam options. . 32

6. TEWI results for a 355 liter Japanese refrigerator/freezer vs. cabinet insulating foam options. .... 33

7. Efficiency ratings of unitary air conditioners in the United States (sales weighted average). ...... 37

8. TEWI for residential heating and cooling options in Miami (1996 efficiency levels). ........... 46

9. TEWI for residential heating and cooling options in Pittsburgh (1996 efficiency levels). ......... 46

10. TEWI for residential heating and cooling options in Atlanta (1996 efficiency levels). ......... 47

11. TEWI for residential heating and cooling options in Pittsburgh (2005 efficiency levels). ........ 47

12. TEWI for residential heating and cooling options in Miami (2005 efficiency levels). .......... 48

13. TEWI for residential heating and cooling options in Atlanta (2005 efficiency levels). ......... 48

14. TEWI for 7.5 ton roof top air conditioning units in the USA. ........................... 50

15. TEWI for residential heating options in Europe. ..................................... 51

16. TEWI for residential cooling-only options in southern Europe (Greece). ................... 52

17. TEWI for commercial heating and cooling options in Europe. .......................... 52

18. TEWI for 2.8 kW room air conditioner/heat pump in Japan. ........................... 53

19. TEWI for a 14.0 kW packaged air conditioner/heat pump in Japan. ..................... 53

20. TEWI for 1200 kW (350 ton) electric driven chillers in an Atlanta. ...................... 65

21. TEWI for 3500 kW (1000 ton) electric driven chillers in an Atlanta office application ........ 66

22. TEWI for 3500 kW (1000 ton) chillers in Atlanta across a range of CO2 emission rates (2005

efficiencies). ............................................................ 67

23. TEWI for 3500 kW (1000 ton) chillers in Atlanta across a range of operating hours (2005

efficiencies). ............................................................ 67

24 TEWI for 1055 kW (300 ton) chillers in Tokyo, Japan. ............................... 68

25 TEWI for 3500 kW (1000 RT) gas-fired chillers in Atlanta.. ........................... 69

26 TEWI for 1200 kW (350 RT) gas-fired chillers in Atlanta, Georgia USA. ................. 69

27. Benefits of heat recovery for chiller plants. ........................................ 71

28. Medium temperature refrigeration in Europe. ....................................... 83

29. Evaporating temperatures. .................................................... 86

30. Air-conditioner system efficiency at low engine speeds. ............................... 88

31. Air-conditioner system efficiency at highway speeds. ................................. 89

32 . Compressor operation and average daily temperature. ............................... 91

33. TEWI for HFC-134a, HC-290, and CO2 air conditioners in selected design climates. ........ 94

34. TEWI for residential gas heating/cooling options in Pittsburgh, Pennsylvania USA. .......... 101

35. TEWI for residential gas heating/cooling options in Atlanta, Georgia USA. ................ 101

36. TEWI for residential cooling options in Miami, Florida USA. .......................... 102

37. TEWI for residential heating only options in Europe. ................................ 102
38. Average power plant emission rates and electricity production for European nations ......... 127

39. TEWI for European refrigerator/freezer using high (0.98 kg CO2/kWh) and low (0.0 kg CO2/kWh)

emission rates for power plants (HFC-134a, HCFC-141b, 15 year lifetime, 200 kWh/y). . 128

40. Sensitivity of TEWI for chillers to power plant CO2 emission rates. ..................... 130

41. Heat flows for an absorption chiller. ............................................ 163

42. Low temperature refrigeration in Europe. ........................................ 181

43. Low temperature refrigeration in Japan. ......................................... 181

44. Low temperature refrigeration in North America.. .................................. 182

45. Medium temperature refrigeration in Europe. ...................................... 182

46. Medium temperature refrigeration in Japan. ....................................... 183

47. Medium temperature refrigeration in North America. ................................ 183

48. Air-conditioner charge with “as manufactured” emission rate of 35 g/year. ................ 195

49. Air-conditioner charge with 11⁄2 recharges during operational lifetime (3 recharges in 22 years). 195

50. Air-conditioner charge with two recharges during operational lifetime (4 recharges in 22 years). 196

51 . Outdoor ambient temperature distribution for Greece (average). ....................... 199

52. Outdoor ambient temperature distribution for Okinawa, Japan. ........................ 199

53. Outdoor ambient temperature distribution for the southwestern U.S. .................... 200
List of Tables

1. Baseline North American refrigerator/freezer. ....................................... 27

2. Baseline European refrigerator/freezer. ............................................ 27

3. Baseline Japanese refrigerator/freezer. ............................................ 28

4. Calculated energy use for refrigerator/freezers using alternative refrigerants and blowing agents

(kWh/day). ............................................................ 30

5. Relative efficiencies for alternative refrigerants for unitary equipment (relative to HCFC-22). .... 43

6. Performance data for unitary Japanese systems used for the TEWI calculations. .............. 44

7. Current technology (1996) efficiency data. ......................................... 45

8. Future system efficiencies (2005). ................................................ 45

9. Refrigerant emissions for 1200 kW (350 refrigeration ton) and 3500 kW (1000 refrigeration ton)

chillers. ............................................................... 62

10. Chiller integrated part load values (IPLVs) for analysis of commercial chillers (kW/RT). ....... 63

11. Assumptions for Japanese chiller calculations. ...................................... 64

12. Chiller integrated part load values (IPLVs) for analysis of commercial chillers (kW/RT). ....... 68

13. Low temperature refrigeration COPs. ............................................ 78

14. Medium temperature refrigeration COPs. ......................................... 78

15. Automobile air-conditioning high-side approach temperatures. .......................... 87

16. Compressor isentropic efficiencies. .............................................. 87

17. TEWI for HFC-134a, hydrocarbon, and transcritical CO2 air conditioners. ................ 93

18 Residential gas options. ...................................................... 100

19. Carbon dioxide emissions from power plants. ..................................... 124

20. Carbon dioxide emissions from power plants in Europe. ............................. 125

21. U.S. regional average CO2 emissions from power plants (Calm 1993). .................. 126

22. 1993 average electric power plant CO2 emissions for other regions. .................... 126

23. U.S. electricity generation mix (1996). .......................................... 129

24. Atmospheric lifetimes and 100 year GWPs for refrigerants and blowing agents. ............ 136

25. Atmospheric lifetimes and 100 year GWPs for zeotropic and azeotropic mixtures. .......... 137

26. TEWI calculations for a North American refrigerator/freezer. ......................... 141

27. TEWI calculations for a European refrigerator/freezer. .............................. 142

28. TEWI calculations for a Japanese refrigerator/freezer. ............................... 143

29. European residential building data. ............................................. 147

30. Average seasonal performance factors for HCFC-22 based heat pumps in Europe. ......... 148

31. European commercial building data. ............................................ 149

32. TEWI calculations for unitary space conditioning equipment: Pittsburgh, Pennsylvania USA. ... 153

33. TEWI calculations for unitary space conditioning equipment: Atlanta, Georgia USA. ........ 154

34. TEWI calculations for unitary space conditioning equipment: Miami, Florida USA. .......... 155

35. TEWI for 7.5 ton roof top air conditioners (RTU): cooling only. ....................... 156

36. TEWI for European residential space heating (data from the International Heat Pump Status and

Policy Review Report, PHC-AR3, IEA Heat Pump Centre, September 1994, Part 1 ... 157
37. TEWI for European residential cooling only (Greece). Source for loads and Seasonal Performance

Factors (SPF): International Heat Pump Status and Policy Review Report, HPC-AR3,

IEA Heat Pump Centre, September 1994. .................................... 159

38. TEWI for unitary heating and cooling equipment in Japan. ............................ 159

39. TEWI for packaged air conditioners (PACs) for Japanese unitary applications. ............ 160

40. Cooling tower fan power. .................................................... 165

41. Cooling tower electrical burden. ............................................... 166

42. Parametric analysis of the indirect effects for chillers ................................ 169

43. Sensitivity of TEWI in chillers to refrigerant make-up rate. ............................ 170

44. TEWI for 3500 kW (1000 RT) chillers using data from Tables 42 and 43. ............... 170

45. TEWI for direct-gas fired absorption chillers. ..................................... 172

46. TEWI for natural gas engine driven chillers using HCFC-22, HFC-134a, or R-717 (screw and

centrifugal compressors, HFC-134a). ........................................ 173

47. Direct effects taken from Table 43. ............................................. 173

48. TEWI for centralized chiller applications in Japan. .................................. 174

49. Secondary loop pumping power for commercial refrigeration applications. ................ 177

50. TEWI for low and medium temperature direct expansion refrigeration systems in North America183 .

51. TEWI for low and medium temperature distributed refrigeration systems in North America. ... 185

52. TEWI for low and medium temperature direct expansion refrigeration systems in Europe. .... 186

53. TEWI for low and medium temperature secondary loop refrigeration systems in North America.187

54. TEWI for low and medium temperature secondary loop refrigeration systems in Europe. ..... 188

55. TEWI for low and medium temperature distributed refrigeration systems in Europe. ......... 189

56. TEWI for low and medium temperature direct expansion refrigeration systems in Japan. ...... 190

57. TEWI for low and medium temperature secondary loop refrigeration systems in Japan. ...... 191

58. TEWI for low and medium temperature distributed refrigeration systems in Japan .......... 192

59. Refrigerant emissions for automobile air conditioning (North America). .................. 196

60. TEWI for automobile air conditioners in the U.S. .................................. 203

61 TEWI for automobile air conditioners in Europe. ................................... 204

62 TEWI for automobile air conditioners in Japan. .................................... 205

63. TEWI calculations for unitary space conditioning equipment: Pittsburgh, Pennsylvania USA. ... 213

64. TEWI calculations for unitary space conditioning equipment: Atlanta, Georgia USA. ........ 214

65. TEWI for European residential space heating (data from the International Heat Pump Status and

Policy Review Report, PHC-AR3, IEA Heat Pump Centre, September 1994, Part 1. ... 215
ACKNOWLEDGMENTS

The authors are deeply indebted to the great many people who contributed to this project and it

would have been impossible to perform this work without their assistance. First and foremost among

these are those who staff the sponsoring organizations, the Alternative Fluorocarbons Environmental

Acceptability Study (AFEAS) and the U. S. Department of Energy (DOE) under contract DE-AC05-

96OR22464 with Lockheed Martin Energy Research Corporation. Esher Kweller, William Noel, and

John Ryan at DOE provided encouragement and financial support. Katie Smythe, assisted by Susan

Minnemeyer, served as project manager for AFEAS. The AFEAS oversight committee worked

diligently providing technical direction to this work with special effort and contributions coming from

Bob Orfeo of Allied-Signal, Tony Vogelsberg of DuPont, and Nick Campbell and Archie McCulloch

of ICI Chemicals & Polymers.

Numerous individuals and organizations from around the world provided data, ideas and

suggestions, and spent time reviewing both the analysis methodology and assumptions, the preliminary

conclusions, and the draft report. Most of these contributors are cited in the report references. The

following persons expended significant effort in support of this work. From Europe: Guy Biesmans of

ICI Polyurethanes, Simon Brasch of Electrolux, Jean-Yves Caneill of Electricité de France, Alberto

Cavallini and Louis Lucas of the International Institute of Refrigeration, Hans Fernqvist of Volvo, Peter

Göricke of RWE Energie, S. Haaf of Linde, Jürgen Köhler of IPEK, Horst Kruse of Universität

Hannover, H. Laue of Fachinformationszentrum, Helmut Lotz of Bosch-Siemens, Jurgen Michorius of

the Dutch Electricity Generating Board, Jostein Pettersen of SINTEF, Jürgen Pannock of Whirlpool

Italia, Bert Stuij of the International Energy Agency, and Jürgen Wertenbach of Daimler-Benz. From

Japan: Kiyoshi Hara of Japan Industrial Conference for Ozone Layer Protection, Toshio Hirata of

DENSO, Shunya Hisasima of Japan Refrigeration and Air Conditioning Industry Association, Osami

Kataoka of Daiken, and Akihide Katata of Japan Electrical Manufacturers' Association. From the U.

S.: Roland Ares (independent consultant), Ward Atkinson of Sun Test Engineering, James Baker of

General Motors, Lee Burgett and Paul Glamm of Trane, Jason Glazer of Gard Analytics, Ken Hickman

of York, Glen Hourahan of the Air-Conditioning And Refrigeration Institute, Fred Keller of Carrier, Ed

Reid of the American Gas Cooling Center, Douglas Reindl of the University of Wisconsin, Bill Ryan of

the Gas Research Institute, Steve Rosenstock of the Edison Electric Institute, and Len Swatkowski of

the Association of Home Appliance Manufacturers.

The authors are also indebted to Lambert Kuijpers (cochair of the UNEP Technology Panel),

R. S. Agarwal and S. Devotta from India, and Deborah Ottinger, Reynaldo Forte, and Cynthia Gage of

the U. S. Environmental Protection Agency for their efforts in reviewing the draft reports and for their

ideas and suggestions.

Last, but not least, the authors wish to express special thanks to Phil Fairchild of ORNL for his

support and encouragement throughout this project and for his helping hand during the tough times.

1AFEAS is an international consortium of fluorocarbon chemical manufacturers.

1

1. EXECUTIVE SUMMARY

This report summarizes the results from the third phase of a project sponsored by the U.S.

Department of Energy (DOE) and the Alternative Fluorocarbons Environmental Acceptability Study

(AFEAS).1 The study is a comparative analysis of the global warming impacts of alternative

technologies for refrigeration, air conditioning, and appliance insulation that could be commercialized

during the phase-out of hydrochlorofluorocarbons (HCFCs) mandated under the Montreal Protocol.

The analysis utilizes a systems approach to determine the overall contribution to global warming from

an individual air conditioner, heat pump, or refrigeration system during its operating lifetime- a concept

known as the "total equivalent warming impact (TEWI)."

Air conditioners, heat pumps, and refrigeration systems can affect global warming through the

release of refrigerants and insulation blowing agents directly into the atmosphere and also through the

release of carbon dioxide from the generation of electricity to power the system throughout its lifetime.

The results of the TEWI analysis for refrigeration and air conditioning applications are presented in this

report.

The TEWI concept provides a useful tool in the assessment of various competing technologies.

However, it is only one of many criteria that must be considered as illustrated in Fig. 1. Safety, health,

and other environmental concerns, system initial and operating costs, regional energy considerations,

and ease of maintenance are among other important factors that must be evaluated in the selection of

the "best technology" for any given application. Finally, TEWI is dependent on the electricity fuel mix in

a given area; as such it should not be considered absolutely, but rather comparatively.

1.1 INTRODUCTION

The initial TEWI study compared the global warming impacts of alternatives to fully

halogenated chlorofluorocarbons (CFCs) in refrigeration, air conditioning, insulation, and solvent

cleaning applications that could conceivably be commercialized before the year 2000. The concept of

total equivalent warming impacts, or TEWI, was developed to combine the effects of the direct

emissions of refrigerants, polymer foam insulation blowing agents and solvents in end use applications

with the indirect effects of energy consumption from the combustion of fossil fuels and generation of

electricity used for heating or cooling. Direct contribution to TEWI is based on the use of the global

warming potentials (GWPs) developed by the Intergovernmental Panel on Climate Change (IPCC) that

use carbon dioxide (CO2) as a reference gas (GWPs of CO2 = 1.0 regardless of time horizon). TEWI

provides a measure of the environmental impact of greenhouse gases from operation, service and endof-life disposal of equipment. 
2

Figure 1. Factors to consider in an Integrated Approach to select an optimum solution for a given

heating, cooling, or refrigeration requirement.

Toxicity

Stratospheric

Ozone

Physical

Properties

Technology

Changes

Flammability

Greenhouse

Effect

(TEWI)

Energy

Efficientcy

(TEWI)

System

Costs

Safety

Environment

Performance

Integrated Analysis

Leading to the Selection of the Product

The second AFEAS/DOE study was initiated to evaluate the energy and global warming

impacts of newly developed fluorocarbon replacements for chlorine-containing refrigerants and blowing

agents as well as not-in-kind (NIK) non-fluorocarbon technologies that could conceivably be

developed or improved in the future to displace vapor compression refrigeration and air conditioning

and foam insulations (AFEAS has continued work separately on solvent cleaning applications, but these

activities are not part of the second and third phases of work sponsored by AFEAS and DOE). Next

generation refrigeration and insulation technologies were classified into three categories:

! zero ODP HFC refrigerants and blowing agents;

! existing NIK commercial technologies (i.e., ammonia (NH3) and hydrocarbon (HC)

compression, desiccant drying, evaporative cooling and absorption chillers); and

! NIK technologies not currently at a commercial level (i.e., acoustic compression, 

adsorption, Stirling, transcritical vapor compression, etc.).
3

The major change in assumptions between the first and second studies was the use of GWPs

based on a 100 year time horizon in the second study. GWPs of different gases depend on the infrared

absorption properties of the gas and the elapsed time before it is purged from the atmosphere, and are

relative to the natural cycle of carbon dioxide. The IPCC published GWPs of many gases relative to

CO2 for short, medium, and long time periods, or integration time horizons. Scientifically, there are

arguments for using GWPs based on 500 year time horizons and these were used as the basis for the

1991 AFEAS/DOE project. Since 1991, however, the prevailing argument has been that the continued

high release rates of greenhouse gases could affect the rate of climate change in the next several

decades and that policy decisions should be based on shorter time horizons.. Policy makers have

therefore chosen a 100 year integrated time horizon for global warming studies. The second study

presented the principle comparison data on a 100 year time horizon basis.

AFEAS and DOE undertook this current study to assess the significant developments that have

occurred in HFC blends (i.e. R-404, R-407C, R-410A, and R-507) and the application of nonfluorocarbons like hydrocarbons (HCs), carbon dioxide (R-744 or CO2

), and ammonia (NH3 or

R-717) as refrigerants or foam insulation blowing agents. New data with regard to thermal performance

of these compounds made it possible to perform an objective evaluation of the energy and global

warming impacts of these “third generation” refrigerants and blowing agents. Refrigerant and equipment

manufacturers have also made significant advances in the use of high pressure blends of HFCs as

alternatives to HCFC-22 in both refrigeration and air conditioning applications. Analytical and

experimental results became available to perform quantitative comparisons between HFC blends and

the application of hydrocarbons, ammonia, and carbon dioxide as refrigerants. Additionally, new

technologies for gas-fired air conditioning systems are being commercialized and operating data are

now available. 

As with the previous studies, analyses for end use applications in North America, Europe, and

Japan include the effects of cultural and technical differences in each region. The differences include

such things as the room and internal compartment temperatures for refrigerator/freezers, the sizes of

refrigerators and thicknesses of insulation used, annual driving distances for automobiles, fuel types used

for generating electricity, and climate differences for building heating and cooling loads. The results in

this report expand the work in the initial studies and indicate important regional differences in TEWI for

some applications.

1.2 TECHNICAL EVALUATIONS

1.2.1 Alternative and Next Generation Technologies

Several alternative technologies for the conventional vapor compression (reverse Rankine)

cycle employing halocarbons were identified in the Phase II (TEWI-2) report that had been laboratory

tested or had been developed to a point where they were considered as potential alternative

technologies for those currently used. For example, employing CO2 in a transcritical vapor compression

cycle or using a high efficiency absorption cycle heat pump showed sufficient promise that they were
4

revisited as potential alternatives in this study. Additional development work has occurred on tripleeffect, direct-fired absorption chillers and on absorption heat pumps for residential or light-commercial

applications so that estimates of TEWI for this equipment are included in the appropriate sections of

this report. Alternative technologies which showed little near commercial promise in the TEWI-2

report, such as thermoelectric cooling, magnetic heat pumps, thermoelastic heat pumps, etc., are not

considered here.

1.2.1.1 Engine Driven Systems

Engine driven chillers and heat pumps are considered which employ the same reverse Rankine

cooling cycle as conventional electric powered systems, but the electric motor is replaced by an internal

combustion engine. This change has very little or no effect on the direct contribution to TEWI resulting

from inadvertent releases of the working fluid to the environment, but it can change the indirect

contribution associated with CO2 emissions. Factors affecting the indirect contribution result from the 

primary energy source used to drive the cycle and any inherent differences in cycle efficiencies such as

changes in part load efficiencies or waste heat recovery, resulting from this substitution. Engine driven

air conditioning and refrigeration equipment allows consumers to select natural gas as the primary

system fuel source in situations where there are significant energy price advantages, utility rebates, or

other incentives. These opportunities have prompted considerable support for research and

development (R&D) at HVAC manufacturers and other organizations and have resulted in the

commercial availability of the newest generation of high-efficiency packaged natural gas, engine-driven

chillers and residential heat pumps. TEWI calculations are presented for gas engine-driven chillers and

heat pumps.

1.2.1.2 Absorption Chillers

Absorption chillers are commercially available and represent a major share of the commercial

air conditioning market in Japan and a portion of the market in North America. Absorption equipment

is often used in “hybrid” plants working together with electric centrifugal chillers to reduce electric peak

demands and utility demand charges. Absorption equipment can be an effective component for

managing end users’ total energy costs. Single effect absorption chillers are also used in applications

powered by waste heat, in which case the lower efficiency may not be as important. Direct-fired,

double-effect chillers can simultaneously provide chilled water for air conditioning and hot water for

space heating. Triple-effect absorption chillers are under development which are expected to be 20-

45% more efficient than current double-effect chillers.

1.2.1.3 Absorption Heat Pumps

Absorption heat pumps are under development for heating and cooling in residential and light

commercial applications. Generator absorber heat exchange (GAX) ammonia-water absorption heat

pumps are under development in Europe and the U.S.; several field test units have been built in Japan.
5

One or two manufacturers have prototype units undergoing laboratory and field testing in 1997, with a

goal to commercialize this technology by around 2000. Seasonal performance data are not available for

these prototype units. These systems have potential to reduce TEWI in areas where heating load

dominates, but may have higher TEWI in areas where cooling dominates.

1.2.1.4 Desiccant Dehumidification 

Two types of desiccant systems are available for meeting the latent cooling load in building airconditioning applications. One of these is based on solid desiccants in rotating wheels and the other on

liquid desiccants pumped between various components in the circuit. Current designs are used

primarily in special applications such as supermarkets, hospital operating rooms, and other niche

markets requiring low humidity. Further improvements are necessary in the efficiency, cost, size,

reliability, and life-expectancy of desiccant components and systems in order for them to penetrate the

broader commercial air conditioning market. Integrated systems which combine desiccant-based

components with more conventional air conditioning equipment are available. System efficiency is

greatly enhanced if the desiccants are at least partially regenerated with waste heat from other system

components. Since these systems tend towards individually-engineered or custom-designed

installations, no calculation of TEWI was attempted with the available data.

1.2.1.5 Advanced Vapor Compression

Conventional vapor compression technologies continue to be improved and efficiencies of

refrigeration and air-conditioning equipment will be higher in the future for both electric and engine

driven systems. Developments leading to these improvements include the use of higher efficiency motors

and compressors, more effective heat exchangers, and adaptive controls. Refrigerant losses from

applications such as automobile air conditioning and supermarket refrigeration have been, and will

continue to be, reduced. Indeed, supermarket equipment now under development and entering the

market, such as electric or engine driven systems that circulate a chilled secondary fluid or distributed

electrically driven compressor racks, show promise of dramatically reduced refrigerant charge and

emissions. Regulations and refrigerant costs provide an incentive to reduce emissions of refrigerant by

eliminating intentional venting during servicing, improving maintenance practices and procedures,

mandating charge recovery and recycling, and minimizing leaks. 

TEWI calculations for fluorocarbon compression systems in this report are based on

demonstrated production efficiencies, modeled efficiencies, or proven performance from R&D research

laboratory tests (which would probably show different efficiencies than optimized production designs).

Refrigerant loss rates used in the previous studies were from a time when it was standard practice to

simply vent the refrigerant charge during servicing. These practices are now prohibited in the United

States and elsewhere and it is clear that historical emission rates are not appropriate for current

calculations. Current and projected refrigerant make-up rates based on information from industry are

used for cases presented in this report. 
6

TEWIs of viable, commercially available compression systems are compared in some sections

of the report to estimated TEWIs for emerging technologies which have not yet reached the stage of

commercial production. In these instances the best available laboratory or computer modeled

performance data are used for new technologies. While helpful in identifying future technologies that

may have lower global warming impacts, the reader must be careful not to attribute too much

significance to comparisons with minor differences in TEWI because of the more uncertain nature of the

data used for these non-commercialized, emerging technologies. 

1.2.1.6 Evacuated Panel Insulation

Evacuated panels which can be used to improve appliance insulations have very low thermal

conductivities. Thin, flat panels are constructed using a filler material such as aerogel, diatomaceous

earth, or open cell filler foam enclosed by one or more plastic or metal membranes under a vacuum.

“Total panel” thermal conductivities are usually significantly higher than the “center-of-panel”

measurements usually cited, because heat transfer through the plastic or metal membranes along the

surfaces and near the edges of the panel is enhanced. Evacuated panels, at last report, are being used

by a Japanese refrigerator-freezer manufacturer for a commercial product. While evacuated panel

insulation could be an effective means to reduce TEWI for refrigerator-freezers, these panels are quite

expensive compared to the foamed polyurethane insulations usually used. There continues to be doubt

that panels used for cabinet insulation will retain a vacuum and maintain high thermal resistances over

the 15 to 20 year lifetime of an appliance. Such panels could be used in conjunction with blown foam

insulation to improve the thermal properties of appliance cabinets or to achieve comparable

performance using thinner walls, thereby permitting more usable internal volume. 

1.2.2 Update of the 1991 and 1994 Analyses

Literature searches, personal contacts, and review meetings were used to obtain up-to-date

information on alternatives to CFCs, HCFCs, and HFCs for the applications evaluated in the previous

studies. This information was then used with similar procedures and methodologies established in the

first two studies to recalculate the TEWI for those applications where significant developments and

changes have occurred since 1991 and/or 1994. Underlying assumptions on equipment lifetimes, CO2

emission rates from power generation, equipment dimensions, are consistent with those used in the

1991 and 1994 studies. Some of the key findings are summarized below:

1.2.2.1 Refrigerator-Freezers

The latest available published information indicates no significant difference between the

measured energy efficiency of refrigeration circuits utilizing HFC-134a or Iso-butane (HC-600a) as the

refrigerant.

Insulating foams blown with cyclopentane or pentane isomers consistently show higher thermal

conductivities than HCFC-141b blown foams and refrigerators produced with these HC foams would
7

have 8 to 10% higher energy consumption assuming the same foam thickness. Most of the R&D work

for this application has centered around finding an alternative for HCFC-141b, which is scheduled for a

2003 phaseout date. Current data shows HCFC-141b blown foam has the lowest thermal conductivity

and highest insulating value of the foam blowing agents investigated, which results in the lowest energy

consumption for the refrigerator design when equivalent wall thicknesses and internal volumes are

assumed. Optimized HFC-245fa or HFC-365mfc blown foam is expected to show similar conductivity

and insulating values. Vacuum panel technology can further improve cabinet thermal performance but

with significantly increased costs. It should be noted that the design of the refrigerator is a major factor

in minimizing the TEWI of the system. Designs consistent with the “average” models prevalent in each

region were postulated based on data from manufacturers and industry associations and consistent

assumptions on components, wall thicknesses, and internal volumes were applied.

The direct impacts of HFC-134a and the various halocarbon blowing agents range from 8% to

15% of the total equivalent warming impact for refrigerator-freezers in North America. One-tenth of the

TEWI for refrigerators using HFC-134a as the refrigerant and HCFC-141b as the blowing agent is due

to fluorocarbon emissions. Almost all of the direct effect is due to the foam blowing agent. Mandatory

refrigerant recovery would result in a 2 to 3% decrease in total lifetime TEWI in North America.

The direct contribution due to fluorocarbons in European refrigerator-freezers is about 19%,

primarily because the refrigerators are smaller and have lower annual energy use. The lower CO2

emissions rate for electric power generation in Europe, which has a higher percentage of nuclear and

hydroelectric power generation than North America, is also a factor. The difference between operating

a refrigerator in a country with a very low emission rate from power generation (e.g., 0.02 kg

CO2/kWh) and in one with a high rate (e.g.., 1.08 kg CO2/kWh) is much greater than the difference in

direct effects between using fluorocarbon or hydrocarbon refrigerants and blowing agents.

In 1996, hydrocarbon refrigerators were available in both manual and automatic defrost models

in parts of Europe, particularly Germany. Previously, the use of HC refrigerants had been limited to

manual defrost models because of concerns that electrically switched devices such as fans, thermostats,

defrost heaters, etc. might provide an ignition source for refrigerant that has leaked inside the cabinet.

Isobutane (HC-600a) frost-free refrigerator designs which incorporate a foamed-in evaporator and

explosion proof electrical devices or switches located outside of the food compartments are now being

built and sold. Additional safety precautions and system components have resulted in higher

manufacturing, purchasing, and servicing costs for refrigerators using HCs. Flammable refrigerants are

not used in United States or Japanese refrigerators due primarily to the higher costs required to mitigate

the associated safety risks.

1.2.2.2 Unitary Air Conditioning Equipment

Unitary air conditioners and heat pumps generally fall into four distinct categories, based on

primary use and capacity: room air conditioners with capacities between 2.0 and 10.5 kW (0.6 and 3.0
23.52 kW = 1 ton of refrigeration

8

tons2); ductless packaged and split systems ranging in capacity from 2.0-20.0 kW (0.6-5.7 tons);

ducted systems with capacities from 5.0 to 17.5 kW (1.4 to 5.0 tons); and single packaged or split

systems for commercial use whose capacity can be from 20 to 420 kW (5.7 to 120 tons). HFC

mixtures have been proposed and tested as substitutes for HCFC-22 in unitary equipment.

Hydrocarbons have been studied for this application as well. Building codes and safety concerns in

most developed countries limit the use of hydrocarbons in applications where a refrigerant leak could

result in explosive mixtures at atmospheric conditions. These restrictions apply to air-to-air heat pumps

and air conditioners in North America and Japan (90-95% of the approximately 214 million unitary

heat pumps and air conditioners in the world). Hydrocarbon refrigerants might be able to satisfy safety

requirements for the air-to-water or water-to-water unitary systems used in Europe where the entire

refrigerant charge remains outdoors.

Comparisons were made for HCFC-22 and two non-flammable HFC mixtures identified as

likely HCFC-22 replacements in the Air-Conditioning and Refrigeration Institute (ARI) Alternative

Refrigerants Evaluation Program. The direct TEWI effects for both HCFC-22 and HFC alternative

mixtures are a small fraction of the total in each case. Energy efficiency is very important for this

application and contributions to global warming from energy usage with HFC blends are expected to be

about the same as or slightly lower than those of current technology air conditioners using HCFC-22;

engineering optimization is expected to reduce energy use and resultant CO2 emissions with the

mixtures in future systems. Propane’s performance as a refrigerant for air-to-air equipment was

reduced by assuming an intermediate heat transfer loop was needed to keep this flammable refrigerant

out of the occupied space. Engine driven heat pumps and the Generator Absorber Heat Exchange

(GAX) absorption heat pumps are evaluated in a residential gas heating/cooling options chapter and

compared with a gas furnace/electric air conditioner combination.

Refrigerant make-up rates and end-of-life losses assumed in this study for 1996 vintage

equipment, which are critical in evaluating the direct effect of emissions on TEWI, were suggested by

industry experts in each region and are the same as those used for the 1994 study; a 4% annual makeup rate and 15% loss of charge upon equipment decommissioning. 

1.2.2.3 Supermarket Refrigeration Systems

These systems have historically used large refrigerant charges and experienced high leakage

rates. The current high costs of refrigerants and environmental regulations are resulting in better efforts

at refrigerant containment and lower loss rates. The most likely substitutes for CFCs and HCFCs in

supermarket refrigeration are mixtures of HFCs, although use of ammonia chillers with indirect heat

transfer loops is seeing some use in Europe. Alternative refrigerants and technologies are considered as

replacements for R-502 in low temperature refrigeration (e.g. freezers and ice cream display cases) and

HCFC-22 in medium temperature refrigeration (e.g. meat, fish, and dairy cases). The alternatives
9

include HFC mixtures in direct expansion systems using remote and distributed compressor racks and

HFC mixtures or ammonia in secondary loop refrigeration systems. 

Secondary loop systems are a means of reducing refrigerant charge and controlling leakage and

emissions, albeit with first cost and efficiency penalties. This approach to commercial refrigeration

avoids long, field erected refrigerant lines which run to individual cases in the store and confines the

refrigerant charge to a smaller, more leak-tight refrigeration circuit in the store’s equipment room. The

secondary loop approach generally must operate over a larger temperature lift to accommodate the

intermediate level of heat exchange (though the magnitude of this extra temperature lift can be minimized

by appropriated design optimization) and has an additional parasitic load associated with a fluid

circulating pump. Building codes in the developed countries make it expensive, or in some cases

prohibitive, to use ammonia in most supermarkets because of the public safety risks in densely

populated areas near the stores. When ammonia is used, secondary loops are mandatory so that the

refrigerant lines do not enter the retail sales areas of the building .

Another improved commercial refrigeration design, usually referred to as the distributed

system approach, moves the compressor with its associated high pressure liquid and suction gas lines

as close as practical to the case evaporator loads and utilizes a closed-loop, water circuit to reject the

heat of condensation. The distributed system with compressors located near or in the refrigeration cases

requires a larger number of smaller compressors located throughout the store. It, too, has a parasitic

load associated with the heat rejection water loop and pump albeit smaller than for the secondary loop

system. Both the distributed system and secondary loop approaches drastically reduce the refrigerant

charge (by as much as 75 to 90%) and make it more practical to minimize refrigerant leaks and

maintain system efficiency. 

1.2.2.4 Chillers

The air conditioning loads of larger commercial buildings are generally met with water-cooled

chillers which employ cooling towers for heat rejection and which distribute chilled water or a

water/antifreeze mixture to building air handlers and fan coil units. Centrifugal or screw compressors are

used for larger, 350 to 35,000 kW (100 to 10,000 ton), chillers because of the high volumetric flow

rates of refrigerant required. Replacement of CFC refrigerants in chillers with HCFC and HFC

alternatives has had the most significant impact on the direct contribution for this equipment. An

increased awareness of the environmental impact of refrigerants, recently enacted legislation which

requires extensive record keeping, increasing refrigerant prices, and improved equipment designs have

all served to dramatically reduce refrigerant loss from chillers. Typical annual loss rates of low pressure

refrigerant from new centrifugal chillers has been reduced more than 50-fold in seven years. New

systems are equipped with electronic alarms alerting operators to the first indications of leaks or unusual

purge pump operation. Refrigerant leak and annual make-up rates have been improved to the point

where the GWP of chiller refrigerants has very little effect on the total TEWI.

The TEWI of chillers has been reduced through simultaneous and substantial improvements in

chiller efficiencies as well. Rating point COPs for new electric chillers have increased by more than

40% (from 5.0 to 7.0) over the last ten years which has resulted in nearly a 30% decrease in the
10

indirect contribution from CO2 emissions. Market competition and a greater emphasis on lower life

cycle operating costs, as opposed to governmental legislation, are responsible for these dramatically

improved performance efficiencies. Even with these improved operating efficiencies, lifetime energy

consumption is the predominant factor influencing TEWI for this equipment.

Efficiency of natural gas engine-driven chillers and of absorption chillers has also been

significantly improved as noted on pages 4 and 14. Gas-powered chillers are sometimes used together

with electric chillers to decrease peak electrical demand and lower building operating costs.

TEWI calculations were made for both vapor compression and absorption chillers of two

discrete capacities - 1,200 and 3,500 kW (350 and 1,000 tons) - in North America. These calculations

were based on the use of Integrated Part Load Values (IPLVs) and annual operating hours for an

Atlanta, Georgia office building. Direct contributions to TEWI were computed for centrifugal and screw

chillers using HCFC-123, HFC-134a, HCFC-22, and NH3 as refrigerants for a range of annual makeup rates up to 4%. 1993 vintage CFC-11 and CFC-12 machines were included for comparison

purposes. Similar computations were made for 1,055 kW (300 ton) chiller options in Japan. In this

case, rated full-load performance data and associated annual operating hours for a Tokyo location

were used. Vapor compression chillers using HCFC-123 and HFC-134a were considered.

For the Atlanta location examined in the report, the direct contribution to TEWI for

fluorocarbon-based technologies is at most about 7% of the total (about 0.5% for the HCFC-123

machines) even when the maximum annual leak/make-up rate is assumed. Of course, both the direct

and indirect impacts are dependent upon the total operating hours assumed for the analysis. For

example, a more northerly location with half the operating hours of Atlanta would yield indirect TEWI

contributions one-half as large and the percentage contribution due to refrigerant emissions would

approximately double.

1.2.2.5 Automobile Air Conditioners

Automobile air conditioning was identified in the two previous studies as one of the few

applications in which the direct contribution of fluorocarbon refrigerant emissions was a significant

fraction of total TEWI. While the conclusion has not been contested, the approach taken in those

studies has been criticized because of reliance on efficiency data at a single design point coupled with

estimates of equivalent full-load operating hours. These two assumptions greatly simplified the analysis

but cannot account for varying performance over a range of operating conditions or the effects of

different climates. The present analysis addresses these concerns by incorporating efficiency differences

across a wide range of operating conditions, regional variations in ambient temperature, and changes in

air-conditioner on-time with ambient temperature.

Three fundamentally different cooling systems were considered; a conventional HFC-134abased system, a propane (HC-290) based system, and a transcritical vapor compression system using

CO2 as the refrigerant. The hydrocarbon system in this study includes the use of a secondary heat

transfer loop to isolate the flammable refrigerant outside the passenger compartment. This safety feature

reduces cycle efficiency relative to direct expansion systems, adds parasitic energy consumption due to

the fluid pump, and increases overall system weight.
11

Though the analysis includes much more detailed information than the earlier studies, it relies on

the same approach of evaluating energy use for operating and transporting the air conditioner and the

direct contribution of refrigerant emissions. Results are computed for thirteen different countries.

Depending on the location and assumptions about lifetime refrigerant emissions, the alternative systems

show potential for lower TEWI than the HFC-134a system. Prototypes of both alternatives are being

studied by manufacturers (the CO2 system to a much greater extent than the HC system). More

extensive prototype and field trial testing will be needed before fully developed, reliable commercial

designs will be available.

1.2.3 Analysis Limitations

Any calculation of TEWI is based on a selection of assumptions about equipment performance,

operating conditions, and CO2 emissions. All of these assumptions contain a degree of uncertainty that

is reflected in the results and this uncertainty must be considered when drawing conclusions about

differences in TEWI for two or more alternatives for a given application. The indirect contributions to

TEWIs in this report are calculated using data from the open literature for average annual power plant

CO2 emission rates for Europe, Japan, and North America. Calculations are sensitive to these CO2

emission rates and different comparisons are observed and conclusions drawn when calculations are

repeated using annual average CO2 emission rates for specific countries (e.g. Norway with virtually all

electricity from hydroelectric power, Denmark with a high percentage of coal-fired power plants) or

emission rates for only peak and intermediate power generation from specific utilities.

It is difficult to calculate an absolute value for TEWI. Most of the benefits for TEWI come

from using it as a comparative tool for assessing the relative global warming impacts of different

technology options under a controlled set of assumptions. Uncertainties exist for all of the assumptions 

(many of which are estimates or averages) that enter into the TEWI calculations. Uncertainties in the

direct effect include estimates of refrigerant emission rates and end-of-equipment life recovery rates as

well as uncertainties in the GWPs themselves. The indirect contributions from CO2 emissions can be

determined with less uncertainty (especially for specific locales where the fuel source can be

characterized accurately) but these values are still not precise. These various uncertainties minimize the

importance of small differences in TEWI comparisons between similar, established technologies that use

the same energy source. When making such comparisons, if the TEWI differences are small, the

technology that shows lower energy use should be favored as long as safety and environmental

considerations are adequately addressed and costs are reasonable. In some such cases, a decision

between options based on lowest energy consumption could result in choosing a technology with a

calculated TEWI that is slightly greater than the minimum. However, in these situations it can be argued

that resource conservation and particularly energy resource conservation should take precedence over

marginal TEWI differences, due to the uncertainties mentioned above.

It is also difficult to make accurate interfuel comparisons using TEWI as computed in this

report. At the very least “local” CO2 emission factors need to be used for electric power generation

instead of the broad regional annual averages used in this study, and some technologies may in fact

require time of year or time of day factors to account for differences in emissions due to peak power.
12

These differences are in part due to the types of generating equipment brought on line for peak demand,

changes in the local fuel mix, and possibly lower transmission and distribution efficiencies during peak

generation periods.

In some cases, sections of this report compare established technologies with known efficiencies

to systems under development for which ultimate production efficiency levels and system configurations

are less precisely known. When making these comparisons, if the emerging technology exhibits a TEWI

10% or more lower than that of the established alternative then further analysis and development may

be justifiable. TEWI, however, would not be the sole determining factor in the ultimate choice between

the alternatives.

1.3 CONCLUSIONS

Several broad conclusions can be drawn from the study.

! TEWI evaluations emphasize the combined environmental effect of the direct emission of

greenhouse gases with the indirect effects of CO2 emissions from energy use by equipment using

these fluids as refrigerants or blowing agents. This is only one criterion in selecting between

technology options. System costs, operating costs, regional energy costs, ease of maintenance,

continuing technology improvements, etc., are equally important factors to consider in selecting the

most appropriate technology for any specific application. 

! Reductions in TEWI through the use of ammonia or hydrocarbons as refrigerants are insignificant

for refrigeration systems with low emissions and may lead to an increase in energy use when

applications of these fluids must meet the same safety design criteria currently defined as acceptable

for fluorocarbon refrigerants.

- Ammonia and some hydrocarbon refrigerants have thermophysical properties comparable to

(and for some applications superior to) those of HCFC or HFC refrigerants. They also have

system irreversibilities and system design features necessary for safe products (e.g., secondary

loops) which reduce their overall efficiency. Such changes often offset much of the TEWI

benefit claimed for non-fluorocarbon refrigerants.

- Insignificant TEWI differences for most applications occur when design and service

requirements, for low refrigerant emissions and safe operation of equipment using flammable or

toxic refrigerants, are applied to systems engineered for non-flammable or non-toxic

refrigerants.

! TEWI for systems with significant direct effects from refrigerant emissions (i.e. supermarket

refrigeration, automobile air conditioning) can be reduced by improved refrigerant containment and

maintenance practices or possibly by alternative technologies. Alternative technologies (e.g.
13

secondary loops, transcritical CO2) with lower direct effects from refrigerant emissions have

potential to reduce TEWI but at lower energy efficiency.

! Efficiencies of conventional technologies are likely to increase as electric and gas-driven equipment

and insulating foam formulations are further optimized for replacement refrigerants and blowing

agents.

! Innovative design and modifications of standard practice can lead to significant reductions in TEWI

for refrigeration systems using ammonia, fluorocarbon, or hydrocarbon refrigerants. These include

mandatory refrigerant recovery and recycling, distributed refrigeration systems, charge reduction,

elimination of flared fittings and reduced numbers of brazed connections, highly efficient purge units,

improved heat transfer surfaces, high-efficiency compressors, etc. Although not included in this

study, both active and passive desiccant dehumidification and heat exchange technology are

expected to have high potential for reducing TEWI. Use of heat pump technologies for water

heating would significantly reduce energy consumption and indirect CO2 emissions as well.

! Average annual CO2 emissions from electricity generation vary widely for individual regions and

countries - from 0.0 to over 1.0 kg CO2/kWh compared to the 1993 World average of 0.58.

Emission rates also vary with season and time of day depending on how the generation fuel mix

changes. Overall TEWI values in any particular location will be peculiar to the local electrical power

generating efficiency and seasonal and time of day generating characteristics. The direct contribution

can range from all (or nearly all) of total TEWI in areas with low CO2 emission rates [using mostly

nuclear or hydro power] to a minor fraction of TEWI for areas with higher rates [using mostly

coal].

1.3.1 Individual Applications

1.3.1.1 Household Refrigeration

The phase-out of HCFCs affects TEWI for household refrigerator/freezers through the choice of

refrigerant and insulating material. High efficiency appliances rely on blown foam insulation to achieve a

high internal volume for given outside dimensions and low heat leakage into the cabinet. No clear

conclusions are possible at this time for foam blowing agents to replace HCFC-141b, although HFC

blowing agents have been identified and are under active commercial development which produce

insulation comparable to HCFC-141b blown foam. Hydrocarbon blown foams continue to have higher

thermal conductivity than HCFC-141b and HFC foams and, consequently, exhibit higher energy use

with increased impact on CO2 emissions. The increased energy use must be balanced against any direct

impact caused by the HFC blowing agent itself. 

TEWI estimates from this analysis for household refrigerator-freezers using HC refrigerants and

foams are about 4-5% lower, in North America and Japan, and about 13% lower in Europe, than those

of HFC-134a refrigerant/HFC foam units assuming refrigerant recovery at end-of-life disposal. Energy
14

use estimates for HC-based refrigerators are about 10% greater than that of the HFC units in all

regions. Use of HCs in refrigerators raises safety concerns and has resulted in higher unit costs.

Applying this cost differential to HFC designs (to incorporate vacuum panel insulation in the cabinet

walls, for instance) could yield a product with potentially superior TEWI characteristics.

1.3.1.2 Automobile Air Conditioners

The direct effect of refrigerant emissions for HFC based automobile air conditioners is a significant

part of the TEWI. The automobile manufacturers have responded with efforts to reduce charge size and

emissions. Research and laboratory development of air-conditioning systems based on both transcritical

CO2 compression and hydrocarbon refrigerants show a potential to reduce TEWI for this application.

Estimated TEWIs for CO2 and hydrocarbon systems are lower than those for HFC-134a systems in

regions with cool climates; TEWI are comparable to higher in climates with high cooling loads. Energy

consumption estimates for HFC systems are consistently lower than those of CO2 and hydrocarbon

systems. The long term performance, lifetimes, viability, and TEWIs of both the alternative systems

must be proven through extensive prototype and field trial testing. Energy consumption and TEWIs for

HC-based systems are negatively affected in all regions due to use of an indirect loop with attendant

efficiency penalties to keep the flammable refrigerant out of the passenger compartment.

1.3.1.3 Chillers

TEWI for this class of equipment has fallen significantly since the early 90's. New electric chillers

have 25% to 30% lower TEWIs than models of 4-5 years ago due to replacement of CFC refrigerants

with HCFC and HFC alternatives and to significant improvements in energy efficiency and reductions in

refrigerant loss rates. The choice of refrigerant makes only a minor difference in direct TEWI in new

equipment. Differences in chiller efficiencies for various refrigerant options can have a significant impact

on the indirect contribution to TEWI, however, which is dominant in this application.

Significant advances have been made in gas-fired chiller technologies. Triple-effect absorption

chillers now under development show potential for 25-30% reductions in TEWI compared to existing

double-effect machines. Engine-driven chillers are now available with rated efficiencies more than 25%

higher than the value used in the TEWI-II report. Estimated TEWIs for these machines are around 25%

lower than those for the triple-effect absorption equipment.

1.3.1.4 Unitary Equipment

Transition away from HCFC-22 in this equipment appears to be achievable with either no change

or a slight reduction in the estimated TEWIs. The HFC-400 blends, R-407C and R-410A, are the

principal alternatives being considered as HCFC-22 substitutes. Laboratory and limited field testing

indicates that R-407C has equivalent performance compared to R-22 while R-410A-based equipment

has potential for slightly better efficiency, lowering the indirect contribution to TEWI. Geothermal heat
15

pumps and premium grade air-to-air heat pumps can significantly reduce TEWI in this application,

albeit with higher purchase costs.

Gas engine-driven and gas-fired absorption heat pumps for space heating and cooling show

potential to reduce TEWI in climates dominated by heating requirements. Long term performance and

reliability of the gas-driven technologies have not been demonstrated.

1.3.1.5 Commercial Refrigeration

Supermarket refrigeration systems have had high direct contributions because of historically high

refrigerant charges and leakage rates. Equipment manufacturers have worked to reduce refrigerant

leaks at the display cases and in the brazed and welded joints in refrigerant lines. The lower emission

rates have reduced TEWI significantly from the values reported previously. New system design

concepts (secondary loop and distributed compressor approaches) also dramatically reduce the direct

effect of refrigerant emissions and result in lower overall TEWI estimates for this application. The

differences in TEWI between the HFC mixtures that have been considered are due primarily to the

GWPs of the refrigerants; the differences in energy use are not considered significant. Ammonia with

secondary heat transfer loops has been shown to be a viable alternative for HFCs in this application,

but there can be an energy penalty associated with necessary secondary heat exchangers. In many

areas system designs will have to comply with regulation and permit requirements intended to ensure

safe use in retail and commercial areas. Some European and developing countries have fewer

regulations and are more open to using ammonia. Refrigerant containment measures necessary for

ammonia and hydrocarbons could also be used with HFCs, resulting in essentially identical TEWI for

these alternatives.

1.4 SUMMARY

The concept of Total Equivalent Warming Impact, or TEWI, was developed as a comparative

index of the global warming impacts of an end use application by accounting for both the direct

contributions from refrigerants and blowing agents to the atmosphere and the indirect contributions from

energy consumption. It provides a useful tool in the assessment of various competing technologies that

could be used to substitute for current technologies as chemical compounds are phased out of use

under the Montreal Protocol. TEWI, however, is not the only criteria that must be considered. Safety,

health, other environmental concerns, initial system and operating costs, regional energy considerations,

ease of maintenance, and system reliability are some other factors that must be considered in evaluating

the “best technology” for any given application. As with any other composite index of performance,

estimated TEWI values are sensitive to an underlying set of key values and assumptions. Results from

these studies show that it is essential to analyze both the direct and indirect contributions of all air

conditioning and refrigeration alternatives to choose the most environmentally acceptable option. The

studies have also shown that energy efficiency and reduced refrigerant emissions are the most effective

means to mitigate future anthropogenic contributions to global climate change.
16
17

2. INTRODUCTION

The first study by AFEAS and DOE (Fischer, et al 1991) compared the global warming impacts of

alternatives to CFCs in refrigeration, air conditioning, foam insulation, and solvent cleaning applications

that could be commercialized before the year 2000. The refrigeration, air conditioning, and foam

insulation portions of that study were conducted by Oak Ridge National Laboratory (ORNL) and the

solvent cleaning portion by Arthur D. Little, Inc. That project developed the concept of total equivalent

warming impact (TEWI) which combined the effects of the direct emissions of refrigerants, foam

blowing agents, and solvents into the atmosphere with the indirect effects of energy consumption due to

combustion of fossil fuels and generation of electricity.

AFEAS and DOE supported a second phase study (Fischer, et al 1994) to evaluate the energy and

global warming impacts of not-in-kind (NIK) and next generation alternative technologies that could

possibly be developed in the long term to replace current commercial practices for vapor compression

refrigeration and air conditioning and foam insulations using CFCs and HCFCs. NIK alternatives

evaluated included thermoelectric refrigeration, evacuated panel insulation, and adsorption and

absorption technologies. Next generation technologies included chlorine-free fluorocarbon refrigerants

and blowing agents (HFCs), non-fluorocarbon refrigerants like ammonia and hydrocarbons, and

alternative compression technologies like acoustic compressors.

Both of the earlier studies examined end use applications in North America, Europe, and Japan and

included effects of cultural and technical differences in each region. These differences included such

things as the room and compartment temperatures for refrigerator/freezers, refrigerator size and

insulation thickness, annual driving distances for automobiles, fuels used for electricity generation, and

climate differences for building heating and cooling loads. In general, the principal results and

conclusions in the prior studies were the same without regard for the geographic region.

AFEAS and DOE decided to sponsor this third study for a number of reasons. Significant

developments have occurred in HFC and non-fluorocarbon refrigerants and blowing agents since

1993-94 which have not been considered in an independent evaluation of their global warming impacts. 

Data on thermal performance have become available that make it possible to do a meaningful evaluation

of the TEWI of these "new generation" fluids. A significant amount of work has been done on HFC

blends as alternatives to R-22 in both refrigeration and unitary space conditioning applications. Much

effort has also been devoted recently on use of non-fluorocarbons (eg, ammonia, hydrocarbons, and

CO2) as refrigerants and foam insulation blowing agents in selected applications. Analytical and

experimental results are available for quantitative comparisons between HFCs and non-fluorocarbons. 

In addition, new technologies for engine-driven compression systems and absorption systems for

residential and commercial space conditioning are being commercialized and operating data are

available. Thermal performance measurements have been made for refrigerator/freezer foam insulations

blown with "new generation" fluids.

The scope of the analyses in this report is limited to:
18

a) NIK or alternative technologies identified in the second study which were concluded to have a

reasonable potential for reducing TEWI and for which there are manufacturer development

activities underway which could lead to commercial products by the early years of the next

century; and

b) evolutionary improvements to existing technologies that have emerged since publication of the

phase II report.

The NIK concepts primarily considered were absorption systems for heating and air conditioning

residential and commercial buildings, the transcritical CO2 cycle for mobile air-conditioning, and

vacuum panel insulation for household refrigerators. Double-effect absorption equipment is already on

the market and more advanced cycles are under development with market entry expected before 2000. 

A number of automakers and auto suppliers (principally in Europe and Japan) are seriously investigating

the transcritical CO2 concept and are testing prototype systems. Hydrocarbon (HC) vapor

compression systems are also considered for household refrigerators, residential space conditioning

systems, and automobile air-conditioners.

End use applications considered in the present study are:

* household refrigeration;

* commercial chillers;

* automobile air conditioning;

* supermarket refrigeration; and

* unitary space conditioning.

TEWI computations in both the earlier studies and in the present study are made using the global

warming potentials (GWPs) for trace gases developed by the Intergovernmental Panel on Climate

Change (IPCC) based on the use of carbon dioxide as the reference (GWP of CO2 = 1). GWPs

based on a 500 year integration time horizon (ITH) were used in the first study while 100 year values

were used in the second study. The 100 year GWPs are about three times the 500 year values and,

thus, yield TEWIs with a higher relative direct contribution due to the refrigerant or blowing agent. The

long atmospheric lifetime of CO2 results in an effect on climate change that accumulates on a time scale

of centuries and, because the calculation of effects is truncated at the ITH, using GWPs at a short ITH

can seriously distort the energy component of TEWI. This is particularly problematic at the 20 year

ITH. More than 90% of the effect of CO2 lies beyond this time horizon and so values calculated using it

do not meet requirements of the Climate Change Convention for intergenerational equity. On the other

hand, despite their better description of the long term effects of CO2, the values calculated at a 500

year ITH are subject to the greatest uncertainty. Most of the international scientific community and
19

national and international policy makers concerned with global climate change appear to have settled on

the 100 year ITH as a reasonable compromise and so those GWP values are used here. GWPs

calculated using a 500 year ITH are approximately one third those at a 100 year ITH, simply due to

discounting the long term effect of CO2.

The TEWI estimates in this report also depend on assumptions made about carbon dioxide

emission from energy use, either from combustion of fossil fuels in automobile engines or furnaces or

indirectly from use of coal, natural gas, or oil for electricity generation. Electric power CO2 emission

rates were obtained from a number of electric utility and independent sources. These data are

summarized in Appendix A. Most of the TEWI estimates presented in the main body of the report

were generated using regional annual average emissions of 0.473 kg CO2/kWh delivered for Japan,

0.470 kg CO2/kWh for Europe, and 0.650 kg CO2/kWh for North America. Electricity transmission

and distribution loss effects are included. The effects of time-of-day variations in electricity generation

fuel mix and of fuel extraction and processing energy consumption were not considered in the main

body of this report. Appendix A includes a discussion of these factors.

Carbon dioxide emissions from fuel use for furnaces or gas-fired chillers and heat pumps are

assumed to be 53.0 g CO2/MJ input heat value for natural gas. This includes an average natural gas

transmission efficiency factor of 96.5%. This is an average for the years 1994-1996 obtained from

natural gas consumption data for the U. S. from the Energy Information Administration (EIA, 1997).

About 3.5% of the natural gas consumed (exclusive of extraction and processing consumption) was

used, primarily by compressors, to distribute the gas to consumers through the pipeline network.

Emissions from gasoline used to drive automobile air conditioners is assumed to be 2.32 kg

CO2/liter of fuel. This assumption is consistent with the data used in the first two studies.

A number of simplifying assumptions are implicit in the TEWI estimates provided herein. It would

be very difficult to compute absolute TEWI values for the full range of locations and alternatives

included in this report. Such an analysis would have to include location-specific values for electric

power plant carbon dioxide emissions (for some applications hourly emissions variations would have to

be considered together with hourly energy use simulations). Conclusions from detailed analyses like this

would be applicable only to the specific locations considered. This conflicts with a major element of the

project’s charter: to examine opportunities to reduce CO2 emissions that are broadly applicable across

regions. Thus, these detailed analyses would have to be repeated for numerous locations in each region

to fairly illustrate how comparisons between given technology options for each end use application

would vary throughout the regions. This is well beyond the scope of, and time available for, this study,

therefore most of the comparisons and conclusions are based upon regional annual average values for

CO2 emissions.

In the previous studies, the energies required to manufacture the greenhouse gases themselves were

not included nor have they been in this study. The climate change impacts of energy used for

production of a number of common refrigerants - HFCs, non-methane hydrocarbons, and ammonia -

have been evaluated directly by Campbell and McCulloch (1997). They show that the global warming

impact of the production energy is insignificant when compared to the potential impact on global

warming of releases of these gases.
20

Most of the value of the TEWI estimates included in this report come from using them as

comparative measures of the global warming impacts of different technology options for meeting a given

need in one of the end use applications. Uncertainties exist for all of the assumptions (many of which

are estimates or averages) that enter into the TEWI calculations. Uncertainties in the direct effect

include estimates of refrigerant emission rates and end-of-equipment-life refrigerant recover rates as

well as uncertainties in the GWPs themselves. The indirect contribution from CO2 emissions due to

energy use can be determined with less uncertainty (especially for specific locales where the fuel source

can be characterized accurately) but these values are still not precise. These uncertainties minimize the

importance of small differences in TEWI between similar, established technologies using the same

energy source (i.e. gas or electricity). When making such comparisons, if the TEWI differences are

small, the technology that shows lower energy use should be favored as long as safety and

environmental considerations are adequately addressed and costs are reasonable. In some cases, a

decision between options based on lowest energy consumption could result in choosing a technology

with a TEWI that is greater than the minimum possible. However, it can be argued that resource

conservation and particularly energy resource conservation should take precedence over marginal

global warming impact reductions when TEWI estimates are close.

In some cases, established technologies with known efficiencies are compared to systems under

development for which ultimate production efficiency levels and system configurations are less precisely

known. When making these comparisons, if the emerging technology exhibits a TEWI 10% or more

lower than that of the established alternative then further analysis and development may be justifiable on

this basis. TEWI, however, would not be the sole determining factor in the ultimate choice between the

alternatives

As in the previous two phases, this study involved experts from industry, government, and

academia around the world to characterize existing HCFC-based practices and to identify the operating

characteristics of alternative technology options. Therefore, much of the initial scoping work prior to the

start of the analyses may be viewed as a survey of the user industries concerning their candidate options

to replace HCFCs in the above applications. Unlike the first two studies, more interaction was

programmed through the course of the study with these experts. They were called on to provide

review and comments at three distinct points: an initial review of the assumptions and methodology

proposed for this round of analyses, a mid-term review of preliminary results and conclusions of the

analyses, and review of this final report itself. The cooperation of these experts has been vital to the

success of this study.
21

3. REFRIGERATOR-FREEZERS

3.1 INTRODUCTION

Residential refrigerator-freezers are one of the most familiar and inherently useful applications of

HCFC and HFC technology which touches our daily lives. Market saturation for domestic refrigerators

in the United States, Western Europe, and Japan is greater than 98% which is an indication of how

useful and necessary this appliance has become for a modern urbanized society. It is anticipated that an

explosive growth in the production and consumption of refrigerators will occur in developing countries

like China and India as these societies become more affluent (Deloitte & Touche 1996). The global

warming impact of a refrigerator-freezer includes the indirect emissions of CO2 resulting from the

combustion of fossil fuels to generate electricity to operate the refrigerator, and the direct global

warming contributions from the refrigerant and insulation foam blowing agents eventually released to the

atmosphere. Prior to 1993, CFC-12 was used as the refrigerant in this application because of its

excellent efficiency, safety, and compatibility characteristics which had been fully developed over 50

years of use. The phaseout of CFC-12 in industrialized countries as a result of the Montreal Protocol

has prompted the development of HFC-134a and hydrocarbons like iso-butane (HC-600a) and

propane (HC-290) as chlorine-free, ozone-safe, alternative refrigerants for CFC-12.

Rigid polyurethane insulating foams blown with CFC-11 were the dominant insulation used in the

cabinets of refrigerators prior to the ban on production and sale of CFCs. After the effective date for

implementation of the Montreal Protocol, CFC-11 blown insulating foam was rapidly eliminated from

refrigerator-freezer production throughout the world. Germany, northern Europe, Australia, and New

Zealand are converting to cyclopentane as an appliance foam blowing agent, with some other

European manufacturers using HFC-134a, HCFC-141b, and blends of HCFC-142b or HCFC-141b

with HCFC-22. U.S. and Japanese manufacturers have predominantly converted to HCFC-141b as

an insulation blowing agent for refrigerators because of its excellent insulating characteristics, nonflammability, and low volatile organic carbon (VOC) emission ranking. 

3.2 ENERGY EFFICIENCY

Domestic refrigerator-freezers, used primarily for food preservation, are an important and major

user of electricity. Approximately 62 million new refrigerators are manufactured world-wide each year

primarily to serve as replacements in saturated markets for the hundreds of millions of units already in

use (UNEP 1995). Energy efficient domestic refrigerators contribute substantially to world-wide efforts

to reduce energy consumption and global warming. Efforts to decrease the energy used by refrigerators

have focused primarily on improving the efficiency of the mechanical refrigeration process through more

efficient compressor designs and improvements in refrigerant-to-air heat transfer. Improved cabinet

insulation which results in a smaller refrigeration load has also significantly decreased the energy
22

Figure 2. Historical and projected refrigerator-freezer

energy improvements for the United States.

1975 1980 1985 1990 1995 2000 2005

200

400

600

800

1000

1200

1400

1600

1800

2000

Year

Electricity Use (kWh/year)

1990

Federal

Standards

1993

Federal

Standards

2003

Next

Round

of Standards

consumption of this appliance (Vineyard

1995). Government mandated minimum

energy-efficiency standards like those

established by the United States under the

National Appliance Energy Consumption

Act have been effective in motivating the

adoption of energy saving technologies

(NAECA 1987). A chart showing the

actual and projected improvements in

electrical energy use for U.S. refrigerators

is shown in Fig. 2. The initial standards

went into effect on January 1, 1990, and

have had one revision in 1993, which

resulted in a 25% reduction in energy

consumption. The next revision is expected

to require about an additional 30%

reduction. This revision was originally

scheduled for 1998 but has been

rescheduled to 2001. 

The efficient use of energy is important

because of its limited availability and also

because of international efforts to reduce

the atmospheric concentrations of

greenhouse warming gases. For refrigerators, which have very small refrigerant charges and perhaps the

most leak-free hermetic refrigeration system, the indirect contribution to TEWI is about an order of

magnitude greater than the direct contribution of the CFC- alternative refrigerants. So, selection of the

most efficient refrigeration fluids for this application, regardless of their individual GWPs, is critical to

minimizing its overall global warming impact.

3.3 REFRIGERANTS

3.3.1 HFC-134a

HFC-134a was initially the refrigerant of choice to replace CFC-12 in refrigerator-freezers

because it is chlorine-free (zero ODP), non-flammable, non-toxic and has comparable cycle efficiency.

Iso-butane, propane, and mixtures of propane and iso-butane were subsequently suggested as

refrigerant alternatives, especially by the Northern Europeans, because they are comparable in

efficiency to HFC-134a, miscible with conventional refrigeration oils and have much lower GWPs. The

boiling point, vapor pressure characteristics, and refrigeration cycle efficiency of HFC-134a are quite

similar to CFC-12, but its volumetric capacity is about 12% less, and it is insoluble in the mineral and
23

Figure 3.. Measured energy use for European 3-star

refrigerator/freezers using CFC alternatives.

100 150 200 250 300

0.20

0.30

0.40

0.50

0.60

0.70

0.80

0.20

0.30

0.40

0.50

0.60

0.70

0.80

Effective Volume (liter)

Energy Use (kWh/day/100 liter)

Pentane / Butane

Pentane / HFC-134a

Blowing Agent / Refrigerant

alkylbenzene oils previously used with CFC-12. Despite a slightly smaller ideal thermodynamic COP

for the domestic refrigeration cycle, extensive energy consumption testing in refrigerator-freezers has

shown that HFC-134a can achieve approximately equal performance to CFC-12 in optimized units

(Radermacher 1996). Synthetic polyolester (POE) oils are usually used in hermetic systems with HCF134a. POE oils can hydrolyze in the presence of high levels of moisture to produce system damaging

acids, therefore additional process controls are required to ensure the maintenance of low system

moisture levels. This new refrigerant/lubricant combination has also forced re-evaluation of

manufacturing and processing fluids used for production of compressors and heat exchanger tubing to

avoid selective solubility and sludge formation that can result in capillary tube plugging (Kruse 1996).

3.3.2 Hydrocarbon Refrigerants

Hydrocarbon refrigerants are flammable which implies safety issues in the manufacture and home

use of units charged with these refrigerants. Product liability concerns preclude the use of hydrocarbon

refrigerants in the larger frost-free designs that are popular in the United States and Japan. The

flammability of hydrocarbons causes increased investments in production factories for additional safety

precautions and in the form of several design changes required for certain models to reposition electrical

switches or change to explosion proof components needed to avoid potential accidents. German

manufacturers have demonstrated that hydrocarbon refrigerants can be handled safely in the

manufacturing process and that European customers are willing to accept products using hydrocarbons.

Normalized energy consumption

data presented in the phase II

DOE/AFEAS report for 3-star

refrigerators produced in

Europe, Fig. 3, and subsequent

testing indicate that roughly

equivalent efficiencies can be

obtained with HC-600a and

HFC-134a refrigerants 

(Fischer 1994, Deloitte &

Touche 1996, Wenning 1996).

In Fig. 3, the solid triangles are

measurements for refrigerators

using cyclopentane as the foam

blowing agent and butane as the

refrigerant. The open circles

represent measurements for cyclopentane as the blowing agent

and HFC-134a as the refrigerant.

The broad scatter makes it

impossible to draw any definitive
24

conclusions about any one combination being more efficient than the other and indicate that equivalent

efficiencies can be obtained with either refrigerant. 

In addition to the favorable GWP of hydrocarbon refrigerants, they are also compatible with the

materials, lubricants, and manufacturing/processing fluids previously used with CFC-12 refrigerators.

The latent heat of vaporization of hydrocarbons is large in comparison to halogenated alkanes and their

low density results in smaller charge requirements. Initial hydrocarbon applications used HC-290 /HC600a (propane/iso-butane, 50/50 mass%) blends to roughly simulate the volumetric capacity and allow

the use of compressors designed for CFC-12 or HFC-134a. Use of this blended formulation has

decreased in favor of using HC-600a alone which has a volumetric capacity .80% lower than HFC134a thereby requiring a larger displacement compressor to obtain the same refrigeration capacity. 

Compressors with larger displacement can be an advantage since compressor efficiency falls off in the

smaller capacity ranges due to manufacturing tolerance limitations.

3.4 Blowing Agents

Heat leakage through the insulated cabinet of a refrigerator is the main source of the refrigeration

load so improvements in the thermal conductivity of insulating foams used in the cabinet and the heat

integrity of other cabinet components significantly improve the energy efficiency of the product (Sand

1994, Vineyard 1995, UNEP 1995). The rigid strength of the foam and its ability to adhere to cabinet

walls and liners is also important because it constitutes an integral element of the cabinet structure and is

essential to maintain dimensional stability for the appliance. The material compatibility of blowing agents

with the polymer liner materials used for interior surfaces of the refrigerator is another factor which must

be taken into consideration when an alternative blowing agent is being evaluated. Care must be taken to

be sure that the insulating, structural, compatibility, and processing characteristics are not severely

degraded by the choice of alternative blowing agents.

 

3.4.1 HCFC-141b and HFC Alternatives

HCFC-141b was chosen as a successor to CFC-11 for U.S. and Japanese refrigerators after

several years of experimentation with various blowing agent options. Its use as an alternative blowing

agent represents a greater than 9 to 1 reduction in the ODP and a 6 to 1 reduction in the GWP over

CFC-11. HCFC-141b blown foams have excellent insulating values and equipment previously used for

CFC-11 was easily converted for use with this alternative. Early material incompatibility problems of

these foams with plastic door liners were solved by using co-extruded barrier on these liners.

In anticipation of the 2003 phaseout of HCFC-141b, appliance manufacturers through their

international associations and members of the Polyisocyanurate Foam Manufacturers Association

(PIMA), have generated and published a great deal of experimental information on the performance of

insulating foams blown with chlorine-free alternative blowing agents (Haworth 1996) (Polyurethane

1995) (Polyurethane 1996). Alternative blowing agents evaluated in the United States and Japan for

HCFC-141b include: HFC-134a, CO2 (water- based), HFC-356mffm, HFC-365mfc, HFC-245fa, 
25

HFC-236ea, and cyclopentane. The Association of Home Appliance Manufacturers (AHAM)

Insulation Technical Advisory Committee (ITAC) and member companies have measured the relative

energy use of 22 ft3, top-mount, refrigerators blown with HCFC-141b and all of these alternatives

using the Department of Energy, 90E F, closed-door test, commonly known as the standard DOE test. 

Parallel cabinet energy consumption tests for the blowing agents listed above and for HFC-152a and

HFC-245ca in Japanese refrigerators were reported by The Japanese Electrical Manufacturers

Association (JEMA).

3.4.2 Cyclopentane and Other Hydrocarbons

Cyclopentane was introduced in 1993 as an alternative blowing agent for appliance foams in

German refrigerators. Cyclopentane and other hydrocarbon blowing agents have zero ODPs and

insignificant GWPs, but foams blown with these chemicals have higher insulation thermal conductivities

compared to HCFC-141b which results in higher energy use for refrigerator-freezers. Foams blown

with cyclopentane have no incompatibilities with other materials used in the refrigerator cabinet. These

blowing agents would contribute to the volatile organic compound (VOC) emission inventory of

industries that are controlled under the Clean Air Act. 

Since hydrocarbons like cyclopentane are flammable, investment costs for changing a production

line to HC use and implementing the manufacturing and production controls to ensure safe operation are

higher than those associated with non-flammable technologies. After the manufacturing process, foams

blown with HCFC-141b and cyclopentane have been reported to have similar flammability

characteristics.

Conflicting data has been reported for the relative aging characteristics of HC versus HCFC or

HFC blown insulating foams (Christian 1997, Heilig 1995). 

3.4.3 Vacuum Panels

Vacuum insulated panels are the leading alternative technology to blown polyurethane foam for

improving the cabinet insulation of refrigerators. These panels can be formed by using air and moisture

tight metal or plastic barriers to contain a low conductivity filler material such as diatomaceous earth,

aerogels, or open cell polyurethane foam. Selection of materials by appliance manufacturers will depend

on material and panel manufacturing costs, weight, and proven performance. For this study, panels are

assumed to be used in the door, each wall, and the top of the freezer cabinet, centered in each surface.

Panels are assumed to extend to within 4 cm (1.57 in.) of each edge resulting in approximately 80%

surface coverage for the freezer. 

The annual energy use for a U.S. refrigerator with composite vacuum panels covering 80% of the

exterior surface area of the freezer compartment was taken from the actual measured performance of a

20 ft3, top-mount refrigerator tested at ORNL. The results indicate about an 8.5% performance

improvement measured for a cabinet with evacuated panel enhancements to the freezer over a baseline

cabinet prepared without panels (Vineyard 1997). TEWI calculations for U.S. refrigerators with

composite wall construction are based on constant overall wall thickness; the combined thickness of the
26

vacuum panel and surrounding insulation for each wall of the refrigerator is the same as the foam

thickness in the simulations for refrigerators without vacuum panels. One of the advantages of vacuum

panel insulation is that thinner walls could be used while maintaining or improving the cabinet

performance and also increasing usable internal volume; this approach was not considered in this study.

The actual performance of evacuated panel insulation is very strongly affected by the dimensions of the

panels because of the edge effects. 

Evacuated panels may be a viable option to be considered in the future along with other energy

saving design changes (e.g. higher efficiency compressors, better door gaskets). Evacuated panels

could be implemented on a commercial scale if they can be made cost-effective and have acceptable

product lifetimes.

3.5 ALTERNATIVE TECHNOLOGIES

3.5.1 Absorption Refrigeration

More than a million thermally activated ammonia/water absorption refrigerators are manufactured

and sold annually in world-wide markets. These products are commonly used in recreational vehicles

because of their ability to use bottled gas as an energy source and in hotel rooms because of their quiet

operation. These products are limited in size because of constraints imposed by the reliance on present

day bubble pumps in their operating cycle and cannot be scaled up to displace conventional vaporcompression products. Recently reported research results indicate that better utilization of waste heat

from the refrigerator rectifier and improvement in the design of the generator can improve the thermal

COP of these refrigerators by as much as 50% (from a COP of 0.2 to 0.3) without degrading cooling

capacity (Chen 1996). While the unique capabilities and operating characteristics will continue to allow

absorption refrigerators to fill niche markets, low efficiencies and inherent design limitations make it very

unlikely that they will be broadly adopted as replacements for vapor compression refrigerators.

3.5.2 Thermoelectric Refrigeration

Data for commercially available thermoelectric modules indicate they have a COP of approximately

0.32 at the same operating conditions as those used to rate compressors for refrigerator-freezers

(Fischer 1994). Some commercial products like portable coolers and small drink cooler cabinets where

portability and convenience of operation are important considerations use thermoelectric refrigeration

systems. But this alternative technology is unlikely to replace vapor compression refrigeration for the

broader domestic refrigerator market because of its extremely low efficiency.

No TEWI results were calculated for absorption refrigerators and thermoelectric refrigerators in

this report.
27

Parameter Value

internal volume 510 liters

foam volume 0.267 m3

refrigerant charge 0.155 kg HFC-134a

annual energy use 646 kWh/year (AHAM 1996)

blown foam insulation 1.08 kg HCFC-141b

energy use/unit internal volume 1.27 kWh/y/liter

Table 1. Baseline North American refrigerator/freezer.

Parameter Value

internal volume 230 liters

foam volume 0.1539 m3

refrigerant charge

 HFC-134a

 HC-600a

0.127 kg

0.045 kg

annual energy use

 HCFC-141b foam

 cyclopentane/pentane foam

300 kWh/year

328 kWh/year

blown foam insulation

 HCFC-141b foam

 cyclopentane/pentane foam

0.520 kg

0.279 kg

energy use/unit internal volume

 HCFC-141b foam

 cyclopentane/pentane foam

1.30 kWh/y/liter

1.42 kWh/y/liter

Table 2. Baseline European refrigerator/freezer.

3.6 ASSUMPTIONS

Refrigerants HFC-134a and iso-butane (HC-600a) or mixtures of hydrocarbons were assumed to

give equivalent performance in the refrigeration circuit for all regions. Measurements of actual

compressor performance with the two refrigerants need to be used to make a more definitive

comparison.

3.6.1 North American Refrigerator-Freezer

TEWI calculations for a North

American refrigerator-freezer

design are based on a 510 liter (18

ft3) frost-free (automatic defrost)

model. TEWI calculations used

refrigerator energy consumption

data at the standard U.S. rating

test; 32.2EC (90EF) room

temperature, -15EC (5EF) freezer

temperature, and 3.3EC (38EF)

fresh food temperature as reported

in the 1996 Directory of Certified

Refrigerators and Freezers

(AHAM 1996). The freezer door

and wall thicknesses are assumed to be

6.00 cm (2.36 in.) and 7.50 cm (2.93 in.),

respectively, while the door and wall

thicknesses for the fresh food section are

assumed to be 4.00 cm (1.77 in.) and 6.00

cm (2.36 in.). The energy use for the

baseline refrigerator/freezer using HCFC141b blown foam and HFC-134a for the

refrigerant is 646 kWh/year (1.77

kWh/day). These data are summarized in

Table 1.

3.6.2 European Refrigerator/Freezer

The typical European refrigerator

chosen for this analysis is assumed to be a

230 liter (8 ft3) model with a small freezer

compartment. Reported energy use from
28

Parameter Value

internal volume 355 liters

foam volume 0.1796 m3

refrigerant charge 0.140 kg HFC-134a

annual energy use 522 kWh/year

blown foam insulation 0.71 kg HCFC-141b

energy use/unit internal volume 1.47 kWh/y/liter

Table 3. Baseline Japanese refrigerator/freezer.

European manufacturers at rating conditions was used as a starting point for TEWI calculations. Typical

temperatures for these tests are set at 25EC (77EF), -18EC (-0.4EF), and 5EC (41EF) for the room,

freezer compartment, and fresh food compartment. Freezer door and wall thicknesses for this

refrigerator would typically be 6.75 cm (2.66 in.), while the door and wall thicknesses for the fresh

food section are about 3.50 cm (1.38 in.) and 4.50 cm (1.77 in.). TEWI calculations were performed

for cyclopentane, HCFC-141b, HFC-245fa, and HFC-134a blown foams and HFC-134a or isobutane (HC-600a) as refrigerants. The energy use for the baseline refrigerator-freezer using HFC-134a

or HC-600a as the refrigerant and HCFC-141b blowing agent is taken as 300 kWh/year (0.82

kWh/day) or 1.31 kWh/y/l. The energy use for a baseline European refrigerator-freezer with either

HFC-134a or HC-600a as the refrigerant and cyclopentane as the insulating foam blowing agent is

assumed to be 328 kWh/year (0.90 kWh/day) or 1.42 kW/y/l. This performance difference between

the HCFC-141b and cyclopentane cabinets is the result of the increased thermal conductivity of

cyclopentane and assumed equivalent wall thickness of the appliance insulating foam. These data are

summarized in Table 2.

3.6.3 Japanese Refrigerator-Freezer

The baseline Japanese refrigerator use for this TEWI evaluation is assumed to be a 355 liter (12.5

ft3) automatic defrost model with an annual energy use of 522 kWh/year (Hara 1996a). The cabinet is

divided into three compartments; one for frozen foods, one for cold storage, and the third for produce

and vegetables (approximately 25%, 50%, and 25% of the refrigerated volume, respectively). The

cabinet is foamed with 47 mm (1.85 in.)

of HCFC-141b blown foam in the

freezer walls and door. The walls and

door of the fresh food compartment have

35 mm (1.38 in.) of insulation (same as

used for the TEWI-1 report). The annual

energy consumption given above is

based on a refrigerator-freezer using

HCFC-141b blown foam and 0.14 kg

of HFC-134a as the refrigerant. Data

for the Japanese refrigerator are

summarized in Table 3.

3.7 METHODOLOGY

Rather than relying on a refrigerator computer model to generate annual energy use, the foregoing

baseline values for “typical” U.S., European, and Japanese refrigerators at operating conditions

reflecting each countries’ rating and certification standard conditions were chosen as the starting point

for this analysis. These rating point energy consumption values were assumed to be representative of
29

field operating conditions for each location. Alan Meier has shown that energy use values generated

using the standard DOE test for U.S. refrigerator-freezers was a reasonably good predictor of field

energy use (Meier 1993). An indication of how the same refrigerator would perform if tested according

to the standard rating procedures for different countries can be estimated from results presented in a

recently published comparison (Bansal 1995). 

The relative energy consumption of refrigerator cabinets prepared using insulating foams blown with

alternatives to HCFC-141b or cyclopentane was taken whenever possible from research results

presented by the AHAM Appliance Research Consortium (ARC) Insulation Technical Advisory

Committee (ITAC) at recent conferences and the parallel Japan Electrical Manufacturers Association

(JEMA) Home Appliances Department program (Haworth 1996, Araki 1996). No internally

consistent energy consumption data for the various combinations of refrigerants and blowing agents was

found for the three- and four-star refrigerators marketed in Europe. “Baseline” energy consumption

results were averaged from data provided by S. Brasch, M. Verhille, and H. Lotz (1996). Baseline

energy consumption values for “typical” refrigerators in North America, Europe, and Japan were scaled

in proportion to measured energy consumption increases or decreases published in the AHAM or

JEMA reports for cabinets prepared with insulating foams blown with alternatives to HCFC-141b or

cyclopentane. This adjusted annual energy use , the appliance lifetime , and the kg CO2/kWh electric

power plant emission rate were used to calculate the indirect TEWI for each refrigerant/blowing agent

combination.

3.8 DISCUSSION

No attempt has been made to evaluate or balance all the advantages or disadvantages of the HFCand hydrocarbon-based refrigerators. One has an impact from direct emissions of blowing agent and

refrigerant, the other uses flammable materials and has a higher energy consumption but no direct

impact on TEWI. Small differences in TEWI between the fluorocarbon- and hydrocarbon-based

refrigerators must be weighed against the risks associated with the hydrocarbons. Japanese and

American risk analyses have questioned the safety of large, automatic defrost refrigerators using hydrocarbons (Deloitte & Touche 1996). 

The actual performance of the appliance will depend on the thermal properties of the insulation at

the operating temperatures. Generally the insulating ability of the foams increases at lower

temperatures, however there is a range where the thermal conductivity goes up for foams blown with

low boiling liquid blowing agents (e.g. HCFC-141b, cyclopentane). As the temperature goes down,

these blowing agents begin to condense in the cells of the foam, reducing the percentage of insulating

gas in the cells and increasing the overall thermal conductivity. This phenomenon does not occur with

foams formed with gaseous blowing agents (e.g., HFC-134a, HCFC-22/HCFC-142b) over the range

of temperatures relevant to a household refrigerator/freezer. This variation of k-factor with temperature

has not been taken into account in any of the calculations in this study.

Where IPCC listed GWPs were not available for some of the newly developed blowing agents,

values estimated by AFEAS member companies were used to calculate the direct TEWI. Considerable
30

Blowing Agent / Refrigerant

North America Europe Japan

HFC-134a HC-600a HFC-134a HC-600a HFC-134a HC-600a

HCFC-141b 1.77*1.77 0.82*0.82 1.43* NC

HCFC-22/HCFC-142b (60/40) 1.86 1.86 NC NC 1.43 NC

HFC-134a 2.01 2.01 0.85 0.85 NC NC

HFC-245fa 1.79 1.79 0.82 0.82 1.43 NC

HFC-365mfc 1.93 1.93 NC NC NC NC

HFC-356mffm 1.84 1.84 NC NC 1.48 NC

HFC-236ea 1.85 1.85 NC NC 1.49 NC

Cyclopentane 1.95 1.95 0.90 0.90*1.57 1.57

CO2 1.98 1.98 NC NC NC NC

HFC-152a NC NC NC NC 1.67 NC

HFC-245ca NC NC NC NC 1.50 NC

Vacuum Panels 1.62 1.62 NC NC NC NC

*

 baseline values

Table 5. Calculated energy use for refrigerator/freezers using alternative refrigerants

 and blowing agents (kWh/day).

controversy exits on how accurate GWP values are calculated. IPCC in their 1995 report chose to use

compound atmospheric lifetimes corresponding to tropospheric degradation mechanisms alone. When

both tropospheric and stratospheric destruction mechanisms are considered, overall lifetimes and the

resulting GWP value for the chemicals in question are considerably decreased. To be consistent, GWPs

chosen for these calculations were taken from values listed in the 1995 IPCC report. Whenever values

were not available in this listing, AFEAS estimates were used. Tabulated values for refrigerant and

blowing agent GWP are given in Tables 23 and 24 in Appendix B.

3.9 RESULTS AND CONCLUSIONS

Detailed tabulations of the TEWI calculations for refrigerator-freezers are given in Appendix C for

all regions. Key results are summarized below.

3.9.1 North American Refrigerator-Freezer

Energy use for North American refrigerator-freezers with various blowing agents are listed in Table

4. These results and data for foam density, weight percent blowing agent, and refrigerator dimensions 
HCFC-141b

Vacuum Panels

HCFC-22/HCFC-142b

HFC-134a

HFC-365mfc

HFC-356mffm

HFC-245fa

HCFC-22/HCFC-141b

HFC-236ea

Cyclopentane Isomers

Carbon Dioxide

Cyclopentane Isomers

Carbon Dioxide

0 2000 4000 6000 8000 10000 12000

Blowing Agents

TEWI (kg CO )

Indirect Effect

Energy

Direct Effect

Blowing Agent

Direct Effect

Refrigerant

Refrigerant:

HFC-134a

Refrigerant:

Isobutane

2

31

Figure 4. TEWI results for a 510 liter (18 ft ) North American, top-mount refrigerator/freezer 3

insulating foam options.

were used to compute the TEWI for each combination of insulation and refrigerants shown in Fig. 4.

TEWI are calculated using 100 year GWPs, a 15 year service life for the appliance and an average

CO emissions rate of 0.65 kg CO /kWh. The refrigerator using HCFC-141b for cabinet foam shows 2 2

12% of the TEWI due to fluorocarbon emissions. HFC blowing agent alternatives for HCFC-141b

show from 11 to 15 % of TEWI due to chemical emissions. As an alternative to HCFC-141b,

refrigerator foams blown with HFC-245fa give the best balance of energy efficiency and lowest TEWI.

With the exception of cyclopentane base insulations, almost all of the direct effect from refrigerators is

due to the foam blowing agent. Assumptions on mandatory recovery of the refrigerant would result in

marginal, 2 to 3% reductions in the TEWI.

An 8 to 9% improvement in TEWI is indicated for a refrigerator with vacuum panels

incorporated into the freezer walls. Essentially all of this improvement is due to a reduction in the

indirect TEWI contribution resulting from decreased energy use. A cost/benefit analysis performed on

this vacuum panel freezer enhancement indicated a 36 year consumer pay-back period and concluded

that other readily available energy saving options were more promising (Vineyard 1997).

Essentially no interest remains in using HCFC blends like HCFC-22/HCFC-142b or

HCFC-22/HCFC-141b as blowing agents because of the impending phaseout dates for chlorine

containing refrigerants and blowing agents. Despite low GWP values, the use of CO2 alone as a

blowing agent for appliance foams is not considered viable because the high initial thermal conductivity

and rapid aging of foams blown with this gas (Deloitte & Touche 1996).
Cyclopentane

HCFC-141b

HFC-245fa

HFC-134a

Cyclopentane

HCFC-141b

HFC-245fa

HC-134a

0 1000 2000 3000 4000

Blowing Agents

Indirect Effect

Energy

Direct Effect

Blowing Agent

Direct Effect

Refrigerant

Refrigerant:

HFC-134a

Refrigerant:

Isobutane

TEWI (kg CO 2)

32

Figure 5. TEWI results for a 230 liter European refrigerator/freezer versus cabinet insulating foam

options.

3.9.2 European Refrigerator-Freezer

Baseline and projected daily energy use for the European style refrigerator/freezer are also

contained in Table 4. All refrigerators with cabinet insulation blown with cyclopentane or pentane

isomers are estimated to use 9 to 10% more energy than similar models using HCFC-141b as the foam

blowing agent (Haworth 1996, Deloitte & Touche 1996). Use of HFC-134a as a blowing agent in

Europe is decreasing because of poor thermal conductivity results and the availability of other blowing

agents (Deloitte & Touche 1996). 

TEWI results shown in Fig. 5 for European refrigerators are calculated using a 15 year service

life for the appliance and CO2 emission rate for electricity generation of 0.470 kg CO2/kWh. These

results show the same trends as those in Fig. 4 for the North American refrigerator/freezer, although the

percentage of direct effects due to fluorocarbon emissions is higher (# 30%), primarily because of the

lower CO2 emissions for the power generation mix in Europe (higher percentage of nuclear and

hydroelectric power generation with virtually zero CO emissions). The lower rate of CO emissions for 2 2

power generation results in a relatively smaller indirect effect, a smaller overall TEWI, and hence a

larger percentage direct effect than the North American refrigerators. Refrigerant recovery at end of life

would reduce TEWI’s for the HFC-134a options by 5-6%, making them equal to those of the HC600a refrigerators.
HCFC-141b

HCFC-22/HCFC-142b

HFC-152a

HFC-236ea

HFC-245ca

HFC-245fa

HFC-356mffm

Pentane Isomers

Pentane Isomers

0 1000 2000 3000 4000 5000 6000 7000

Blowing Agents

Indirect Effect

Energy

Direct Effect

Blowing Agent

Direct Effect

Refrigerant

Refrigerant:

HFC-134a

Refrigerant:

Isobutane

TEWI (kg CO 2 )

33

Figure 6. TEWI results for a 355 liter Japanese refrigerator/freezer vs. cabinet insulating foam options.

3.9.3 Japanese Refrigerator-Freezer

Projected daily energy use for a 355 liter (12.5 ft ) Japanese refrigerator-freezer is shown with 3

the North American and European refrigerators in Table 4. TEWI results presented in Fig. 6 are

calculated using a 15 year service life for the appliance and a 0.473 kilogram of CO per kilowatt-hour 2

power plant emission rate for electricity in Japan. No recovery of the blowing agent or refrigerant is

assumed. The relative performance of cyclopentane blown appliance foams for these refrigerators was

extrapolated from the AHAM ITAC energy consumption results for North American refrigerators and

confirmed with JEMA representatives, with the exception of the HCFC-22/HCFC-142b blown foam

(Hara 1996b). The direct TEWI contribution from fluorocarbon blowing agents and refrigerant are

about 5 to 15% of the total TEWI (Fig. 6).

Despite the equivalent insulating characteristics reported by JEMA for HCFC-141b foam and

foams blown with a 60/40 mass% HCFC-22/HCFC-142b mixture, the larger GWP of the HCFC

blend results in an 18% greater overall TEWI. The low GWP for HFC-152a is not enough to

overcome the poor performance of foams blown with this fluorocarbon, so no TEWI advantage is seen

over refrigerators using foams blown with HCFC-141b, HFC-236ea, HFC-245ca, HFC-245fa, or

HFC-356mffm which have larger GWPs. 

No clear TEWI advantage is shown for refrigerator-freezers using HC-600a as the refrigerant

and cyclopentane or hydrocarbon blown insulations because of the diminished thermo-insulating
34

No clear TEWI advantage is shown for refrigerator-freezers using HC-600a as the refrigerant and

cyclopentane or hydrocarbon blown insulations because of the diminished thermo-insulating

performance seen with these foams and the corresponding increase in indirect TEWI.

Assumptions on recovery of the refrigerant would result in a 3-4% reduction in the TEWI, which

would essentially equate the calculated TEWIs of the HFC and HC refrigerators. 

Basic research on recovery of the blowing agent is being conducted in Japan. If the technology is

established and foam blowing agent is partially recovered, the direct TEWI of refrigerators with HCFC

or HFC blown insulating foams could decrease by . 10%, but indirect TEWI would be increased in

proportion to the amount and type of energy utilized in this recovery process.

3.10 CONCLUSIONS

Much of the R&D effort associated with improving the environmental acceptability of refrigeratorfreezers has been directed at improving the efficiency of the vapor compression refrigeration cycle

through more efficient circuit components (Vineyard 1997). No alternative cooling technology appears

ready to displace vapor compression for domestic refrigeration. Vacuum panels can be used to

improve cabinet insulation and decrease TEWI but current costs make this approach unattractive

compared to other options. The use of HC refrigerants and foams in refrigerators and freezers as

compared to HFC-134a refrigerant with HFC foams results in estimated TEWI reductions of about 6%

in North America, 9% in Japan, and 18% in Europe under the baseline assumptions used in this

analysis. If the HFC-134a is assumed to be fully recovered at disposal of the refrigerator, TEWI

reductions for the HC option become about 4% in North America, 5% in Japan, and 13% in Europe.

Use of HCs in these products also raises additional safety concerns and has resulted in higher unit costs

(Deloitte and Touche, 1996).

Consistent improvements in refrigerator efficiencies lead to lower TEWIs but accentuate the direct

contribution. Energy consumption estimates for the HFC-based refrigerator alternatives average about

10% lower than those of the HC alternatives in all regions.

A concerted research effort has been directed to finding an alternative blowing agent for HCFC141b which contains chlorine and will be phased out of use in 2003. Some general conclusions relating

to refrigerator-freezer applications that can be drawn from this effort are listed below:

! Several of the alternative HFC blowing agents being considered give insulating foams comparable

to HCFC-141b, but have higher GWPs than HCFC-141b which result in larger direct TEWIs

relative to the HCFC-141b baseline. The alternative HFC blown foams have increased energy

efficiency compared to hydrocarbon blown foams which gives a lower indirect contribution to

TEWI and offsets some of the differences in direct effects.

! HCFC-141b and all the alternative HFC blowing agents with the exception of HFC-134a have

produced appliance insulating foams with better k-values than HC blown foams.
35

! In this comparison, insulating foams resulting in refrigerators with better energy efficiencies may be

favored because decreased electrical energy use has other long range environmental implications.

Uncertainties, as discussed in the executive summary, are embedded in the parameters that factor

into these calculations. Further, as stated in the opening paragraphs of the executive summary, TEWI

should not be considered the sole determining factor in deciding between technology options for this or

any other application. Not withstanding it must be noted that the 13-18% estimated TEWI reduction

from HC use in refrigerators for Europe should be considered significant. It must also be noted that the

lesser HC TEWI advantage estimated for the North American and Japanese cases is not considered to

be significant. Substantial added costs would be needed for production safety and environmental

concerns at the factories and for reducing/eliminating spark sources in the products themselves to

produce a frost-free refrigerator (typical of the major type used in the latter two regions) using

hydrocarbons. These higher costs could be applied to HFC designs (to incorporate vacuum insulation

in the cabinet walls, for instance) to yield a product with potentially superior TEWI.
37

Figure 7. Efficiency ratings of unitary air conditioners in the

United States (sales weighted average).

1975 1980 1985 1990 1995

0

5

10

15

YEAR

EER/SEER (Btu/kWh)

EER SEER

UNITARY AIR CONDITIONER

NAECA efficiency

standards

4. UNITARY AIR CONDITIONING

4.1 INTRODUCTION

Estimates of the installed unitary air conditioning capacity worldwide range from 763x103 kW

(217x103 tons) (Fischer 1991) to 1,450x103 kW (410x103 tons) (UNEP 1995) exceed the worldwide

installed chiller capacity by at least a factor of four. Over 90% of this unitary equipment is in North

America and Japan, although air conditioning is becoming more common in Europe and other countries.

By its broadest definition, the unitary term encompasses all equipment that cools or heats building air

via direct heat exchange with a refrigerant coil or indirect heat exchange through a hydronic loop. It

applies to air conditioners which only cool

and heat pumps which can only heat or

heat and cool. The refrigeration capacity of

unitary systems can range from 2 to 420

kW (0.6 to 120 refrigeration tons). As

indicated in Fig. 7, the efficiency of U.S.

central air conditioning products has

steadily increased over the last twenty

years with slight spurts in performance

brought on by governmental regulation and

market competition (ARI 1996). 

Improvements in performance and

reliability of this product over the years

have not been matched by an increase in

constant dollar cost to the consumer

because the market for unitary equipment

is highly competitive. Alternative

refrigerants for this market segment or

alternative technologies must deliver

equivalent or better performance than

current systems with similar or lower costs

in order to compete effectively with existing

technologies.

4.2 TECHNOLOGY DESCRIPTION

Ducted air-to-air residential systems using HCFC-22 constitute most of the North American

unitary market. A compressor and heat exchanger located outside of the building supply refrigerant to

a single indoor coil. Building air is cooled or heated by the indoor coil and distributed throughout the
38

building via ducts. UNEP estimates that 59 million ducted, air conditioning systems are installed in

North America and world wide. Due to differences in building practices, ductless split systems are

more popular in Japan and many other countries. In ductless systems, a compressor and heat

exchanger are installed outside the conditioned space and refrigerant is distributed to one or more fan

coils inside the building. There may be one refrigerant-to-air fan coil unit for each room. This ductless

approach requires a larger refrigerant charge per unit of capacity, usually 0.32 to 0.34 kg of HCFC-22

per kW as compared to a centralized ducted system average of about 0.26kg per kW. An estimated

80 million ductless units are installed world wide. 

Because of the wide diversity of systems covered by this category, TEWI calculations in this

chapter are limited to 10.5 kW (3.0 ton) ducted residential systems and 26.4 kW (7.5 ton) ducted roof

top units in North America; a 2.8 kW (0.80 ton) room air conditioner/heat pump and 14.0 kW (4.0

ton) ductless, package air conditioner/heat pump in Japan; and a residential heating or cooling-only

system and commercial heating/cooling unit in Europe defined only by seasonal COPs, relative

refrigerant efficiencies, and annual loads (IEA 1994). 

4.2.1 Refrigerants

HCFC-22 is the refrigerant used in virtually all unitary equipment because of its inherent

efficiency and high refrigeration capacity. Provisions of the Clean Air Act (CAA) in the United States

call for HCFC-22 to be phased out of “new equipment” by 2010, and allow production of smaller

amounts of the refrigerant until 2020 for servicing installed equipment. This date has prompted a flurry

of research and development work at universities, research institutions, and by refrigerant manufacturers

and HVAC companies to find suitable HCFC-22 replacements. 

No single-component refrigerant or blend has been identified which can successfully replace

HCFC-22 in all of these unitary applications. The Air-Conditioning and Refrigeration Institute (ARI),

through its Alternative Refrigerants Evaluation Program (AREP), led an international effort to measure,

evaluate, and disseminate data on potential HCFC-22 and CFC-502 alternatives. The primary goal of

the program was to measure the performance of alternatives under conditions indicative of the

refrigeration and air conditioning application currently served by these refrigerants. Issues such as

flammability, toxicity, fractionation, etc. were not addressed. Based on testing performed under AREP,

several alternative refrigerants were identified which gave similar or slightly improved performance when

compared to HCFC-22 (Godwin 1994). The program also clearly indicated that no single alternative

studied was superior to HCFC-22 in all of its current applications, and that several refrigerants or

blends of refrigerants may be needed to fill the void left when HCFC-22 is phased out of production.

TEWI results from previous reports have indicated that the direct GWP of the refrigerant used

for unitary equipment contributes less than 10% of the total TEWI for these products depending on

what assumptions are used for the analysis, and that the direct GWP of the refrigerant is less important

than the overall efficiency of the unitary system (Fischer 1991) (Fischer 1994). This indicates that any

refrigerant or refrigerant blend proposed as an alternative for HCFC-22 must provide good cycle

efficiency in addition to a low to moderate GWP.
39

4.2.1.1 HFC-134a

HFC-134a is one commercially available refrigerant initially considered by AREP as an

alternative candidate for HCFC-22. It is a single component refrigerant that has been widely adopted 

by the domestic refrigeration, automotive air conditioning, and chiller air conditioning market segments,

so the major refrigerant manufacturers have constructed manufacturing plants and have extensive

product application literature. HFC-134a has a 40% lower refrigeration capacity than HCFC-22 under

unitary operating conditions and has shown a 5% decrease in efficiency for typical unitary applications,

so it cannot be considered as a “drop-in” replacement. Refrigerants with markedly lower cooling

capacities than HCFC-22 are at a disadvantage because losses in refrigerant volumetric capacity

necessitate larger compressors and heat exchangers to maintain system capacity. These translate into

higher equipment costs to the consumer. 

HFC-134a is not expected to be widely used in U.S. or Japanese unitary equipment. Some

heating-only, HFC-134a based units are included in the TEWI results for European unitary equipment

with the assumptions that it was used in units specifically designed for this refrigerant, and that it

performed at an efficiency equal to equipment using HCFC-22 (Fischer 1994). With the current

information on prospective HCFC-22 alternatives and the increasing emphasis on more efficient cooling

performance, it is unlikely that HFC-134a will be extensively used in unitary air conditioning or heat

pumps (UNEP 1995).

4.2.1.2 HFC Mixtures

The most likely replacements for HCFC-22 which came out of the AREP program are binary

or ternary mixtures which are ozone-safe, non-flammable, non-toxic, efficient, and have performance

levels close to that of HCFC-22. 

R-407C

R-407C is a 23/25/52 mass % blend of HFC-32, HFC-125, and HFC-134a. It showed

equivalent capacities to HCFC-22 but efficiencies averaged about 5% lower in soft-optimized

equipment (Godwin 1994). R-407C is a zeotropic blend which will fractionate or change composition

during evaporation and condensation in vapor compression refrigeration applications and will show

about a 5E C (9E F) change in temperature (temperature glide) across the heat exchangers due to this

composition change. This heat exchanger temperature glide and tendency to fractionate make zeotropes

less attractive commercially. It is a departure from the isothermal phase change behavior of pure

refrigerants to which the industry has become accustomed, and system leaks with zeotropes may result

in composition changes making service and repair more difficult. 

Testing of R-407C in laboratory and soft-optimized commercially produced equipment has

established capacity and system efficiency levels relative to HCFC-22 which allow TEWI evaluations

for unitary equipment (Hwang 1995, Murphey 1995, Judge 1995, Berglof 1996, Linton 1996).
40

R-407C has an ASHRAE safety classification “A1/A1” which designates it as a commercially available

refrigerant with low toxicity and no flame spread.

R-410A

R-410A is a mixture of HFC-32 and HFC-125 with a 50/50 mass % composition. AREP

found the blend’s capacity was essentially the same as HCFC-22 given compressors appropriately

sized for the difference in volumetric capacities of the two refrigerants. The ARI results also indicated

that system cooling efficiencies averaged from 1% to 6% higher than HCFC-22 (Godwin 1994).

Extensive testing of R-410A has also occurred subsequent to the AREP reports (Hwang 1996,

Murphey 1995, Feldman 1995, Linton 1996). The latest results from these tests indicate that R-410A

can be used in redesigned unitary equipment with no decrease in system capacity and a 4 to 7%

increase in system efficiency. Most of this system efficiency gain is attributed to improved

thermophysical properties of the blend over HCFC-22. System efficiency results from this latest series

of tests were used for the TEWI results presented here.

R-410A is considered a “near azeotrope” in that it does not fractionate during a phase change

in refrigeration equipment. One drawback of the mixture is that it has a system operating pressure

approximately 50% higher than HCFC-22 so it cannot be considered as a drop-in replacement or for

retrofit into existing unitary systems. Design changes will be required to accommodate these higher

operating pressures. Another more subtle drawback of this HFC mixture is a critical temperature

significantly lower than that of HCFC-22 ( 73.3E C for the HFC blend versus 96.1E C for HCFC-22).

This could diminish efficiency relative to HCFC-22 at higher condensing temperatures. R-410A has an

“A1/A1” ASHRAE safety classification. 

4.2.1.3 Propane

Propane (R-290) can be a good refrigerant and it is attracting attention as an alternative to

HCFC-22. The major disadvantage with propane, naturally, is that it is flammable Due to their

flammability, hydrocarbon refrigerants like propane, are seriously considered in low-charge systems

such as refrigerators, freezers, and packaged coolers (Stene 1996). The most significant use of 

hydrocarbons as working fluids is found in the United Kingdom and Germany. There is a strong

reluctance on the part of manufacturers in the United States and Japan to expose customers to the

hazards of flammable refrigerants in residential and commercial applications, which will make it difficult

for propane to gain wide acceptance in these markets.

One evaluation of propane conducted by an equipment manufacturer reported a slightly better

efficiency and capacity for a 9 kW (21⁄2 ton) air conditioner compared to an HCFC-22 system

(Treadwell 1994). Part of this work involved a cost estimate for a 12 kW (31⁄2 ton) unitary air

conditioner using propane which came out be 30% higher than a comparable system using HCFC-22.

The increased costs are due to modifications necessary to handle a flammable refrigerant:

! completely sealing refrigerant tubing, 
41

! sealing or relocation of potential indoor-side ignition sources from the blower motor, wiring,

and motor capacitor, 

! possibly increasing the wall thickness of heat exchanger return bends, 

! a pressure relief valve, and 

! a propane leak detector. 

These cost items are not required when using a non-flammable refrigerant. In addition, there are costs

to the manufacturer associated with safe storage, charging, and handling of a flammable refrigerant in

the plant. More recent evaluations on the feasibility of substituting propane for HCFC-22 in ducted

residential air conditioners have also focused on the relative costs versus environmental benefits, an

approach suggested by Kuijpers (1995). Three engineering solutions proposed to mitigate the

flammability risk of propane are to: 1) prevent leakage and remove all sources of ignition as suggested

by Treadwell above; 2) add a flame suppressant like HFC-227ea or a fluoroiodocompound to

propane in sufficient quantities to make the mixture non-flammable; and 3) use a secondary loop to

prevent the propane from entering residences (Douglas 1996, Keller 1996). Conclusions from these

studies indicate that TEWI could be reduced if propane is used as a refrigerant in leak-tight air

conditioner/heat pumps designed according to specifications suggested in option number 1 above, but

that applying the additional costs to improve the efficiency of systems using non-flammable refrigerants

would be more cost effective. The market for unitary equipment is very price sensitive and any cost

premium would put air conditioners at a great disadvantage in free, competitive markets.

Several recent research reports indicated that propane is being substituted for HCFC-22 with

slight (. 5%) increases in system efficiency in hydronic, heating only heat pumps commonly used in

Europe (Lystad 1996, Rodecker 1996).

In this report, TEWI calculations are performed for propane in a ducted air conditioning/heat

pump system which employs an intermediate heat exchanger and indirect loop to prevent flammable

refrigerant from entering the conditioned space. Since the steady-state COP of propane and HCFC-22

are essentially equivalent, the TEWI for a propane system that does not use an intermediate loop or a

mixed flame suppressant can be reasonably estimated from the indirect contribution for HCFC-22

systems.

4.2.1.4 Ammonia

Ammonia (R-717) is a good refrigerant which is likely to experience broader application as

CFCs and HCFCs are phased out, but it is not a choice well suited for unitary equipment (Fairchild

1995). The chemical properties of ammonia present material compatibility problems not experienced

with the fluorocarbon and hydrocarbon alternatives. 

Residential and light commercial air conditioning systems are mass produced using copper for

refrigerant tubing, in the heat exchangers and connecting components, using hermetic compressors
42

whose electric motors employ copper wire windings, and using direct heat exchange evaporators.

Ammonia is incompatible with copper in the presence of water, and materials need to be changed to

eliminate the use of copper to ensure acceptable equipment lifetimes. Aluminum or steel are typically

used for piping and tubing in industrial ammonia refrigeration systems with accompanying material and

installation cost increases. Efforts are underway to develop electric motors for use in ammonia systems

but have not yet resulted in hermetic compressors for unitary applications. 

Direct heat transfer evaporators are not considered feasible with ammonia in residential or other

high-occupancy applications, so a secondary heat transfer loop and fluid (e.g. brine, glycol) would be

needed. This additional heat transfer loop increases the cost and complexity of the system and reduces

the efficiency. Ammonia also has high discharge temperatures that need to be reduced to avoid

damaging the system; this can be done economically on larger refrigeration systems that are used in

centralized cold storage or food processing applications but the cost could be prohibitive on small

systems.

In view of the above considerations, it is not considered likely that ammonia will be used

commercially in unitary equipment as a replacement for HCFC-22. No TEWI calculations are

performed for unitary systems using ammonia as a vapor compression refrigerant. Ammonia-water

absorption equipment options are considered.

4.3 ALTERNATIVE TECHNOLOGIES

Electric resistance heat is evaluated in appropriate locations to provide TEWI comparisons to

heat pump options. Residential gas heating and cooling options are evaluated in a subsequent section.

4.4 ASSUMPTIONS

National averages are used for the power plant emission rates in these calculations. The

averaged electrical power plant emission rates are 0.650 kg CO2/kWh for the United States, 0.470 kg

CO2/kWh for Europe, and 0.473 kg CO2/kWh for Japan , Appendix A. These CO2/kWh emission

rates are compiled from open literature data rather than calculated from the fundamental heat content of

fuels, fuel mix used in power production, plant efficiencies, and transportation and distribution losses.

Published measurements for steady-state COP data relative to HCFC-22 and fixed values for

the seasonal energy efficiency ratio (SEER) or heating seasonal performance factor (HSPF) of HCFC22 were used to calculate SEERs and HSPFs for propane and the R-407C and R-410A mixtures. The

relative efficiency values used for these calculations are summarized in Table 5. These values are based

on recent test results in “soft optimized” systems. Further development of air conditioners specifically

designed to use these alternative refrigerants could lead to more favorable comparisons relative to

HCFC-22.

Propane’s performance for a 10.6 kW (3 ton) ducted unitary systems was attenuated by

assuming an intermediate heat transfer loop to keep this flammable refrigerant out of the occupied
43

Refrigerant Components (mass % composition)

Efficiency Relative to HCFC-22

1996 2005

Cooling Heating Cooling Heating

HCFC-22 HCFC-22 (100%) 100% 100% 100% 100%

HFC-134a HFC-134a (100%) 100% 99% 100% 99%

HC-290

 secondary hx at 2.8ECÄT

 secondary hx at 5.6ECÄT

HC-290

HC-290

85%

72%

88%

77%

85%

72%

88%

77%

R-407C HFC-32/HFC-125/HFC-134a (23/25/52) NA* NA* 100% 100%

R-410A HFC-32/HFC-125 (50/50) NA* NA* 105% 105%

* not available

Table 5. Relative efficiencies for alternative refrigerants for unitary equipment (relative to HCFC-22).

space. A 2.8E C (5E F) or 5.6E C (10E F) lower evaporator or higher condenser temperature were

assumed to drive this intermediate loop, resulting in lower cycle efficiencies. A parasitic pump load of

one-fourth horse power was assumed for the 10.6 kW (3.0 ton) residential systems. Since hydronic

heating systems with intermediate, secondary heat exchange loops are more common in Europe, the

additional energy used and CO2 generation associated with this intermediate loop was not applied to

results for European residential heating.

Fifteen year lifetimes are assumed for U.S. and European unitary equipment. Based on

information provided by Hara from JICOP and Hisajima of JRAIA, the lifetime of the Japanese 2.8 kW

(0.8 ton) split, room air conditioner/heat pump was taken as 12 years and the 14.0 kW (4.0 ton) 

packaged heat pump was set at 10 years (Hara 1996b).

Based on information assembled from ARI member companies, the maximum residential heat

pump and air conditioner annual leak rates of 4% of the charge for 1996 equipment and 2% per year

for equipment available in 2005 were used for the direct TEWI calculations (Hourahan 1996a). This

same group recommended a 0.25 to 1.5% of charge annual loss rate for 1995 roof top equipment

(1.5% was used) and a 0.25 to 1.0% rate for 2005 equipment (1.0% was used). Since roof top units

require no field refrigerant line connections, there is less opportunity for leakage. An end-of-life

(E.O.L.) charge loss rate of 15% was rationalized for both residential and roof top units on the basis of

recovering 90% of the charge from 95% of the field units, but allowing for a 100% charge loss from

about 5% of field stock (Hourahan 1996a). The 1996 annual leak or make up rate and 15% E.O.L.

loss were also used for the European equipment calculations. 

Hara and Hisajima recommended that an annual charge loss rate of 0.5% be used for the

Japanese 2.8 kW (0.8 ton) split, room air conditioner/heat pump and 0.2% for the 14.0 kW (4.0 ton) 

packaged heat pump (Hara 1996a). End-of-life charge loss rates were taken as 15% of the total

charge for both of these heat pumps (see Table 6).
44

Room Air Conditioner /

Heat Pump System

Packaged Air Conditioner /

Heat Pump System

Base System HCFC-22; 2.8 kW split HCFC-22; 5 PS split

COP

 Cooling

 Heating

2.67

3.20

2.50

3.00

Annual Equivalent Full Load

Operating Hours (Tokyo)

 Cooling

 Heating

350

450

600

900

Charge Weight HCFC-22 (kg) 1.0 4.45

Annual Refrigerant Make Up Rate

 (rates suggested by Hara 1996a)

0.50% 0.20%

Charge Loss at End-of-Life 15% 15%

Equipment Lifetime 12 years 10 years

COP Relative to HCFC-22

 R-407C

 R-410A

---

1.0

0.95

0.95

Charge Weight Relative to HCFC-22

 R-407C

 R-410A

---

1.0

1.0

1.0

Table 6.. Performance data for unitary Japanese systems used for the TEWI calculations.

4.5 METHODOLOGY

Total equivalent warming impacts were calculated for baseline 10.5 kW (36,000 Btu/h) heat

pumps with SEERs of 10, 12, and 14; and corresponding HSPFs of 7, 8, and 9 with a refrigerant

charge of 2.8 kg (6.2 lbs) of HCFC-22 for three locations in the U.S. In calculations where air 

conditioning is combined with some other heating technology, a central air conditioner with SEERs of

10 and 12 was used. SEERs and HSPFs for a geothermal heat pump were chosen from information

provided by major manufacturers and results of standard rating/certification tests (ARI #330-93 1993). 

Seasonal energy use is computed based on a “typical” 1,800 sq. foot residence with a 74.7x106 Btu/yr

(78.8x106 kJ/yr) heating load and 16.1x106 Btu/yr (17.0x106 kJ/yr) cooling load in Pittsburgh; a

34.8x106 Btu/yr (36.7x106 kJ/yr) heating load and 33.8x106 Btu/yr (35.7x106 kJ/r) cooling load in

Atlanta; and a 0 Btu/yr (0 kJ/yr) heating load and 82.2x106 Btu/yr (86.7x106 kJ/yr) cooling load in

Miami (Ballou 1981).
45

System

Efficiencies

Cooling / Heating

Air-to-Air Heat Pumps

 HCFC-22: 

 minimum efficiencies

 high efficiencies

SEER-10 / HSPF-7

SEER-12 / HSPF-8

Premium Technologies

 Air-to-Air Electric Heat Pump

 Geothermal Heat Pump

SEER-14 / HSPF-9

SEER-15.8 / HSPF-12

Table 7. Current technology (1996) efficiency data.

System

Efficiency

Cooling / Heating

Air-to-Air Electric Heat Pumps

 HCFC-22

 R-407C

 R-410A

 HC-290

 2.8EC ÄT in secondary loop

 5.6EC ÄT in secondary loop

SEER-12 / HSPF-8

SEER-12 / HSPF-8

SEER-12.6 / HSPF-8.4

(SEER-12 / HSPF-8)

SEER-10.9/HSPF-7.2

SEER-10.2/HSPF-6.2

Premium Technologies

 Air-to-Air Electric Heat Pumps

 Geothermal Heat Pumps

SEER-14 / HSPF-9

SEER-17.2 / HSPF-12.8

Table 8. Future system efficiencies (2005).

For ducted residential equipment available in 2005, the baseline heat pump was assumed to have

an SEER of 12 and a HSPF of 8 and the SEER of a central air conditioning unit was taken to be 12.

Cooling-only TEWI calculations for 26.4 kW (7.5 ton) roof top units with full load rating EERs of 

10 for 1996 and 11 for 2005 with a HCFC-22 charge of 6.9 kg (15.2 lbs) were calculated for

Pittsburgh with 600 equivalent full load cooling hours, Atlanta with 1,400 equivalent full-load cooling

hours, and Miami with 2,700 equivalent full-load cooling hours. Annual energy use for these units was

calculated based on their capacity (90,000 Btu/hr), the EER, and equivalent full load hours.

Residential heating and cooling-only TEWI calculations were performed for Europe using averaged

loads and seasonal performance factors (SPFs), Appendix D. An average heating load for all European

countries was used for heating TEWI calculations but only southern Europe (Greece) was used for the

residential air conditioning calculations. Heating and cooling TEWIs for a composite European

commercial building were calculated for heat pumps using averaged loads and heat pumps with various

refrigerants and an HCFC-22 air conditioner used with electric resistance heat. The European average

of 0.47 kg CO2/kWh was used to convert electric use to CO2 emissions with the exception of the

residential air conditioning calculation which used Greece’s emission rate of 0.976 kg CO2/kWh.

TEWI results for Japanese

unitary equipment were

calculated using data from Table

6. This data was provide by 

Hara and Hisajima (Hara

1996b). The Japanese average of

0.473 kg CO2/kWh was used to

convert electric use to CO2

emissions for these calculations.

4.6 RESULTS AND

DISCUSSION

Total equivalent warming impacts

for various residential heating/cooling

options were calculated for

Pittsburgh, Atlanta, and Miami in the

United States, and the results are

shown in Figs. 8 to 13. These results

are computed using the efficiency

data in Table 7 and 8. Each figure has

two sections, the upper portion

shows “standard technologies,” or

heating/cooling options that represent

baseline cost for a residential system
Air-to-Air HP

 HCFC-22

 R-407C

 R-410A

 HC-290: 2.8°C

 HC-290: 5.6°C

Premium Technologies

 HCFC-22

 Geothermal HP

0 25,000 50,000 75,000 100,000 125,000 150,000 175,000 200,000

Indirect Effects

electricity

resistance heat

refrigerant emissions

Direct Effects

pump power

TEWI (kg CO ) 2

Air-to-Air Heat Pump

 HCFC-22: minimum

 HCFC-22: hi eff

Premium Technologies

 Air-to-Air HP

 Geothermal HP

0 25,000 50,000 75,000 100,000 125,000 150,000 175,000

TEWI (kg CO )

Indirect Effects

electricity

resistance heat

refrigerant emissions

Direct Effects

2

46

Figure 9. TEWI for residential heating and cooling options in Pittsburgh, Pennsylvania USA (2005

efficiency levels).

Figure 8. TEWI for residential heating and cooling options in Pittsburgh, Pennsylvania USA (1996

efficiency levels).
Air-to-Air Heat Pump

 HCFC-22: minimum

 HCFC-22: hi eff

Central A/C & Furnace

hi eff & resistance

Premium Technologies

 Air-to-Air HP

 Geothermal HP

0 25,000 50,000 75,000 100,000 125,000 150,000 175,000

TEWI (kg CO ) 2

Indirect Effects

electricity

resistance heat

refrigerant emissions

Direct Effects

Air-to-Air A/C

 HCFC-22

 R-407C

 R-410A

 HC-290: 2.8°C þT

 HC-290: 5.6°C þT

Premium Technologies

 HCFC-22

 Geothermal HP

0 25,000 50,000 75,000 100,000 125,000 150,000

Indirect Effects

electricity

refrigerant emissions

Direct Effects

pump power

TEWI (kg CO ) 2

47

Figure 10. TEWI for residential heating and cooling options in Atlanta, Georgia USA (1996

efficiency levels).

Figure 11. TEWI for residential heating and cooling options in Atlanta, Georgia USA (2005

efficiency levels).
Air-to-Air Heat Pump

 HCFC-22: minimum

 HCFC-22: hi eff

Premium Technologies

 hi eff a/c

 geothermal a/c

0 25,000 50,000 75,000 100,000 125,000 150,000

Indirect Effects

electricity

refrigerant emissions

Direct Effects

TEWI (kg CO ) 2

Air-to-Air A/C

 HCFC-22

 R-407C

 R-410A

 HC-290: 2.8°C

 HC-290: 5.6°C

Premium Technologies

 HCFC-22

 Geothermal A/C

0 25,000 50,000 75,000 100,000 125,000 150,000

TEWI (kg CO ) 2

Indirect Effects

electricity

refrigerant emissions

Direct Effects

pump power

48

Figure 12. TEWI for residential heating and cooling options in Miami, Florida USA (1996

efficiency levels).

Figure 13. TEWI for residential heating and cooling options in Miami, Florida USA (2005

efficiency levels).
49

in each of these cities, while the lower indicates a “premium technologies” section which is a

heating/cooling option that is significantly more expensive than the baseline technology. Figures 8, 10,

and 12 show unitary systems currently available, whereas Figures 9, 11, and 13 suggests what TEWI is

possible from technologies anticipated by the year 2005. Attaching specific prices to each option is

difficult and often-times misleading because HVAC manufacturers, dealers, installers, and local utilities

can all influence the final price paid by the consumer. While specific prices are problematic, it is clear

that newly developed and more efficient options shown will cost more than conventional systems, and 

dividing these technology options into standard and premium categories gives some indication of the

added investment required to obtain a TEWI benefit (Kuijpers 1995). 

Each segment of the bar graphs plotted in these figures indicates TEWI contributions from different

sources. The initial, darker gray section of most bar graphs is the indirect TEWI contribution from

electric power used for the vapor compression heating and/or cooling process. In the Pittsburgh results,

Figs. 8 and 9, electric resistance heat to supplement heat pump operation is also shown. The lighter

section on some bars shows the TEWI contribution from auxiliary electric technologies such as pumps

or fans that are not included in the SEER or HSPF ratings of the equipment. Electricity used for pump

power on the secondary loop for the HC-290 heat pump/air conditioner would be an example of

contributions added here. The moderate gray section on the ends of most bar graphs is the direct

TEWI contribution caused by refrigerant losses.

Using Figs. 8 and 9 as examples, the advantages of increasing unit efficiencies become quite

obvious if the HCFC-22, SEER=10/HSPF=7, SEER=12/HSPF=8, and SEER14/HSPF=9 (in the

premium technologies section) are compared. Total TEWI values for these three heat pump options in

Pittsburgh are about 126,000; 111,000; and 100,000, respectively. A 10 to 12% improvement in

TEWI is indicated for each step of efficiency improvement. Relative TEWI decreases with increased

efficiency are greater in climates with a higher cooling/heating ratio.

Figure 9 also shows the benefits in TEWI and relative energy savings associated with the added

expense of a geothermal or ground-source heat pump

In all three U.S. cities, use of propane as the refrigerant with a secondary loop incurred a TEWI

penalty. Direct propane systems with all the added safety precautions and increased costs needed to

make them safe for the U.S./North American market are not shown, but are assumed to perform

similar to the HCFC-22 heat pump/air conditioners with essentially no direct TEWI contribution from

the refrigerant (Keller 1996).

Figure 10 shows a TEWI reduction resulting from residential heating in Atlanta via a heat pump

rather than electric resistance heat. Similar reductions in energy use and indirect TEWI are possible if

heat pump water heaters replace conventional electric water heaters (resistance heat) for domestic hot

water applications in residential and commercial buildings (IEA 1993).

There are bars for the TEWI results for HFC mixtures R-407C and R-410A as well as for

HCFC-22 in Figs. 9, 11, and 13. None of the systems represented were specifically designed for

optimal performance with these mixtures and consequently the energy efficiencies reflect test results in

“soft optimized” systems. In all the cases presented, the direct contribution of refrigerant to the TEWI

was no larger than 6% of the total. The average direct contribution is more like 3 to 4%. Essentially no

difference is seen in the TEWIs for HCFC-22 systems and those where R-407C or R-410A are used
Pittsburgh

HCFC-22

HCFC-22

R-407C

R-410A

Atlanta

HCFC-22

HCFC-22

R-407C

R-410A

Miami

HCFC-22

HCFC-22

R-407C

R-410A

0 50,000 100,000 150,000 200,000 250,000

TEWI (kg CO )2

Indirect Effects

electricity

refrigerant emissions

Direct Effects

1996

2005

2005

2005

1996

2005

2005

2005

1996

2005

2005

2005

50

Figure 14. TEWI for 7.5 ton roof top air conditioning units in the USA.

as substitutes, because unit efficiencies are very similar and the 100 year ITH refrigerant GWPs are

1700 for HCFC-22, 1530 for R-407C, and 1730 for R-410A. The smaller charge sizes per unit of

capacity for R-410A and early indications of system efficiency improvements over HCFC-22 will help

reduce TEWI even more for this option.

Nearly 80% of the direct effect is due to the assumption on annual emissions from leakage,

accidents, and maintenance practices. As regulatory procedures requiring conscientious maintenance

and repairs of leaks and strict adherence to refrigerant recovery come into common usage and are

followed, the direct effect will diminish in significance.

TEWI results calculated for 26.4 kW (7.5 ton) roof top air conditioning units operated with

HCFC-22, R-407C, and R-410A in Pittsburgh, Atlanta, and Miami are shown in Fig. 14. In this graph,

the indirect TEWI contribution from electricity consumed is indicated by the darker gray bar segment.

Direct TEWI contribution from the refrigerant is indicated by the lighter gray portion of the bar. It is

apparent from these results that the direct contribution to TEWI due to refrigerant emissions from this

line of products is less than 8% in Pittsburgh which has the least cooling hours and correspondingly low

indirect TEWI contribution. The relative contribution of refrigerant emissions to the total TEWI

decreases rapidly in locations with more operating hours. 

 Increasing the EER of this equipment from 10 to 11 results in a 9 to 10% decrease in indirect

TEWI. Also indicated in Fig. 14, is the similarity of TEWI results for HCFC-22 and the HFC

alternatives under the assumptions used for these calculations.
Air-to-Air Heat Pump

HCFC-22

HFC-134a

R-407C

R-410A

Electric A/C & Resistance

0 200,000 400,000 600,000 800,000 1,000,000 1,200,000

electricity

refrigerant emissions

Indirect Effect

Direct Effect

TEWI (kg CO ) 2

51

Figure 15. TEWI for residential heating options in Europe.

Figure 15 shows the relative magnitude of TEWI for several residential heating options in

Europe. Direct TEWI from refrigerants for the vapor compression technologies average about 10% of

the total if leak rates and end-of-life losses similar to North American equipment are assumed. The

relative direct TEWI effect for this equipment in Europe is larger than for comparable North American

systems because a smaller load is assumed and the average CO emission rate from electric power 2

plants, 0.47 kg CO2/kWh, is less than the U.S. average. Even with this lower power plant CO2

emission rate, the use of electric resistance heat results in a significantly larger TEWI when compared to

the other technologies.

TEWI results for residential, cooling-only system options in Europe based on building loads and

system performance values listed in Appendix D, are shown in Fig. 16. Since cooling loads given in the

reference are specific to Greece, the power plant carbon dioxide emission rate for Greece, 0.976 kg

CO2/kWh, was also used for the calculations and results displayed. The direct TEWI portion of the

various vapor compression technologies shown ranges from 10 to 15% of the total TEWI.

Figure 17 summarizes TEWI results calculated for a composite commercial building averaged

from tabulated results presented for commercial building unitary applications in Europe (IEA 1994).

With annual charge leak rates of 4% and a 15% loss of refrigerant on unit disposal, the direct TEWIs

obtained in these results are less than 2% of the total. Clearly, the direct TEWI contribution for this

application is not significant. No significant difference in TEWI is seen between systems using HCFC22 and those using alternatives. Electric resistance heat shows increased TEWI.
Air-to-Air HP/AC

HCFC-22

HFC-134a

R-407C

R-410A

Geothermal HP/AC

HCFC-22

HFC-134a

R-407C

R-410A

0 20,000 40,000 60,000

electricity

refrigerant emissions

Indirect Effect

Direct Effect

TEWI (kg CO )2

Air-to-Air Heat Pump

HCFC-22

HFC-134a

R-407C

R-410A

Electric A/C & Resistance

0 200,000 400,000 600,000 800,000 1,000,000 1,200,000

electricity

refrigerant emissions

Indirect Effect

Direct Effect

TEWI (kg CO )2

52

Figure 16. TEWI for residential cooling-only options in southern Europe (Greece).

Figure 17. TEWI for commercial heating and cooling options in Europe.
HCFC-22

R-407C

R-410A

HCFC-22

R-407C

R-410A

0 1,000 2,000 3,000 4,000 5,000 6,000 7,000

TEWI (kg CO )

Indirect Effects

Cooling

Heating

Direct Effects

Refrigerant

Emissions

2

Emission Rates:

 1⁄2% per year

 5% end-of-life

Emission Rates:

 1⁄2% per year

 15% end-of-life

Lower Eff. Assumption

HCFC-22

R-407C

R-410A

Higher Eff. Assumption

R-407C

R-410A

0 10,000 20,000 30,000 40,000 50,000 60,000 70,000

electricity

refrigerant emissions

Indirect Effect

Direct Effect

TEWI (kg CO ) 2

53

Figure 18. TEWI for 2.8 kW room air conditioner/heat pump in Japan.

Figure 19. TEWI for a 14.0 kW packaged air conditioner/heat pump in Japan.
54

Results for TEWI calculations performed on a 2.8 kW room air conditioner/heat pump and a

14.0 kW packaged air conditioner/heat pump typical of those used in Japan, under operating conditions

indicative of Tokyo are shown in Figs. 18 and 19, respectively. Direct TEWI accounts for less than 8%

of the total under the most pessimistic leak and refrigerant recovery scenarios for the room heat pump

units. Decreasing leak rates and improved refrigerant recovery estimates provided for this equipment by

JRAIA (Hara 1996b) lowers the direct effect to less than 5% for the room heat pump system and to

approximately 1% for the packaged system.

4.7 CONCLUSIONS

TEWIs for HFC mixtures proposed as HCFC-22 alternatives are not significantly different

from those calculated for HCFC-22, and with optimization of equipment design efficiency should

continue to improve. Refrigerant leakage, and the corresponding global warming impact of the

refrigerant, from hermetic unitary equipment is very small and future service losses will be low because

maintenance and replacement practices mandating refrigerant recovery and recycling are in place or

under consideration in many countries. 

The direct contributions to TEWI for all vapor compression systems presented are small fractions of the total in each case. These contributions should not be ignored, however. Procedures for

handling refrigerants and accounting for refrigerant usage currently being adopted should be effective in

reducing the direct effect from that shown in results presented in this chapter. TEWIs of fluorocarbons

system are less than those of a propane vapor compression cycle with a secondary heat exchange loop.

Arguments for not using propane in direct systems with adequate safety precautions to prevent fires and

explosions, center on the relative effectiveness of additional investments and costs.
55

5. CHILLERS

5.1 INTRODUCTION

Comfort air conditioning in larger single-story and multistory commercial buildings is usually and

most efficiently provided by chillers that cool water or a water/antifreeze mixture which is then

circulated to fan coil units or air handlers in individual rooms to cool and dehumidify the air. Chillers are

manufactured in cooling capacity ranges of 7.0 kW to 35,000 kW (2 to 10,000 refrigeration tons),

but the larger capacity, centralized units which use screw or centrifugal compressors for the refrigeration

circuit are of most interest in this chapter. Screw compressors are used in chillers with cooling

capacities ranging from 250 kW (70 tons) up to a maximum of about 2300 kW (650 tons) because

they are more compact, efficient, and reliable than comparable reciprocating compressors which

dominate the field in lower capacity units. In chiller sizes greater than 700 kW (200 tons) centrifugal

chillers are favored because of the larger volumetric flow rate of refrigerant that has to be circulated to

achieve these levels of refrigeration (Stoecker 1982). Out of an estimated 211x106 kW of installed

chiller capacity worldwide in 1991, roughly 134x106 kW or 64% was in North America, 39x106 kW

(18%) in operation in Japan, and 23x106 kW (11%) in Europe (Fischer 1991).

Chillers using centrifugal compressors comprise approximately 70% of this installed capacity, with

about 30% using a combination of reciprocating, screw, or scroll compressors. Historically, nearly twothirds of the centrifugal chillers have been low pressure machines using CFC-11 or HCFC-123, about

12% are high pressure systems using CFC-12 or HFC-134a, with the balance using primarily

HCFC-22. The discussion in this section focuses on capacity ranges served by electric-driven chillers

of 1,200 and 3,500 kW (350 and 1000 tons) and alternative cooling technologies with equivalent

capacities for large commercial and institutional buildings.

Commercially proven gas powered absorption chillers are currently available with cooling

capacities up to 6000 kW (1700 tons). Absorption machines account for about 8% of large tonnage

chiller sales in the United States and about 80 to 90% in Japan.

5.2 REFRIGERANTS

5.2.1 HCFC-123 

HCFC-123 has been developed as a replacement for CFC-11 in low pressure chiller applications.

Prior to the CFC phaseout, the vast majority of centrifugal chillers used CFC-11 because of its inherent

efficiency and because the maximum working pressure in these machines was low enough to exempt

systems from stringent pressure vessel code regulations in most countries. HCFC-123 has been the

refrigerant of choice for retrofitting installed CFC-11 machines as well as for new low-pressure chillers

because its thermodynamic properties are similar to those of CFC-11. In retrofit situations, HCFC-123

is a more aggressive solvent than CFC-11, so, as with other alternatives, hermetic motor insulation and
56

other material compatibility concerns have to be resolved. HCFC-123 has a very short atmospheric

lifetime (.1.4 years) and consequently has extremely low ozone depleting and global warming

potentials. The most efficient commercially available centrifugal chillers use HCFC-123 (UNEP 1995).

Ideal cycle, steady-state efficiencies were computed using thermodynamic properties at an

evaporator temperature of 4.4EC (40EF) and condensing temperature of 40.6EC (105EF) to compare

performance with other refrigerants (Smith 1993). These "ideal" COPs are 6.74 (0.52 kW/ton) for 0EC

(0EF) subcooling and superheat, 6.89 (0.51kW/ton) for 2.8EC (5EF) subcooling and superheat, and

7.03 (0.50 kW/ton) for 5.6EC (10EF) subcooling and superheat.

Chiller efficiencies have improved significantly since the Montreal Protocol was first adopted, and

chillers using alternative refrigerants are now more efficient than the products using CFCs that they

replaced (Glamm 1996). Chillers using HCFC-123 are commercially available with COPs as high as

7.82 (0.45 kW/ton) at the ARI rating conditions. Significant progress has also been made in reducing

refrigerant losses from purge units for low pressure chillers. Chillers with refrigerant emissions of less

than 0.5% of the charge per year are being marketed (Smithart 1993); this is significantly less than

projected "future loss rates" of 4% used in the parametric analysis in the Phase I study. ARI (the Air

Conditioning and Refrigeration Institute) estimated that the total refrigerant make-up rate resulting from

losses in purge units, leakage, maintenance, and recovery upon equipment retirement for a wellmaintained, normally operating machine would average 0.5% of the charge per year over the lifetime of

the chiller (Hourahan 1996a). 

5.2.2 HFC-134a

HFC-134a is the refrigerant of choice for the new generation of chillers that evolved from those

designed to use CFC-12 and for retrofitting chillers originally designed to use CFC-12. HFC-134a has

a moderate GWP, and since it contains no chlorine, has no ozone depleting potential. Very efficient

HFC-134a chillers based on screw and centrifugal compressors are commercially available from a

number of manufacturers (E Source 1995). 

Steady-state efficiencies were computed using thermodynamic properties at an evaporator

temperature of 4.4EC (40EF) and condensing temperature of 40.6EC (105EF) (Smith 1993). These

"ideal" COPs are 6.29 for 0EC (0EF) subcooling and superheat, 6.48 for 2.8EC (5EF) subcooling and

superheat, and 6.66 for 5.6EC (10EF) subcooling and superheat (0.56 kW/ton, 0.54 kW/ton, and 0.53

kW/ton, respectively).

Efforts to remain competitive with HCFC-123 chillers have led to developments and innovations

that have boosted HFC-134a chiller efficiencies to COPs of 6.76 (0.52 kW/ton) at ARI rating

conditions. For example, some machines use cycle modifications such as expansion turbines to recover

some of the refrigerant throttling losses and apply them to supplement energy furnished by the chiller

motor. Additionally, the heat exchangers are designed for very close approach temperatures to boost

efficiencies. The reported COP of 6.76 is not directly comparable to the “ideal” efficiency calculations

given above which do not include such cycle modifications. Ideal COPs are given to provide a rough

sense of relative refrigerant efficiencies.
57

5.2.3 HCFC-22

HCFC-22 is a high-capacity refrigerant used in lower tonnage chillers with scroll, reciprocating and

screw compressors and in the largest, 5,000 to 35,000 kW, capacity chillers with centrifugal

compressors. Its low volumetric flow rate and favorable transport properties make it useful in

compressors with small displacements and heat transfer surfaces. In the largest chiller size ranges, lower

vapor pressure refrigerants like HCFC-123 require large compressor passages and interconnecting

pipe diameters so a high-capacity, higher pressure refrigerant like HCFC-22 is favored. Under

revisions of the Clean Air Act which implements the Montreal Protocol in the United States, the

phaseout date for HCFC-22 in new equipment is scheduled for 2010 and a total production phaseout

is scheduled for 2020. Each country is approaching this HCFC phaseout differently.

Azeotropic and zeotropic blends of HFCs, ammonia, and certain hydrocarbons have been

considered as alternatives for HCFC-22. ARI, under AREP (Alternate Refrigerant Evaluation

Program), has conducted a focused research effort to find and characterize suitable substitutes for

HCFC-22. Most of this work was directed at finding alternatives for smaller, unitary equipment with

direct expansion (DX) evaporators. A zeotropic blend of HFC-32/HFC-125 (R-410A) that is

seriously being considered as a HCFC-22 alternative, operates at a significantly higher pressure which

would result in new design implications for large chillers. One manufacturer has announced a line of

230-300 ton, flooded evaporator, screw chillers that will use R-410A (ACH&RN Oct. 1996). No

zeotropic blends investigated in the AREP program are considered suitable for use in flooded

evaporators where they would tend to fractionate. From information available at the time this report is

being written, no suitable short-term replacement for HCFC-22 has been identified for use in larger

capacity chillers with flooded evaporators (UNEP 1995).

As with HCFC-123 and HFC-134a, the steady state ideal efficiencies were calculated for HCFC22 using its thermodynamic properties at an evaporator temperature of 4.4EC (40EF) and condensing

temperature of 40.6EC (105EF). "Ideal" COPs under these evaporating and condensing conditions are

6.33 for 0EC (0EF) subcooling and superheat, 6.46 for 2.8EC (5EF) subcooling and superheat, and

6.58 for 5.6EC (10EF) subcooling and superheat (0.56 kW/ton, 0.54 kW/ton, and 0.53 kW/ton,

respectively).

Market pressures force HCFC-22 chillers to have efficiencies similar to the HCFC-123 and HFC134a based equipment against which they compete. Many of the same efficiency-enhancing design

improvements applied to machines using other CFC alternative refrigerants are equally applicable to

HCFC-22. A survey of currently available commercial equipment in the 350—3,500 kW (100—1000

ton) capacity range indicates that HCFC-22 screw compressors with full load COPs of 6.39 (0.55

kW/ton) are readily available. HCFC-22 centrifugals are certified with COPs of 6.51 (0.54 kW/ton) at

the ARI rating point conditions (Hourahan 1996b).

5.2.4 Other Refrigerants

Other refrigerants suggested for large chiller applications are ammonia (R-717, NH3), propane

(HC-290), and HFC-245ca. Ammonia is an excellent refrigerant that is routinely used for large
58

refrigerated warehouses. Since ammonia is moderately flammable and moderately toxic, its use has

been confined to systems that are easily isolated from the general public. The technical expertise to

develop and adapt ammonia to comfort cooling applications in chillers exits and is being applied

(Fairchild 1995). Capital investment in redesign, retooling, training, maintenance, service, and marketing

needed to introduce this product as a major competitor for an established non-flammable, non-toxic

technology will be considerable (Stene 1996). TEWIs for ammonia chillers are calculated in this report

to show what effect a zero-GWP refrigerant has on this application. An assessment of some of the

performance and commercial potential considerations for ammonia chillers are given in the following

section. 

Theoretical efficiency estimates for hydrocarbons like propane (HC-290) in chiller applications

have been made (Hayes 1989). Commercial chiller systems using hydrocarbon refrigerants in a

capacity range from 10 to 90 kW (3 to 26 tons) are available in the United Kingdom (Morris & Young

1996), but hydrocarbon refrigerants are undesirable for bigger chillers because of the large refrigerant

charge required in these machines. No TEWI calculations are performed for hydrocarbon refrigerants

because they are an unlikely alternative technology for the capacity ranges considered for this study

(UNEP 1995).

HFC-245ca has been proposed as a chlorine-free, zero-ODP alternative for HCFC-123 (Sand

1991, Smith 1993). Laboratory comparisons of HFC-245ca performance relative to CFC-11 and

HCFC-123 have been performed in a 700 kW (200 ton) three-stage centrifugal chiller (Keuper 1996),

and it was concluded that this HFC can obtain performance comparable to HCFC-123 in a redesigned

centrifugal compressor. There are, however, some flammability concerns with HFC-245ca. This

refrigerant has been found to be marginally flammable, and its flammability is dependent on the moisture

content of the refrigerant/air mixture and the ignition source. In addition it has not been evaluated under

the demanding toxicity tests required for a new commercial refrigerant and no chemical manufacturer

has announced plans to build a production facility for HFC-245ca. Unless the situation changes, it will

not be commercially available. Consequently, HFC-245ca was not evaluated as an alternative

refrigerant for commercial chillers in this study.

5.2.4.1 Ammonia Chillers

Major chiller manufacturers have looked at using ammonia in chilled water systems, and while some

machines are commercially available in Europe there is a great difference in attitudes between the U.S.

and Europe on the use of ammonia as a chiller refrigerant (Lindborg 1997).

5.2.4.2 Theoretical Performance

Steady-state efficiencies were computed using thermodynamic properties at an evaporator

temperature of 4.4EC (40EF) and condensing temperature of 40.6EC (105EF) (Smith 1993). These

"ideal" COPs are 6.60 for 0EC (0EF) subcooling and superheat, 6.64 for 2.8EC (5EF) subcooling and

superheat, and 6.68 for 5.6EC (10EF) subcooling and superheat (0.53 kW/ton, 0.53kW/ton, and 0.52

kW/ton, respectively).
59

No hard data were located comparing operating efficiencies of ammonia and other chiller systems.

Ammonia chillers are commercially available in Europe and they are reported to be more efficient than

HFC-134a chillers in this market (Mosemann 1993). Ammonia and HCFC-22 are expected to have

comparable efficiencies in chillers using screw compressors (Calm 1994). Ammonia is incompatible

with copper alloys, so chillers using ammonia would require an open drive and external electric motor

(Mosemann 1993).

Screw chillers using ammonia are commercially available with good efficiencies. Some

manufacturers are producing screw chillers for refrigeration equipment designed for ammonia which are

initially using HCFC-22 (Fairchild 1995). Building and fire codes that have evolved restricting the use

of ammonia out of concerns for public safety make it difficult to obtain the permits and licenses

necessary to use ammonia chillers in urban areas of the U.S. and many other countries. These

institutional factors limit the potential for ammonia chillers to replace fluorocarbon refrigeration in this

application. An appraisal of the potential benefits from expanded use of ammonia as a refrigerant and

the risks associated with liberalizing regulations affecting its use are beyond the scope of this study

(Fairchild 1995).

5.3 ALTERNATIVE TECHNOLOGIES

5.3.1 Natural Gas Engine Driven Chillers

Engine driven chillers employ the same cooling process as conventional electric-powered systems

except the electric motor is replaced by an internal combustion engine usually powered by natural gas.

Since gas is substituted for electricity as the power source for this air conditioning equipment, the

indirect TEWI is likely to be different for any given application. TEWI values are calculated for gas

engine-driven chillers to provide a comparison with electrically driven technologies.

Some other differences in the application of electrically driven versus gas engine-driven chiller

applications would be variations in maintenance and service costs, additional provisions for rejecting

more heat from gas combustion technologies via the cooling tower, and the opportunity to effectively

utilize the high quality waste heat from gas-engine and gas fired absorption applications for space or

water heating and possibly desiccant-based air conditioning regeneration.

The theoretical cycle performance of gas-engine driven equipment is virtually the same as that of

electric chillers using the same compressors and refrigerants. Full load gas-input-based cooling COPs

for engine driven chillers ranges from 1.50 to 2.08 depending on the capacity and combination of

compressor and refrigerant used (AGCC 1996). Full load COPs are generally smaller than the IPLV

COPs used in TEWI calculations. These COPs are expressed in terms of the kWh of cooling output

per kWh of high thermal input from the gas. They do not include auxiliary electric power required by

electrical components on the chiller or (as with other chillers) any energy associated with cooling tower

operation or chilled water circulation pumps.

Commercially proven, packaged, natural gas engine-driven water chillers are available today. Most

natural gas engine-driven chillers in the 350 to 3,500 kW capacity range use HCFC-22 in screw
60

compressors. Screw and/or centrifugal compressors are also available which use HFC-134a. Local

market conditions and regional factors, such as the relative costs of gas and electricity, peak load

charges and local rebates are the important criteria used to choose between electric and engine chillers.

5.3.2 Absorption Chillers

Gas-fired absorption chillers are a commercially available alternative to vapor compression

centrifugal chillers. Kohler reported on the status of absorption chiller technology and the opportunities

to substitute gas-fired chillers for electric driven centrifugal chillers using HCFCs or HFCs (1993):

"Absorption water chillers are sold in the U.S. in capacities from 100 to 1500 tons (350 to

5300 kW).... The principal competition for absorption chillers is electrically driven vapor

compression equipment. The decision as to which product to use is generally made based on

life cycle costs. Since absorption chillers have higher first costs, operating cost savings or other

incentives are required to justify their purchase.

"Double effect machines represent the most efficient commercially available absorption

technology.... As a rough guide, double effect chillers sell for a $250/ton ($71/kW) premium

over electric chillers with COPs of 5.75 or better. The incremental cost of electricity for the

locality in question is a key factor in the life cycle costing [of electric and absorption chillers]"

One approach to a direct-fired, triple-effect absorption chiller seeks to improve efficiency by operating

a direct-fired generator at a temperature high enough so that the condenser waste heat and the

absorber waste heat from this high temperature stage are enough to operate two lower stages of

conventional LiBr single-stage absorbers (Kujak 1996). A new water/absorption brine pair with new

corrosion inhibitors and chemical performance additives are required for the upper stage of this unit.

As with natural gas engine-driven chillers, primary energy efficiencies (gas-input-based COPs) are

used in characterizing the performance of absorption chillers and these are not directly comparable to

the COPs listed for electrically driven chillers. Gas-fired COPs do not include electrical use for pumps,

fans, or blowers and may or may not include combustion losses.

Theoretical efficiencies of single-effect absorption chillers are not given. Although cooling COPs of

1.2 to 1.3 have actually been achieved for double-effect chillers, this serves as a reasonable theoretical

limit because of the very large heat exchanger surfaces required; COPs of 2.0 have been reported, but

these are based on 0.6EC ÄT's (1EF) in the heat exchangers instead of the 6EC (10EF) ÄT's for

realistic surface areas (DeVault 1994). In speaking of the triple-effect absorption chiller, Kohler says:

"Assuming 'practical' amounts of heat exchange surface and 'reasonable' maximum operating

temperatures, various researchers have claimed full load COPs from under 1.3 to over 1.8.

Whether or not machines with COPs over 1.8 are practical remains to be seen as several

obstacles to their construction exist. These COPs do not account for combustion losses in the

case of direct fired equipment." (1993)
61

Actual COPs of single-effect, indirect fired chillers typically peak around 0.7 with COPs of 1.2 widely

available for steam-fired double-effect chillers (Kohler 1993, AGCC 1996). COPs for double-effect

direct fired absorption chillers are on the order of 1.0. The lower COP for the direct-fired chillers result

from losses in the combustion process. The projected efficiency for a direct-fired, triple-effect

absorption concept whose market introduction is scheduled for 1999, is 1.45 COP (Kujak 1996).

Absorption chillers are a viable technology, favored under certain economic circumstances.

Three U.S. HVAC companies are developing direct-fired, triple-effect absorption concepts (DeVault

1997). The efficiencies which are ultimately achieved in commercially available equipment depend on

technology and market conditions. Factors such as relative costs of natural gas and electricity, peak

load charges, and equipment first costs enter into the chiller selection process. Absorption chillers are

common in Japan where electricity costs reflect the expense of investing in additional generating

capacity. They are used to a lesser extent in the U.S., often in combination with centrifugal chillers and

controlled in such a way as to minimize operating costs. 

5.4 ASSUMPTIONS

5.4.1 Carbon Dioxide Emission Rates From Power Plants

National averages are used for the power plant emission rates in these calculations. The

average electrical power plant emission rates are 0.650 kg CO2/kWh for the United States, 0.470 kg

CO2/kWh for Europe, and 0.473 kg CO2/kWh for Japan , Appendix A. These emission rates are

compiled from open literature data rather than calculated from the fundamental heat content of fuels,

fuel mix used in power production, plant efficiencies, and transportation and distribution losses.

The heat content and carbon dioxide emission rate used for natural gas were 38,200 kJ/m3

and 51.1 g/MJ, respectively. A 96.5% distribution efficiency was assumed for natural gas which raised

the CO2 emission rate to 53.0 g/MJ (55.9 g CO2/1000Btu) at its point of use (EIA, 1997).

5.4.2 Chiller-Specific Assumptions

Analysis assumptions used for calculating the direct contribution to TEWI are summarized in

Table 9.

5.4.2.1 Chiller Capacities

According to information reported to ARI, centrifugal chillers have capacities ranging from 260

kW to 28,000 kW (75 tons to 8,000 tons). The weighted average capacity of the roughly 10,000

chillers produced in the United States for domestic and foreign consumption in 1995 was 1,670 kW

(475 tons). Chillers in the range from 350 kW to 2,100 kW accounted for 79% of the chillers

produced in 1995(Hourahan 1996a). 
62

Chiller

Refrigerant Charge

Annual Emission Rate

 (percent of charge / kg/y)

(kg/kW) (kg)

0.5%

(kg/y)

2%

(kg/y)

4%

(kg/y)

1200 kW Screw or Centrifugal Chiller

 CFC-11

 CFC-12

 HCFC-123

 HFC-134a

 HCFC-22

 R-717

0.28

0.35

0.30

0.23

0.23

0.13

336 

420 

360 

276 

276 

156 

1.68

2.10

1.80

1.38

1.38

0.78

6.72

8.40

7.20

5.52

5.52

3.12

13.4 

16.8 

14.4 

11.4 

11.0 

6.2 

3500 kW Screw or Centrifugal Chiller

 CFC-11

 CFC-12

 HCFC-123

 HFC-134a

 HCFC-22

 R-717

0.28

0.35

0.30

0.23

0.23

0.13

980 

1225 

1050 

805 

805 

455 

4.90

6.12

5.25

4.03

4.03

2.27

19.6

24.5

21.0

16.1

16.1

9.1

39.2 

49.0 

42.0 

32.2 

32.2 

18.2 

Note: additional data for computing direct effect includes equipment lifetime of 30 years,

refrigerant GWPs, and assumed end-of-life loss (refrigerant lost when the equipment is

retired).

Table 9. Refrigerant emissions for 1200 kW (350 refrigeration ton) and 3500 kW (1000

refrigeration ton) chillers.

5.4.2.2 Refrigerant Charge

Refrigerant charge varies by chiller size, refrigerant, machine vintage, compressor type,

manufacturer, heat exchanger options, and other variables. The refrigerant charge used for these

calculations was a composite kg/kW value for screw and centrifugal compressors integrated over the

size range of interest. Individual sources for these values were 1) those published in the original TEWI

report, 2) those estimated by an Ad-Hoc subcommittee formed by ARI member companies, and 3)

those published in the 1995 UNEP Report (Fischer 1991, Hourahan 1996a, UNEP 1995). 

5.4.2.3 Refrigerant Loss Rates – End of Life Charge Loss

Annual loss rates for new chillers were estimated by the ARI Ad-Hoc committee at 0.5% of the

total charge. To allow for large, catastrophic charge losses and some loss as a result of routine service,

TEWIs were also calculated for 1.0%, 2.0%, and 4.0% annual rates. In the United States, losses from

refrigeration equipment have been dramatically reduced in response to the Clean Air Act (US 1993). 

One of the provisions of this act requires the “prompt” or “timely” repair of leaks showing an annual
63

Equipment

1200 kW (350 RT) 3500 kw (1000 RT)

1996 2005 1996 2005

Screw Chillers

 HCFC-22

 HFC-134a

 R-717 (ammonia)

0.56

0.60

0.56

0.54

0.54

0.54

0.60

NA

0.59

0.58

NA

0.57

Centrifugal Chillers

 HCFC-22

 HFC-134a

 HCFC-123

0.59

0.56

0.52

0.53

0.52

0.47

0.54

0.54

0.47

0.48

0.48

0.45

Table 10. Chiller integrated part load values (IPLVs) for analysis of commercial

chillers (kW/RT).

loss of more than 15% of the charge in systems containing over 50 pounds (23 kg) of refrigerant. The

ARI committee estimated that approximately 5% of the original charge would be lost at the end of

chiller life, reflecting current refrigerant prices and recovery, reclaim, and recycling laws.

5.4.2.4 Chiller Life

Chiller lifetime used in this report, 30 years, is the same as that used in the previous two studies

and is consistent with estimates provided by ARI. The chiller life assumption affects estimates for

direct TEWIs because refrigerant loss rates are expressed as a percent of charge lost per year. The

indirect TEWI is also impacted because CO2 emissions resulting for fossil energy consumption used

for electricity or gas to operate the chiller are based on annual loads multiplied by the number of years

of life. 

5.4.3 Chiller Energy Consumption Calculations

5.4.3.1 U.S. Chiller Calculations

Chiller performance was expressed in terms of Integrated Part Load Value (IPLV) as defined

by ARI Standards 550-92 for centrifugal and rotary equipment and 560-92 for absorption equipment. 

IPLVs were chosen for this work rather than Rated Full Load Performance and Integrated Full Load

Hours because this equipment operates at part load most of the time, and the IPLVs purport to give a

more accurate indication of chiller performance under part-load conditions (ARI #550-92 1992) (ARI

#560-92 1992). Table 10 summarizes the IPLV used to calculate TEWIs for this report. These IPLVs,

which are listed for the

best new equipment

currently available and

for equipment

projected to be on the

market by the year

2005, were taken from

estimates provided by

ARI members, an

equipment guide

published by the

American Gas Cooling

Center, and a listing of

currently available

electric chiller models

(Hourahan 1996b,

AGCC 1996, E Source

1995).
64

Chiller Parameter HCFC-123 HFC-134a

Refrigerant

 charge (kg/kW)

 annual make-up rate (% of charge)

 end-of-life recovery rate (% of charge)

0.24

1%

95%

0.24

0.5%

95%

Cooling Capacity

 kW

 refrigeration tons (RT)

1055

300

1055

300

Efficiency

 COP

 kW/RT

5.0

0.70

5.0

0.70

Annual Operating Rate (full-load hours) 700 700

Auxiliary Electricity Demand

(cooling tower and cooling water pumps) 20% 20%

Equipment Lifetime (years) 25 25

Table 11. Assumptions for Japanese chiller calculations.

5.4.3.2 Japanese Chiller Calculation

Some TEWI calculations were made on current electric chiller options available in Japan based on a

unique set of parameters provided by K.Hara (1996a) which is summarized in Table 11. These

calculations used full load rating performance efficiencies and annual full load hours for a Tokyo

location.

 

5.5 METHODOLOGY

Direct and indirect global

warming contributions in kg of CO2

are calculated and combined to

estimate total TEWI. Direct TEWIs 

are calculated from estimates of

refrigerant losses. The total

refrigerant charge, measured in kg, 

is multiplied by the annual make-up

rate and the lifetime of the unit in

years. This is added to the number

or kg of refrigerant lost when the

chiller is scrapped or

decommissioned and multiplied by

the 100 year ITH GWP, whose

units are in equivalent kg of CO2

per kg of refrigerant, to get the

direct TEWI.

Indirect TEWI results from

the CO2 released to the atmosphere

as a result of fuel or electrical

energy use by the chiller over its useful lifetime. An estimate of annual energy use for the chiller is based

on its efficiency and the load. This is multiplied by the lifetime of the chiller and an appropriate factor for

converting this energy into the kg of CO 2 released while providing that energy. The units of indirect

TEWI are also kg of CO2. Published equipment rating efficiencies in (kW/ton or gas COPs), IPLVs, an

estimate of annual operating hours, an equipment lifetime, and power plant CO2 emission rates per

kilowatt of delivered electricity or the kg CO2/1000 Btu value for gas are used to calculate indirect

TEWIs.

IPLVs with the total number of annual cooling hours estimated for an office building in Atlanta

(2125 hr/yr) are used for the bulk of the TEWI calculations presented in this section. 

An additional electrical consumption burden was added to all the chillers for energy used by

cooling tower pumps, chilled water pumps, and fans (air cooled chillers were not considered). A unified
65

method to quantify the electrical energy associated with heat rejection equipment was used which is

based on the quantity of heat that must be rejected by each of the chiller technologies. The rationale and

supporting assumptions for this method are given in Appendix F. 

Auxiliary electric loads of 0.035 and 0.046 kW/ton, respectively, for controls and solution

pumps were added to the direct-fired double- and triple-effect absorption chiller indirect TEWI

calculations based on data from the AGCC Natural Gas Cooling Equipment Guide (1996) and

preliminary specifications for the triple effect product (Fischer 1994). An IPLV of 1.50 COP was used

for the triple-effect chiller based upon a 1.45 rating COP and estimates obtained from ARI member

companies. No auxiliary electrical burden (other than that assigned for cooling tower operation) was

given to natural gas engine-driven chillers.

No additional electrical burden was assigned for ventilation or air movement energies

associated with fan coil units or central air handlers in buildings with centralized chillers. The capacity,

electric demand, and energy use of these subsystems are comparable for electric and gas chillers of the

same cooling capacity. In making energy or TEWI comparisons between chillers and large unitary airconditioning equipment, additional energy use should be assigned to chiller options reflecting energy

required for air circulation within the building. HVAC fan energy consumption ranges from 32 to 43

kWh/m (3 to 4 kWh/ft ) per year, which amounts to 10% to 15% of the total building electrical energy 2 2

use and 20% to 40% of the energy used for heating and cooling (McLain 1988).

5.6 RESULTS

TEWI results for large electric driven chillers in the U. S. are summarized in Figs. 20 and 21.

Two chiller capacity ranges were used for these estimates, 350 refrigeration tons (1,200 kW) and 1000

refrigeration tons (3,500 kW). Estimates of TEWI were made for “best” electric driven chiller

equipment available in 1996 and improved chillers likely to be available by 2005, Figs. 20 and 21.

Improvements in chiller performance over this time range were estimated by ARI member companies

that manufacture this equipment (Hourahan 1996b). 

Segmented bar lengths are shown for the indirect TEWI which was calculated from the chiller

IPLV, its capacity, the annual operating hours for an Atlanta office building (2125 hr/yr), the average

annual CO emission rate for U.S. electricity production including transmission and distribution losses 2

(0.65 kg CO2/kWh), the kg of CO2/1000 Btu of delivered natural gas (where appropriate), and a 30

year lifetime, as explained in the Methodology section. Shorter segments on the ends of these bars

show varying leak rate scenarios for these chillers, 0%, 0.5%, 1.0%, 2.0% and 4.0% (annual

refrigerant leak or make-up rates taken as a percentage of the total chiller charge). In addition we

assume a 5% loss of charge when the chiller is eventually scrapped. Some CFC-11 and CFC-12

chiller data of 1993 vintage equipment efficiencies are shown on Figs. 20 and 21 for comparisons with

the CFC-free technologies and because many chillers are still operating with these refrigerants. IPLVs

of 0.58 for CFC-11 and 0.59 for CFC-12 and an annual leak rate of 4.0% were assumed for this

equipment. Tabulated results used for these figures are given in Appendix G. 

Figure 22 shows the sensitivity of TEWI values to the CO emission rates from electric power 2
CFC-11

CFC-12

HCFC-123

HFC-134a

HCFC-22

HCFC-22

R-717

HCFC-123

HFC-134a

HCFC-22

HCFC-22

R-717

0 10,000,000 20,000,000 30,000,000 40,000,000 50,000,000

indirect effect

end-of-life

0.5%

1%

2%

4%

direct effect

increments for make-up rates

1993

Eff.

1996

Efficiencies

2005

Efficiencies

centrifugal

centrifugal

centrifugal

centrifugal

centrifugal

centrifugal

centrifugal

centrifugal

screw

screw

screw

screw

TEWI (kg CO ) 2

CFC-11

CFC-12

HCFC-123

HFC-134a

HCFC-22

R-717

HCFC-123

HFC-134a

HCFC-22

R-717

0 5,000,000 10,000,000 15,000,000 20,000,000

TEWI (kg CO )

indirect effect

end-of-life

0.5%

1%

2%

4%

direct effect

increments for make-up rates

centrifugal

centrifugal

centrifugal

centrifugal

centrifugal

centrifugal

screw

screw

screw

screw

1993

Eff.

1996

Efficiencies

2005

Efficiencies

2

66

Figure 21. TEWI for 3500 kW (1000 ton) electric driven chillers in an Atlanta office application.

Figure 20. TEWI for 1200 kW (350 ton) electric driven chillers in an Atlanta.
Power Plant Emission Rate (kg CO /kWh)

TEWI (kg CO )

2

2

0.00 0.20 0.40 0.60 0.80 1.00 1.20

0

10,000,000

20,000,000

30,000,000

40,000,000

50,000,000

HFC-134a

Centrifugal

Engine Driven

HCFC-22 Screw

HCFC-123

Centrifugal

U.S. Average

0.65

67

Figure 22. TEWI for 3500 kW (1000 ton) chillers in Atlanta across a range of CO2 emission rates

(2005 efficiencies).

plants. Data for 3500 kW (1000 ton) equipment likely to be on the market in 2005 are presented.

Values chosen for the annual operating hours have a proportional effect on the indirect TEWI.

Figure 23 shows a plot of TEWI for popular 3500 kW (1000 ton) chiller options as a function the

annual operating hours. While indirect TEWI is directly related to the number of operating hours, the

absolute value of direct TEWI remains fairly constant. Longer chiller operating hours minimize the direct

TEWI contribution. 

Figure 24 shows TEWIs for the electric chiller options suggested by Hara and JRAIA for the

Japanese market. Power plant CO2/kWh emission rates, equipment performance characteristics, and

operating conditions suggested by Hara were used for the calculations (Hara 1996a). The large

difference in total TEWI between Japanese and U.S. chillers is due to differences in hours of operation

rather than any fundamental difference in technology.

No unique chiller calculations were carried out for Europe because large centralized chiller

installations are not as popular in European countries, as indicated by the breakdown of installed chiller

capacity presented in the introduction. To perform a calculation for a European application, chiller

efficiencies, the chiller capacity, annual cooling hours, and the power plant CO2 emission rate have to

be substituted for those used for North American calculations. Using the same performance factors for

North American and European chillers is reasonable because chiller operating conditions do not vary

widely (Fischer 1991).
0 1,000 2,000 3,000

0

10,000,000

20,000,000

30,000,000

40,000,000

Annual Chiller Operating Hours (h/y)

TEWI (kg CO ) 2

HCFC-123 Centrifugal

(0.45 kW/Rton)

HFC-134a Centrifugal

(0.48 kW/Rton)

HFC-134a Engine-Driven

Centrifugal (COP=2.40)

HCFC-123

HFC-134a

0 1,000,000 2,000,000 3,000,000

Indirect Effect (Electricity)

Direct Effect (Refrigerant)

TEWI (kg CO 2)

68

Figure 23. TEWI for 3500 kW (1000 ton) chillers in Atlanta across a range of operating hours (2005

efficiency levels)

Figure 24. TEWI for 1055 kW (300 ton) chillers in Tokyo, Japan.

The TEWI of 1200 kW (350 RT) and 3500 kw (1000 RT) engine driven and absorption

chillers based on the efficiency data in Table 12 are presented in Figs. 25 and 26. These results show

the low indirect effect from

energy use for engine driven

chillers as well as the

significant reductions in

indirect effects as the

technology develops from

1996 double effect

absorption chillers to triple

effect chillers. The results for

engine driven chillers using

HFC-134a or HCFC-22

show a possible direct effect

from refrigerant emissions at

end-of-product life and annual

losses, but it remains small

even under the highest

refrigerant loss rates
0 5,000,000 10,000,000 15,000,000 20,000,000 25,000,000

HCFC-22 and

Double Effect

Engine

Driven

Direct

Fired

Engine

Driven

Direct

Fired

Direct

Fired

1996

Efficiencies

2005

Efficiencies

Absorption

Triple Effect

Absorption

Double Effect

Absorption

HFC-134a

HCFC-22 and

HFC-134a

TEWI (kg CO ) 2

indirect effect

end-of-life

0.5%

1%

2%

4%

direct effect

increments for make-up rates

0 10,000,000 20,000,000 30,000,000 40,000,000 50,000,000 60,000,000 70,000,000

HFC-134a

Double Effect

HCFC-22

HFC-134a

Double Effect

Triple Effect

Absorption

Absorption

Absorption Direct

Fired

Engine

Driven

1996

Efficiencies

2005

Efficiencies

Engine

Driven

Engine

Driven

Direct

Fired

Direct

Fired

indirect effect

end-of-life

0.5%

1%

2%

4%

direct effect

increments for make-up rates

TEWI (kg CO ) 2

69

Figure 25. TEWI for 1200 kW (350 ton) gas-fired chillers in Atlanta, Georgia USA

Figure 26. TEWI for 3500 kW (1000 ton) gas-fired chillers in Atlanta, Georgia USA.
70

Equipment

1200 kW (350 RT) 3500 kw (1000 RT)

1996 2005 1996 2005

Engine Driven Chillers

 HCFC-22 Screw

 HFC-134a Centrifugal

1.95

1.95

2.10

2.10

2.30

2.30

2.40

2.40

Absorption Chillers

 Direct fired, double effect

 Direct fired, triple effect

1.07

NA

1.15

1.50

1.07

NA

1.15

1.50

gas COPs are given for part-load gas chiller IPLVs, defined as the delivered cooling

effect divided by the heat content (HHV) of the gas consumed (no waste heat

recovery); NA signifies equipment/refrigerant combination is not available.

Table 12. Chiller integrated part load values (IPLVs) for analysis of commercial

chillers (kW/RT).

considered.

5.7 DISCUSSION

Figures 20

and 21 illustrate the

environmental

benefits of HCFC123 as a refrigerant

for this application.

It does contain

chlorine which gives

it an ODP rating of

0.02. However, it

shows the best

cycle efficiency of all the alternative refrigerants currently being used in large chillers with flooded

evaporators, and it has the lowest GWP (93) of all the non-flammable refrigerants. The one HFC

alternative which has been suggested and preliminarily evaluated for use in low pressure chillers is

HFC-245ca which has a GWP estimated at 560 and is slightly flammable in humid air.

TEWI for chillers is unlikely to improve markedly from 1996 to 2005 because annual

refrigerant leakage and recovery rates are reaching their lowest practical limits, and dramatic

improvements in chiller cycle efficiency within the last 10 years has left very little room for substantial

improvements in this area (Glamm 1996).

Natural gas engine-driven chillers with COPs or IPLVs of 2.0 or greater have total TEWIs

indistinguishable from comparable electric equipment in the United States when an annual average CO2

emission rate of 0.65 kg CO /kWh is used. 2

Several industry experts have suggested that the TEWI definition needs to include all the energy

consumed in extraction, transportation, and processing of the fuel before it arrives at the power plant. It

has also been suggested that time-of-day CO2/kWh should be used for calculation of air conditioning

TEWIs because different plants are used to meet the “peak loads” associated with summer air

conditioning use (Beggs 1996). These two questions are partially addressed in Appendix A.

Direct-fired, triple-effect absorption shows a substantial TEWI advantage over current directfired, double effect equipment, especially for the lower, 350 ton, capacity ranges.

Cooling tower burdens for chiller technologies can be decreased by increasing the temperature

difference for the condensing water from 5.6E C (10E F) to 8.3E C (15E F). A 5.6E C (10E F) )T

was used for both the electrical and gas-driven technologies because the condenser or

absorber/condenser water flow rates specified in ARI 550-92 and 560-92 Standards are consistent

with a 5.6E C (10E F) condenser )T. The new 560 Standard which is out for comment calls for a

lower absorber/condenser water flow rate with would raise the condenser water )T for direct fired

absorption. Industry sources indicate that gas and newer, more efficient electric chillers are being
75% Heat Recovery

50% Heat Recovery

25% Heat Recovery

Engine Chiller (HFC-134a)

0 5,000,000 10,000,000 15,000,000 20,000,000 25,000,000 30,000,000

TEWI (kg CO 2)

Indirect Effects (Energy)

Gas

Electricity

Direct Effects (Refrigerant)

HFC-134a

71

Figure 27. Benefits of heat recovery for chiller plants.

installed in the field with condenser flow rates considerably less than those specified in the appropriate

ARI Standard to help minimize life cycle costs.

No equitable method was found which permitted including a heat recovery benefit from engine

driven equipment and/or a CO2 credit for direct-fired absorption equipment. These b/enefits are quite

site specific and depend heavily on the type of building, its geographic location, and the type of activity

in the building. Heat recovery is most beneficial when it can be used to avoid consumption of additional

fossil fuel for heating or regeneration or when there is a fixed ratio of cooling and heating demands in an

application. Figure 27 shows the relative TEWI reductions possible with fixed percentages of waste

heat recovery.

5.8 CONCLUSIONS

From the data presented in Figs. 20 and 21, it is clear that TEWI values for new electric chillers

have improved by about 20 to 30% since the early 1990’s. Approximately 75% of this improvement

is due to a change from the CFC refrigerants which substantially reduced the direct TEWI. Substantial

improvements in refrigerant handling and recovery practices have also contributed significantly to this

direct TEWI decrease (Smithart 1996). Consequently, there is no significant TEWI advantage from

using zero GWP refrigerants in a chiller.
72

The choice of working fluid among those currently being used as refrigerants for vapor

compression chillers makes only a minor difference in direct TEWI. This is especially true now that low

leakage rates, low service loss rates, and high efficiency purge systems have been implemented on this

equipment. However, the differences in chiller efficiencies for various refrigerant options can make

significant differences in the indirect TEWI. The indirect TEWI term remains dominant in this high

energy use, long-life application with HCFC-123 machines showing a 10 to 15 percent advantage in

total TEWI.

Longer annual operating hours decrease the relative direct TEWI contributions of refrigerants.

Direct-fired absorption chillers can be a preferred choice where they have greater value in their

ability to use lower cost fuels and/or provide high quality waste heat for a useful application. In many

chiller installations, absorption equipment is used for a much shorter time to handle peak air conditioning

loads and avoid high electrical peak demand charges. 

Gas fired technologies can show lower TEWI when simultaneous heating and cooling are

required and heat recovery is used. Double bundle condensers used with electric chillers for waste heat

recovery can also result in lower TEWI values.
73

6. COMMERCIAL REFRIGERATION SYSTEMS

6.1 INTRODUCTION

Commercial refrigeration is very a broad category which can include many diverse applications.

Among these are supermarket refrigeration systems using twenty or more display cases for dairy

products, meat and fish, frozen foods, and ice cream. These systems typically employ large racks of

compressors in a machine room removed from the sales floor and condensers on the roof or in back of

the store. Commercial refrigeration also includes “stand-alone” display cases at convenience and

grocery stores that contain the entire mechanical package (e.g. compressor, condenser, evaporator),

walk-in refrigerators and freezers at restaurants and hospitals, and ice makers in hotels, among others.

Supermarket refrigeration systems, unlike the other applications mentioned, have long refrigerant lines

which are susceptible to leaks and require very high refrigerant charges. Consequently, the direct effect

of refrigerant emissions can be very high for supermarkets with distributed cases and long refrigerant

lines; on the other hand, stand-alone display cases, walk-in freezers, and ice makers have smaller

charges and much lower emission rates and are more like household refrigerators and unitary air

conditioning in this regard. This study focuses exclusively on supermarket systems, and it includes the

current standard technology as well as two alternative refrigeration systems.

6.1.1 Direct Expansion Systems

The traditional design for supermarket refrigeration systems is referred to as “direct expansion”

because the high pressure refrigerant is circulated throughout the store directly to display cases where it

is expanded and absorbs heat (maintaining the cold product temperature). Thousands of meters of

piping can be required to connect the display cases to the remote compressor. The large internal

volumes resulting from all of the piping require very large refrigerant charges, while the number of

welded and brazed joints provide a great many sources for leaks. Also, the refrigerant itself must meet

very strict safety requirements covering toxicity and flammability since it is being circulated throughout

the retail sales floor and is in close proximity to customers and employees.

6.1.2 Secondary Loop Systems

Efforts to reduce the refrigerant charge required by direct expansion systems led to some test

installations of refrigeration systems which use secondary heat transfer loops to pump cold brine

solutions to the display cases on the sales floor. Replacing the piping of refrigerant with piping of brine

allows the refrigeration system to operate with a refrigerant charge that can be as small as 10% of what

is required for a comparable direct expansion system. Besides reducing the charge, secondary loop

systems are less prone to leak because there are fewer welded and brazed joints in pipes containing

refrigerant. A second potential advantage of these systems results from isolating the refrigerant from the
74

sales floor; the refrigerant is no longer in proximity with the customers and less restrictive health and

safety requirements may permit the use of flammable or toxic refrigerants. The introduction of a

secondary loop, however, introduces a )T between the refrigerant and the brine which may require a

lower evaporating temperature than is required for direct expansion systems. The lower evaporating

temperature means lower relative efficiency and higher power consumption. Secondary loop systems

also require a pump to circulate the brine, which is another addition to power consumption and affects

the amount of heat that must be rejected by the condensers.

6.1.3 Distributed Systems

A second approach to reducing refrigerant charge and emissions has resulted in the

development and marketing of what is termed here as “distributed systems.” In these installations, one

or more compressors on small racks are distributed throughout the supermarket near the display cases

they are serving. The compressors can be in attractive cabinets on the sales floor or they can be

installed in unused space in back rooms or on top of walk in freezers and refrigerators (Broccard

1995). Commercially available distributed systems rely on a water loop to connect all of the

compressor installations with a single cooling unit (e.g. cooling tower, evaporative cooler) on the roof or

outside behind the store to reject the waste heat of the system. As with the secondary loop systems, this

water loop introduces an additional )T and thermodynamic loss and pumping power not associated

with direct expansion systems.

While TEWI are calculated and presented for all three classes of equipment, direct expansion,

secondary loops, and distributed systems, insufficient information is provided to make absolute

comparisons between the different technologies. Care was taken to select thermodynamic cycle

conditions that were representative so that energy use calculations could then be compared; however,

no effort was directed toward installation and maintenance costs comparisons for the three systems.

This is a significant omission which needs to be covered before concluding that one system or another is

“better” than another.

6.2 ASSUMPTIONS

6.2.1 Temperature Ranges

Refrigeration needs for supermarkets break down into three broad categories:

! high temperature refrigeration for air conditioning and cooling of prep rooms, typically

providing cold air around 10EC (50EF),

! medium temperature refrigeration for meat and fish and dairy cases and walk-in coolers

for meats and produce, with air temperatures from -2E to 7EC (28E to 45EF), and

! low temperature refrigeration for freezers and ice cream cases and walk-ins, -18E to -
75

32EC (0E to -25EF).

High temperature refrigeration is not considered in this study, since any analysis needs to consider

interactions between the air conditioning system, the refrigerated display cases, and moist air entering

the store. That degree of analysis is beyond what can be done in this study. Medium temperature

refrigeration is simplified by assuming a single fixed evaporating temperature for all the display cases; -

7EC (20EF) for direct expansion systems. A single evaporating temperature is also used to characterize

all of the low temperature refrigeration; -32EC (-25EF) for direct expansion systems. Evaporating

temperatures for the secondary loop calculations are assumed to be 3EC (5.5EF) lower than the

corresponding temperatures for direct expansion systems (Likes 1996).

The high-side temperatures where heat is rejected depend on the outdoor ambient temperature,

naturally, as well as the type of heat rejection unit used; an evaporative cooler will have lower

temperatures than will an air-cooled unit. These calculations are based on a single condensing

temperature, 36EC (97EF), for an entire year of operation and the same temperature is used for

Europe, Japan, and North America. The condensing temperature for the distributed systems is assumed

to be 3EC (5.5EF) higher to accommodate the condenser water loop mentioned earlier (Ares 1996).

These assumptions are significant simplifications that vastly reduce the amount of work required in this

analysis, but as a result small differences in calculated energy use should not be viewed as being

significant.

6.2.2 Refrigeration Loads

Baseline refrigeration loads are assumed for Europe, Japan, and North America. It is assumed

that a typical supermarket in North America requires 88 kW (300,000 Btu/h) of low temperature

refrigeration and 264 kW (900,000 Btu/h) of medium temp (Hourahan 1996). A store in Japan is

assumed to have 24 kW (82,000 Btu/h) and 127 kW (433,000 Btu/h) of low and medium temperature

loads (Hara 1996). Store loads for Europe are assumed to be 50% of those in North America; 44 kW

(150,000 Btu/h) low temp and 132 kW (450,000 Btu/h) medium temp.

6.2.3 Alternative Refrigerants

Historically, CFC-12, HCFC-22, and R-502 were used in supermarket refrigeration systems.

Each of these either has been, or is being, phased out of use as a result of the Montreal Protocol. Many

different refrigerants have been promoted as replacements for these three fluorocarbons, and

manufacturers have gravitated toward marketing equipment using only two or three of the alternative

refrigerants. The analysis presented here, however, includes more than just those two or three

alternatives in an effort to determine if any other refrigerant has a clear advantage in reducing TEWI.

Direct expansion and distributed systems are considered which include:

! five alternatives for medium temp; R-404A, R-507, R-134a, R-407A, and R-410A,

and
 ASHRAE has assigned the official designation of R-507A to the 50/50 mixture of HFC-125 1

and HFC-143a; that composition of 125/143a was commonly referred to as “R-507" prior to 1997

and is marketed in commercial products with that designation. The R-507 designation consequently is

used throughout this report.

76

! four alternatives for low temperature refrigeration; R-404A, R-507 , R-407A, and 1

R-407C.

Each of these refrigerants is also included in the analysis of secondary loop refrigeration systems, but

those calculations and results also include ammonia, R-717, as an alternative. While the use of ammonia

can be limited by building codes and ordinances at all levels of government, it is a viable alternative in

this application and secondary loop systems with ammonia are being built and used in Europe (Haaf

1996). The compressor discharge temperature for ammonia can be much higher than it is for other

refrigerants, and effort must be made in the design and application to keep temperature safely below the

point where the oil would be damaged or destroyed (Ares 1996). These are important considerations

in applying ammonia in commercial refrigeration and in comparing systems using ammonia with those

using refrigerants with lower discharge temperatures, but they are primarily cost and design

consideration and are not included in this analysis.

6.2.4 Refrigerant Charge

Assumptions relative to refrigerant charge and leakage rates are critical to any calculation of

TEWI for this application, yet these numbers are either not well known and documented or they are

widely variable. A general rule of thumb in American industry is that the number of pounds of refrigerant

in a supermarket will be from 8% to 12% of the floor space in square feet (Broccard 1995); 40% to

60% of the floor space in kg refrigerant/m . Some of the large “mega-stores” being built can have 2

refrigerant charges as high as 4000 kg (9000 lb). A recent survey showed the average store in a

supermarket chain to be about 2300 m (25,000 ft ) (Progressive Grocer 1990); the average 2 2

independent supermarket is smaller 1300 m (14,000 ft ) . 2 2

6.2.4.1 Direct Expansion

The baseline systems used to define the analysis are assumed to use 400 kg (880 lb) R-502

and 464 kg (1020 lb) HCFC-22 for North America; 200 kg (440 lb) R-502 and 232 kg (510 lb)

HCFC-22 for Europe; and 79 kg (174 lb) R-502 and 277 kg (610 lb) HCFC-22 for Europe. The

North American data correspond to an average sized store in a major supermarket chain based on the

rule of thumb mentioned earlier (Progressive Grocer 1990); European data are arbitrarily set at 50%

that of North America. A recent study by an industry trade group reported that the average

supermarket in the U.S. uses 2700 kg (6000 lb) of HFCs to provide 530 kW (150 tons) of

refrigeration (Bittner 1995); while this refrigeration load is consistent with the data used in this analysis,
77

the refrigerant charge is significantly higher. This discrepancy has not been reconciled, but may be the

result of the store size reflected by the survey and it may include refrigerant needs for air-conditioning

and prep rooms. Refrigerant charges for alternative refrigerants are estimated using the baseline charge

and the ratio of the liquid density of the alternative refrigerant at 21EC (70EF) to the density of the

baseline refrigerant (i.e. R-502 or HCFC-22).

6.2.4.2 Secondary Loop Systems

The refrigerant charge for secondary loop systems ranges from 8% to 14% of the charge for a

comparable direct expansion system (Hourahan 1996, Kruse 1993). An average value of 11% of the

corresponding direct expansion charge is used in this analysis. The selection of a refrigerant charge for

secondary systems is not crucial in the analysis because the combination of lower charge and lower

emission rates results in a relatively insignificant direct effect in the TEWI.

6.2.4.3 Distributed Systems

Refrigerant charges for distributed systems can be reduced by 75% or more from what they

would be for comparable direct expansion systems (Broccard 1995). This analysis assumes that the

refrigerant charge for distributed systems is 25% of that for direct expansion systems.

6.2.5 Refrigerant Emission Rates

6.2.5.1 Direct Expansion Systems

Historically direct expansion supermarket refrigeration systems have had emission rates of 30%

of the total charge per year or higher. These high rates were the result of using flared fittings in the

piping to expedite service work on the sales floor, mixed metal welds, thermal shocks at flared fittings

due to defrosting, and poorly or improperly supported pipes in long runs (Richey 1995). These rates

were also the consequence of inexpensive refrigerant and expensive labor. Significantly higher

refrigerant costs and enhanced environmental awareness have led to design changes in display cases

and maintenance practices that are resulting in lower emission rates (Broccard 1996). Major

supermarket chains have found that aggressive preventive maintenance programs can reduce emissions

to 10% of the charge per year or less (Richey 1995). Haaf also reported a 10% emission rate for

Europe (1996) and the Air Conditioning and Refrigeration Institute (ARI) reported current ranges of

12% to 15% of the charge which can be reduced to 4% to 8% of the charge per year in 5 to 10 years

(Hourahan 1996). Calculations are reported using the averages of the ranges given by ARI for current

and near term emission rates, 131⁄2% and 6% respectively.

6.2.5.2 Secondary Loop Systems

Secondary loop systems are known to have lower leakage rates than direct expansion systems
78

Refrigerant

Direct

Expansion

Secondary

Loops

Distributed

Systems

R-502 1.798 1.636 1.726

R-404A 1.762 1.603 1.692

R-407A 1.708 1.554 1.640

R-407C 1.654 1.505 1.588

R-507 1.762 1.603 1.692

R-717 1.638

Table 13. Low temperature refrigeration COPs.

Refrigerant

Direct

Expansion

Secondary

Loop

Distribute

d Systems

HCFC-22 3.605 3.281 3.389

HFC-134a 3.317 3.018 3.118

R-404A 3.425 3.117 3.219

R-410A 3.569 3.248 3.355

R-507 3.515 3.199 3.304

R-717 3.205

Table 14. Medium temperature refrigeration COPs.

because there are fewer brazed and welded fittings on pipes containing refrigerant and the piping

connections are more accessible to maintenance personnel for leak checking and repair. Leakage rates

are also lower because there is not a refrigerant fitting at the display case where it would be exposed to

the thermal shocks associated with defrosting the case evaporator. TEWI calculations are based on

emission rates of 4% for current equipment and 2% for near term (5 to 10 years) systems (Hourahan

1996).

6.2.5.3 Distributed Systems

Distributed systems can also have lower leakage rates than direct expansion systems, but not as

low as can be achieved in secondary loop systems. TEWI calculations for current technology are based

on a 5% per year loss rate and 2% for near term (Hourahan 1996).

6.2.6 COPs

6.2.6.1 Direct Expansion Systems

Assumptions about the relative

efficiencies of equipment using one

refrigerant or another are probably the

most controversial choices made in this

analysis because there is an apparent

advantage to one supplier’s product

over another. There are theoretical

differences between the refrigerants, and

differences have been demonstrated in

laboratory and instrumented store

testing. However, there are so many

factors involved in store operation that

the theoretical and laboratory

differences in efficiency for refrigerants

do not necessarily show up as

differences in energy consumption

comparisons between stores.

Different efficiencies, COPs, are

assigned for each alternative refrigerant

in this analysis for low and medium

temperature refrigeration based on data

reported in the open literature. First, a

baseline COP is selected for R-502 for

low temp and HCFC-22 for medium
79

temp, and then literature values for performance relative to those baselines are applied to get a COP for

each alternative (Shiflett 1994, ASW 1994, Borhanian 1994, Haaf 1996). These data are summarized

in Tables 13 and 14; COPs for ammonia, R-717, are shown only for the secondary loop systems

because it is not considered a viable choice of refrigerant for applications where the refrigerant is in

close proximity with customers or employees (note that R-502 and HCFC-22 are listed in Tables 13

and 14, but they are only used as baseline refrigerants to define COPs for the HFC mixtures and

ammonia).

6.2.6.2 Secondary Loop Systems

While there are many comparisons of the performance of HFC mixtures in direct expansion

systems, there is only a small body of information available for secondary loop systems and it is

inadequate for comparing different refrigerants. Consequently this analysis relies on a theoretical

degradation from the performance of direct expansion systems due to assumed differences in

evaporating temperatures to estimate performance of refrigerants in secondary loop systems. The

estimated COPs for secondary loop systems are also listed in Tables 13 and 14. As mentioned

previously, it is assumed that the evaporating temperatures for systems using secondary loops are 3EC

(5.5EF) lower than evaporating temperatures for direct expansion systems. The relative performance

of secondary loop and direct expansion systems were determined using cycle calculations based on:

! low temp conditions of 36EC (97EF) condensing, 2.8EC (5EF) evaporator superheat,

4.4EC (40EF) return gas, 18EC (65EF) liquid line, and -32EC (-25EF) evaporating for

direct expansion and -35EC (-30.5EF) for secondary loops to determine an average

theoretical degradation of 9% for secondary loops relative to direct expansion, and

! medium temp conditions of 36EC (97EF) condensing, 2.8EC (5EF) evaporator

superheat, 18EC (65EF) return gas, 18EC (65EF) liquid line, and -7EC (20EF)

evaporating for direct expansion and -10EC (14.5EF) for secondary loops to determine

an average theoretical degradation of 9% for secondary loops relative to direct

expansion.

These factors are applied to the COPs for direct expansion to estimate the compressor power of

secondary loop systems.

6.2.6.3 Distributed Systems

Similar calculations were performed to estimate degradation factors for distributed systems,

although in this case the 3EC (5.5EF) )T is applied to the condensing temperature of 36EC (97EF).

Theoretically, based on these calculations, low temp distributed systems will have COPs 4% lower than

direct expansion systems; medium temp COPs are 6% lower.
80

6.2.7 Compressor On-Time

Energy use calculations, as explained later, are based on compressor operation for some

estimated equivalent full-load hours. Field measurements on installed equipment and annual energy use

reported by Hara (1996) and Oas (1991) both correspond to the compressors running at full load 40%

of the time. 

6.2.8 Condenser Fan Power

Each of the alternative systems is required to meet its own corresponding set of thermal loads.

The loads are comprised of the refrigeration load of the display cases, the compressor power, and

pumping power for the secondary loop and distributed systems. The condenser fan power necessary to

reject this heat is calculated for each system based on 18.3 W per kW of heat rejected.

6.2.9 System Lifetime

The assumption on equipment lifetime is important in evaluating the results presented later on an

absolute basis, but perhaps surprisingly it is not important in the calculations. Power consumption and

refrigerant emissions are all computed on an annual basis, and the assumed system lifetime is no more

than a scaling factor applied to the annual results. This simplified analysis is a consequence of omitting

any refrigerant losses at the end of product life; it is assumed that essentially all of the refrigerant charge

is recovered when the refrigeration system is replaced. Estimates of equipment lifetime could be based

on the frequency of remodeling the store and replacing the display cases (approximately seven years) or

the approximately 20 year useful lifetime of the compressors and condensing unit (Fischer, et al 1991).

The calculations and results presented in this report are based on an equipment lifetime of 15 years

(Rosenstock 1997).

6.3 METHODOLOGY

TEWI is computed separately for low and medium temperature refrigeration for each of the

three alternative technologies using each of the alternative refrigerants. The indirect effects are

computed by estimating the annual and lifetime power consumption for the compressor and auxiliary

fans and pumps; the total power consumptions are converted to lifetime CO2 emissions using a single

regional average factor for emissions from electrical power production. Direct effects from refrigerant

emissions are computed using the annual emission factors for each technology, the equipment lifetime,

and the GWPs for each refrigerant.

6.3.1 Power Consumption

Annual power consumption is computed as the sum of estimated compressor power, pumping
81

power if any for the secondary loop or condenser loop, and condenser fan power. Power consumption

for the evaporator fans has not been included. This omission is not significant in comparing TEWI for

alternative refrigerants because including evaporator fan power affects each of the alternatives equally;

including evaporator fan power would reduce the fraction of TEWI from direct effects slightly.

6.3.1.1 Compressor

Compressor powers are calculated by dividing the assumed refrigeration load by the COP

corresponding to the application (i.e. low or medium temp; direct expansion, secondary loop,

distributed system) for each alternative refrigerant. The annual energy inputs are then determined by

multiplying these ratios of load and COP by the assumed number of equivalent full load hours, 3504

hours (i.e. 40% run-time and 8760 hours per year).

6.3.1.2 Secondary Loop Pumping Power

An effort is made to include an estimate of the secondary loop pumping power in the TEWI

calculations, but it must be recognized that while attempting to be “realistic” that this effort is at best

cursory. An accurate calculation of pumping power requires a detailed design of the secondary loop

and an economic evaluation of trade-offs between materials and construction costs and operating costs.

This analysis, while necessary for greater precision, is beyond the scope of this report. Whether good

or bad, assumptions are made for these calculations in order to include an estimate of pumping power.

These include:

! an equivalent length of piping between the display cases and the machine room

(different lengths are assumed for Europe, Japan, and North America),

! the )T in the secondary loop acceptable at the display case to provide the cooling,

! the numbers of equal parallel loops servicing the low temperature and the medium

temperature display cases, and

! the diameters of the piping used.

The allowable )T and the refrigeration load determine the flow rate necessary to meet the load on each

parallel loop. The equivalent length, diameter, and flow rate determine the pressure drop and pumping

power (for given coolant properties, several coolants were considered including both organic and

inorganic salt solutions). In reality, the pipe dimensions and number of parallel circuits would be

determined by cost factors; in this analysis these variables were chosen to approximate a European

design value for a )P of 2 to 2.5 bar (30 to 36 psi) (Haaf 1996). Systems were selected for North

America and Japan which gave pumping powers which are about the same fraction of compressor

power as determined for the European systems. The details for these calculations are included in
82

Appendix H.

6.3.1.3 Distributed System Condenser Loop Pumping Power

The calculational procedure for estimating pumping power for the condenser water loop of

distributed systems is similar to that used for computing the pumping power for secondary loops. 

6.3.2 Direct Effect

The portion of TEWI for each refrigerant in each application is a straightforward calculation

using the equipment lifetime, the annual emission rate (as a percentage of refrigerant charge), the

refrigerant charge, and the GWP for each refrigerant. It is assumed that essentially all of the refrigerant

is recovered at the end of product life. 

6.4 RESULTS

Tabular results and bar charts of all of the calculated TEWI are listed in Appendix I. One set of

results is shown in Fig. 28 which shows groups of bars for the three alternative technologies; direct

expansion systems, secondary loops, and distributed systems. Within each group, there are separate

bars for the alternative refrigerants considered, and each bar is broken into three segments. Each of

these segments corresponds to one factor in the global warming from supermarket refrigeration:

! the leftmost, heavily shaded segment, corresponds to indirect global warming from CO2

emissions resulting from energy use,

! the center segment (where it can be seen) corresponds to the direct global warming

effects of refrigerant emissions at the reduced rates that may be achieved in 5 to 10

years (2002 to 2007), and

! the lightly shaded segment on the right also due to refrigerant emissions, it is the

increment over the near term rate due to the current (1997) higher emission rate.

The total length of each bar, then reflects the TEWI at current emission rates and the left two segments

the TEWI at the lower future emission rates.

In a broad sense the information contained in Fig. 28 and Appendix I represent a reasonable

comparison of the three technologies and alternative refrigerants considered, but it relies on many

simplifying assumptions that affect the results. First, all of the energy use calculations are based on a

single, average condensing temperature to estimate annual energy use. Second, although the assumed

)Ts and calculated pumping powers for the secondary loop systems are comparable to what has been
R-404A

R-507

R-134a

R-410A

R-404A

R-507

R-134a

R-410A

R-717

R-404A

R-507

R-134a

R-410A

0 500,000 1,000,000 1,500,000 2,000,000 2,500,000

Indirect (Energy)

Direct (minimum)

Direct (incremental)

TEWI (kg CO )2

DIRECT

EXPANSION

SECONDARYLOOP

DISTRIBUTED

SYSTEMS

83

Figure 28. Medium temperature refrigeration in Europe.

experienced in actual installations (Boyko 1997), poor selection of heat transfer fluid could lead to

higher pumping powers (Ares 1997, Boyko 1997) and design optimization for energy efficiency could

reduce or eliminate the increases in compressor power (Boyko 1997, Gage 1997). While a more

rigorous analysis and more optimal designs can lead to more precision in the estimated TEWI for each

alternative, they are unlikely to change the relative comparisons shown for the three technologies or the

different refrigerants.

6.5 CONCLUSIONS

Several conclusions can be drawn from the data presented in Appendix I, but it is essential to

remember that a major factor has been omitted in this analysis. There are different equipment,

installation, and maintenance costs associated with each of the three technologies presented (not to

mention factors affecting the quality of the refrigerated products). The figures and tables show that:

! direct expansion systems: reduction of the emission rates from the currently

achievable rates to what is believed to be achievable in 5 to 10 years (2002 to 2007)

represents a significant reduction in TEWI,

! direct expansion systems: the major factor affecting differences in TEWI between
84

alternative refrigerants is the GWP of each refrigerant,

! secondary loop systems: significant reductions in TEWI from current levels might be

able to be achieved using secondary loop systems, but with a possible 13 to 26%

increase in energy use relative to direct expansion systems (as shown in Appendix I). A

more detailed analysis is required that includes construction, maintenance, and energy

costs in order to evaluate the value of this technology in reducing TEWI, based on the

assumptions used in this report,

! secondary loop systems: the use of a zero GWP refrigerant, ammonia (R-717), in a

secondary loop system could result in lower TEWI than using a fluorocarbon

refrigerant; TEWI for the fluorocarbons are 0 to 15% higher than the TEWI with

ammonia, 

! distributed systems: significant reductions in TEWI may also be achieved from current

levels for direct expansion through the use of distributed refrigeration systems, but with

an 8 to 15% increase in energy use. Evaluations of the equipment, installation, and

operating costs and the possible loss of useful space on the sales floor is necessary to

assess the value of this technology in reducing greenhouse gas emissions, and

! secondary loop and distributed systems: differences in TEWI for distributed and

secondary loop systems are small, regardless of refrigerant choice, and in most

instances are insignificant.

Three different methods of achieving lower TEWI are shown in these figures and information

concerning the costs of each approach is needed to determine its merits as a technology for reducing

emissions of greenhouse gases. The costs in equipment and preventive maintenance necessary to

reduce refrigerant emissions from the currently achievable levels to those believed possible in 5 to 10

years are different from the costs associated with secondary loop systems. Unfortunately, a cost based

analysis is beyond the scope of the current project and it is only possible to conclude that alternative

technologies have the potential for dramatic reductions in TEWI for this application. 
85

7. AUTOMOBILE AIR CONDITIONERS

7.1 INTRODUCTION

Automobile air conditioning was identified in previous studies as one of the few applications of

fluorocarbon refrigerants where the direct effects of refrigerant emissions is a significant fraction of the

TEWI (Fischer, et al 1991, Fischer, et al 1994). While not contesting that conclusion, those studies

have been criticized because they relied on efficiency data at a single design point and an estimated

equivalent full-load hours of operation. Those two simplifying assumptions cannot account for varying

performance over a range of operating conditions or the effects of different climates. The present

analysis addresses these concerns; it incorporates efficiency differences across a wide range of

operating conditions, regional variations in outdoor ambient temperature, and changes in air-condition

on-time with outdoor temperature. While not perfect, this more detailed analysis should provide a

better comparison of fundamentally different cooling cycles.

Three different cooling systems are considered in this analysis; a conventional system using

HFC-134a, a conventional system employing a hydrocarbon refrigerant, and the transcritical CO2

(R-744) system. All three systems are assumed to consist of an orifice tube, a positive displacement

compressor with a constant displacement, an accumulator, and a plate fin evaporator. The HFC-134a

and hydrocarbon systems have a tube and fin condenser while the CO system uses a high-side gas 2

cooler. Flammable refrigerants, and hydrocarbons in particular, have not attracted a great deal of

support in the automobile industry because of the increased hazards from circulating the refrigerant

through the evaporator in the passenger compartment. Hydrocarbon refrigerants have been specifically

excluded from use for automobile air conditioning in the U.S. and are looked on with disfavor in other

countries. The proposed hydrocarbon system in this study includes the use of a secondary heat transfer

loop which would isolate the flammable refrigerant in the engine compartment so there would be no

significant increased risk to the passengers. This essential safety feature affects the system

thermodynamics, reducing its efficiency relative to traditional direct expansion systems, adds parasitic

power consumption to pump the heat transfer fluid, and increases the overall system weight. Advocates

of hydrocarbon air-conditioning systems have considered both propane (HC-290) and isobutane

(HC-600a) as refrigerants; there are very minor differences between these two refrigerants, as far as

this study is concerned, and only HC-600a is considered here.

Although the analysis includes much more detailed information than earlier studies, overall it

relies on the same approach of evaluating energy use for air conditioning, energy use for transporting the

system, and the direct effect from refrigerant emissions. Results are computed for thirteen different

countries 
0 10 20 30 40 50

-15

-10

-5

0

5

10

15

Outdoor Ambient Temperature (°C)

Evaporating Temperature (°C)

R-744

HFC-134a

HC-290

A

B

86

Figure 29. Evaporating temperatures.

7.2 ASSUMPTIONS

7.2.1 Cycle Efficiencies

Cycle efficiencies are calculated using refrigerant properties at specified operating conditions.

These include evaporating temperature, high-side heat transfer temperature (i.e. condensing

temperature, gas cooler exit temperature), liquid subcooling, and compressor suction temperature. The

transcritical cycle also includes high-side to low-side heat transfer.

Previous calculations of TEWI for automobile air conditioning were based on equivalent full

load hours of operation at a single design condition (Fischer, et al 1991, Fischer, et al 1994). The

conclusions based on this simplified approach may overstate the effects of energy use because the

efficiency at design conditions is lower than it is for most of the operating hours. A more detailed

calculation is performed in this analysis that accounts for changes in the evaporating and condensing

temperatures, and subcooling based on the vehicle speed and ambient temperatures.

7.2.1.1 Evaporating Temperatures

The outdoor ambient

temperature affects the refrigerant

temperature in the evaporator and the

evaporating temperatures are not the

same for HFC-134a, hydrocarbons,

and CO because of heat transfer 2

characteristics and additional )T’s

imposed by secondary loops.

Evaporating temperatures for

HFC-134a vary between -4EC and

7EC (25EF and 45EF) for the lowest

and highest ambient temperatures

(Patti 1996); additionally, typical

evaporating temperatures are -1EC, -

11EC, and 4EC (30EF, 12EF, and 40EF) for HFC-134a, hydrocarbons, and CO2, respectively

(Köhler 1996). These data are shown in Fig. 29 where the dotted line represents the evaporating

temperatures for an air conditioner using a hydrocarbon refrigerant (propane), the dashed line for

HFC-134a, and the two solid lines for CO . While computer simulations using the temperatures in Fig. 2

29 show similar evaporator air-side performance for the hydrocarbon and HFC-134a systems, this is

not the case for the CO2 air conditioner using the evaporator temperatures for prototype systems (line

A in Fig. 29). Calculated evaporator discharge air temperatures in that instance are 1.7E to 4.4EC (3E

to 8EF) higher than those computed for the HFC-134a and hydrocarbon systems. These differences in

supply air temperatures, and accompanying differences in latent capacity, may not seem significant in

mild climates, but they could have large impacts on passenger comfort in climates with high humidity
87

Refrigerant Idle Highway

HFC-134a 20EC (36EF) 7EC (13EF)

hydrocarbons 20EC (36EF) 7EC (13EF)

CO2 20EC (36EF) 7EC (13EF)

Table 1.. Automobile air-conditioning high-side approach

temperatures.

and/or large air conditioning loads. A second set of evaporating temperatures was defined for the CO2

air conditioner that would provide simulated air-side performance equivalent to the HFC-134a system

(based on evaporator effectivenesses of 85% and 90% for the HFC-134a and CO systems, 2

respectively). These evaporating temperatures for CO are shown as line B in Fig. 29. 2

7.2.1.2 Approach Temperatures

This report defines the approach temperature as the difference between the ambient air

temperature and the temperature of the refrigerant (vapor or liquid, saturated or subcooled) leaving the

condenser or the gas cooler. The approach temperature affects the high-side operating pressure for all

three air conditioning systems. It determines the condensing temperature, and hence the condensing

pressure, for the subcritical HFC-134a and hydrocarbon systems. The high-side pressure is chosen for

the transcritical CO system to give the maximum compressor COP for each gas cooler exit condition, 2

and hence for each desired ambient temperature. 

There are differences in the approach temperatures when vehicles are moving and when they

are idling in a stationary position. While moving, the approach temperature is essentially the )T

between the ambient air and the

refrigerant. When stationary, there is a

significant amount of engine heat

recirculated back through the radiator

and condenser as well as heat radiated

from the pavement. Approach

temperatures are also different

depending on the refrigerant because of

their individual thermophysical properties

and heat transfer characteristics. The

assumed approach temperatures are

listed in Table 15 (Pettersen 1997).

7.2.1.3 Subcooling And Return Gas Temperatures

The subcritical systems either have subcooled refrigerant or saturated liquid leaving the

condenser. This exit state depends on the operating conditions as well as the refrigerant charge; a

system designed to have subcooling with a full charge may not have subcooling at lower charges

experienced as refrigerant leaks from the system (Burk 1996). This analysis uses the simplified

assumption that HFC-134a and hydrocarbons operate with no subcooling under idle conditions and

8EC (15EF) subcooling at highway speeds. Subcooling is not a factor in transcritical systems. The

compressor suction gas temperature is 18EC (65EF) for the HFC-134a and hydrocarbon air

conditioners; it is 5EC (9EF) below the gas cooler exit temperature for the CO2 system because of the

assumed high/low-side heat exchanger. There is an assumed pressure drop for the CO2 system of 60
88

Refrigerant Idle Highway

HFC-134a 65% 60%

hydrocarbons 65% 60%

CO2 70% 65%

Table 2.. Compressor isentropic efficiencies.

kPa (8.7 psia) at idle conditions and 150 kPa (21.7 psia) under highway conditions for the accumulator

and high/low side heat exchanger.

7.2.2 Compressor Efficiency

Compression efficiencies vary with the shaft speed and between refrigerants. Only two shaft

speeds are considered in this analysis; a low

speed corresponding to idle and vehicle speeds

below 16 km/h (10 mph) and a high speed for all

other operating conditions. These two shaft

speeds have been arbitrarily chosen to be 900

and 2000 rpm, respectively. These values affect

the compressor efficiency, condenser subcooling,

approach temperatures, and volumetric

efficiencies. The assumed compressor

efficiencies are listed in Table 16 (Köhler 1997).

7.2.3 Clutch / Transmission / Belt Efficiency

Each of the air-conditioning systems considered requires some sort of drive mechanism to

operate the compressor from the engine drive shaft. Traditionally, a clutch and belt system is used for

subcritical HFC-134a air conditioners and a similar system could be employed for hydrocarbons. This

equipment is inappropriate to modulate the operation of a transcritical CO2 system because of the high

operating pressure and the relatively constant volumetric compressor efficiency. This analysis does not

consider alternatives to traditional drive systems that may be suitable for CO2; the calculations are

based on a drive efficiency of 95% for all three types of air conditioners.

7.2.4 Auxiliary Power Consumption

Electric parasitics degrade system performance from the compressor only efficiency or COP. It

is assumed that each of the systems considered uses an evaporator blower drawing 250 W and that the

blower has the same operating time as the compressor, which is not strictly true. It is also assumed that

a pump is used to circulate the secondary fluid for the hydrocarbon systems that draws 150 W,

inclusive of a 50% pump efficiency. The alternator efficiency is 50%.

7.2.5 COP At Idle And Highway Conditions

Cycle efficiencies are calculated using the assumptions listed above with refrigerant properties

calculated by commercially available computer software (F-Chart 1996, NIST 1996). Figures 30 and

31 show the system COPs at highway and low speed/idle conditions, respectively. Two curves are

shown for CO corresponding to the evaporating temperatures for prototype equipment and equivalent 2
15 20 25 30 35 40 45 50

0.0

0.5

1.0

1.5

2.0

2.5

3.0

Outdoor Ambient Temperature (°C)

System COP

HFC-134a

HC-290

R-744

A

B

15 20 25 30 35 40 45 50

0

1

2

3

4

5

Outdoor Ambient Temperature (°C)

System COP

HFC-134a

HC-290

R-744

A

B

89

Figure 30. Air-conditioner system efficiency at low engine

speeds.

Figure 31. Air-conditioner system efficiency at high engine

speeds.

air-side performance discussed in

Section 7.2.1.1. Requiring equivalent

evaporator discharge air temperatures

reduces the COPs for CO2 by 9% to

15% from those computed using the

evaporating temperatures from line A

in Fig. 29. As mentioned earlier, the

high-side pressure of the CO2 system

was selected in order to obtain the

highest compressor only COP within

limits of 7,600 to 15,000 kPa (1100

to 2175 psia). The lower pressure

limit is set in order for the refrigerant

to remain above the critical point, the

upper limit is from the design of a

prototype system (Pettersen 1994,

Pettersen 1997, Fernqvist 1997,

Denso 1997). 

7.2.6 Air-Conditioner Weight

Fuel is burned throughout the

year, discharging CO2 into the

atmosphere, simply as a result of

transporting the weight of the air

conditioner and associated

equipment. Heavier air conditioners

result in greater amounts of CO2

being discharged than lighter systems.

The HFC-134a air conditioner is assumed to weigh 11.4 kg (25 lb) (Taylor 1996); while the

hydrocarbon and CO air conditioners is assumed to weigh 15.4 kg (39 lb). The additional weight for 2

the hydrocarbon air conditioner is due to the refrigerant-to-brine and brine-to-air heat exchangers, the

intermediate brine itself, and the pump (Pettersen 1996). The additional weight for the CO system is 2

believed to be necessary in order to provide the cooling capacity and COPs used in the TEWI

calculations (Fernqvist 1997). The system weights used should exclude components of the heating and

ventilating systems which are not specific to the air-conditioning system (e.g. blower motor) and add

1.5 to 2 kg (3.3 to 4.4 lb) to the vehicle weight (Nonnenmann 1997)

The weight effect of CO emissions is directly proportional to annual vehicle usage (km or 2

miles). The average distances for 1996 are assumed to be 16,100 km/y for Europe, 10,300 km/y for

Japan, and 21,900 km/y for North America based on trends for the past 10 years in each region (Davis

1994).
90

7.2.7 Refrigerant Emissions And Direct Effect

The emission of HFC-134a from automobile air conditioners through leaks from hoses, fittings,

and shaft seals, servicing, and accidents can contribute to global warming because of the relatively high

GWPs of this gas; the corresponding direct effect from hydrocarbons or CO2 would be negligible

because of the very small GWPs of these gases. The magnitude of the direct effect from HFC-134a

depends directly on how well the charge is contained and how it is handled during servicing.

Historically, a single automobile air conditioner that used CFC-12 resulted in the release of

approximately five full charges of refrigerant during its operating lifetime:

! systems required servicing every three years, on an average, 

! any charge remaining in the air conditioner was vented to the atmosphere during servicing,

and 

! a full charge of refrigerant was used to diagnose the problem with the air conditioner before

repairing the system.

The redesign of air conditioners to use HFC-134a included improvements to hoses and fittings which

dramatically reduced refrigerant emissions and lengthened the service interval. It is estimated that

vehicle owners will bring their cars in for servicing after the air conditioner has lost 40% of its charge

and that systems will average 11⁄2 to 2 service visits during their operating lifetime (Baker 1997). 

Service practices, as mentioned, were also a major source of refrigerant emissions in the past

when refrigerants were inexpensive and their environmental effects were unknown. Economics and

attitudes toward the deliberate venting of refrigerants have changed significantly since the Montreal

Protocol was signed and most major developed countries have either adopted or are considering

regulations prohibiting the deliberate release of refrigerants to the atmosphere. These regulations are

also likely to mandate the recovery of refrigerant remaining in air-conditioning systems when they are

scrapped. Changing service practices should limit emissions of refrigerant to only what is required for

recharging (40% of the original equipment charge) and recovery should limit end-of-life emissions to no

more than 10% of the charge remaining in the system when it is scrapped.

Three different scenarios are considered for the direct effect of refrigerant emissions based on

the assumptions listed above:

! a minimum value calculated using leakage of 35 g/y for a new vehicle (Zietlow 1997, Hara

1996, Fernqvist 1997),

! emissions from 11⁄2 recharges during the operating lifetime of the air conditioner plus

recovery of 90% of the remaining charge when the vehicle is scrapped, and

! emissions from 2 recharges during the lifetime plus recovery of 90% of the remaining charge
91

when the vehicle is scrapped.

These scenarios are illustrated in Figs. 48, 49, and 50 in Appendix J; they correspond to lifetime

emissions of 415 g, 860 g, and 1080 g of HFC-134a, respectively, for an original equipment charge of

1000 g (Baker 1996, Nonnenmann 1996). No effort is directed in this analysis toward the technical

problems associated with containing hydrocarbon or CO refrigerants in systems using flexible hoses; 2

these refrigerants have smaller molecules than HFC-134a and may operate at significantly higher

pressures (as much as 5 times as high for CO2) and leakage rates could be quite higher than they are

projected to be for HFC-134a. Higher emissions of these low GWP refrigerants would have a strong

effect on system performance, maintenance requirements, and customer satisfaction, but not on TEWI.

7.3 METHODOLOGY

7.3.1 Ambient Temperature Distributions

Ambient temperature profiles are used for locations; twelve in Europe (two in the U.K., four in

Germany, two in Greece, two in Italy, and two in Spain), four in Japan, and four in North America

(USAF 1978). Data for Europe are averaged to obtain five national values which are used in the TEWI

calculations. Automobiles are manufactured for broad markets, and consequently the air-conditioning

systems are designed according to the requirements necessary to satisfy customers in the hottest, most

humid climates where the cars will be sold. TEWI are presented only for the most demanding

temperature profiles for Europe, Japan, and North America.

Figures 51 to 53 in Appendix K show the number of hours per year that the outdoor ambient

temperature is in a 5.6EC (10EF) temperature bin for a typical meteorological year. The temperate

climate of Greece in Fig. 51 shows a distribution shifted toward lower temperatures while the profiles

for Okinawa, Japan and the Southwestern U.S. show more hours at higher temperatures. Figure 53

shows a significant number of hours at the design condition of 35EC (95EF).

7.3.2 Compressor On-Times

The outdoor temperature is related to compressor on-times using a curve derived from

manufacturer’s fleet data in Phoenix, Arizona USA (Fernqvist 1996). The compressor percent on-time
0 10 20 30 40 50

0

20

40

60

80

100

Outdoor Ambient Temperature (°C)

On-Time (%)

92

Figure 32. Compressor operation and average daily

temperature.

is correlated against the daily average

temperature, as shown in Fig. 32.

7.3.3 Idle and Highway Operation

The final results of the TEWI

calculation incorporate a weighted

average of energy consumption for

idle and low vehicle speeds (e.g. 900

rpm) and highway speeds (e.g. 2000

rpm). It is assumed that driving

patterns for Europe and North

America are similar and that 15% of

vehicle use is at low engine speeds

(and low compressor shaft speeds)

and 85% at high engine speeds (Siewert 1983). Vehicle usage in Japan is significantly different than in

Europe or North America. Approximately 47% of vehicle operation is at idle conditions and 53% at

highway speeds (Denso 2 1996).

7.3.4 Cooling Capacity

Compressor power at each ambient temperature is computed using the COPs shown in Figs.

30 and 31, the compressor on-time as a function of ambient temperature shown in Fig. 32, and

constant cooling capacities of 7 kW (24,000 Btu/h) and 3.27 kW (11,000 Btu/h) at highway and idle

conditions, respectively. The compressor on-time determines the cooling load as a function of the

ambient temperature, in these calculations, and the analysis does not differentiate between the sensible

and the latent loads. A more rigorous evaluation is required to account for the effects of the latent load

(Bhatti 1997).

7.3.5 TEWI Calculations

7.3.5.1 Power Consumption For Air Conditioning

System power consumption is calculated at each ambient temperature using the temperature

distributions (e.g. Figs. 51 to 53), annual driving distance, and a correspondence between driving

distance and hours of vehicle operation. Davis (1994) reported data for average annual vehicle driving

distances for the U.S., Japan, and many European countries. This information is used in conjunction

with data on hours of vehicle use to determine air-conditioning power consumption (Zietlow 1997,

Hara 1996, Siewert 1983). It is assumed that the annual hours of operation have the same distribution

as the ambient temperature in order to estimate the vehicle hours at each ambient temperature. Annual

cooling output at each ambient is calculated at low and high speeds using the number of driving hours,
93

compressor percent on-time, and cooling capacities at low and high speeds. Annual power

consumption is then the sum of the cooling output at each ambient divided by the corresponding system

COP.

7.3.5.2 Incremental engine efficiency

Overall the thermal efficiency of gasoline and diesel engines is between 25% and 30% for

operating the vehicle under full load conditions (Fischer, et al 1991, Nonnenmann 1996, Denso 1996).

Although used in the past (Fischer, et al 1991, Fischer, et al 1994) this range of values is considered

inappropriate for accounting for the incremental energy use resulting from operating the air conditioner;

values of 40% (Nonnenmann 1996) and 50% (Wertenbach 1996) are cited. TEWI calculations in this

analysis use an incremental engine efficiency of 40%.

7.3.5.3 CO2 Emissions From Air-Conditioning Energy Use

The total energy requirement for air-conditioner operation is divided by the incremental engine

efficiency to obtain the energy input to the vehicle engine for air conditioning. The CO emissions 2

associated with this energy input (kWh) are determined based on 0.243 kg CO per kWh of input 2

(Fischer, et al 1994). The annual value is multiplied by the assumed air-conditioner lifetime to obtain a

lifetime value for CO emissions from air conditioning. 2

7.3.5.4 CO2 Emissions From Transporting The Air Conditioner

The fuel use necessary for transporting the air conditioner is calculated for each of the three

systems by multiplying the assumed weight by 57x10 liters/kg/km times the regional annual vehicle use -6

(km) (Fischer, et al 1991). The fuel use is converted to CO emissions based on 2.32 kg CO2 / liter of 2

gasoline (Fischer, et al 1994). This value is also multiplied by the lifetime to get total CO emissions for 2

system weight.

7.3.5.5 Direct Effect Of Refrigerant Emissions

Direct effects of refrigerant emissions are calculated based on the estimated lifetime emission of

the refrigerant and the refrigerant GWP. The direct effects of the hydrocarbon and CO2 systems are

considered negligible and no effort is given to refining estimates of emission rates.

7.4 RESULTS

TEWI are calculated for the HFC-134a air conditioner using the three refrigerant emission scenarios

identified earlier and for the hydrocarbon and CO2 systems with emissions relative to two system
HFC-134

HC-290

R-744 (A)

R-744 (B)

HFC-134a

HC-290

R-744 (A)

R-744 (B)

HFC-134a

HC-290

R-744 (A)

R-744 (B)

0 1,000 2,000 3,000 4,000 5,000 6,000

TEWI (kg CO )2

Indirect

(A/C energy)

Indirect

(weight)

Direct

(minimum)

Direct

(1.5 recharges)

(2 recharges)

Direct

North America Europe Japan

94

Region

Refrigerant

HFC-134a Hydrocarbon

CO2

A B

Europe

A/C Energy

Weight

Direct Effect

TEWI

1255

245

540 to 1404

2040 to 2904

1838

331

 4

2173

1547

331

 1

1878

2048

331

 1

2380

Japan

A/C Energy

Weight

Direct Effect

TEWI

972

154

507 to 983

1633 to 2109

1408

167

 3

1578

1147

167

 0

1314

1694

167

 0

1861

North America

A/C Energy

Weight

Direct Effect

TEWI

2889

363

540 to 1404

3792 to 4656

4160

491

 4

4655

3759

491

 1

4240

4902

491

 1

5393

Table 17. TEWI for HFC-134a, hydrocarbon, and

transcritical CO air conditioners. 2

Figure 33. TEWI for HFC-134a, HC-290, and CO2 (R-744) air conditioners in

selected design climates.

recharges for HFC-134a. These

results are summarized in Table 17 for

three climates with high airconditioning demand (i.e. the

southwestern U.S., Greece, and

Okinawa, Japan). Results for CO2 are

shown for calculations based on

prototype equipment (column A) and

equivalent discharge air temperatures

(column B). All four sets of results

differ substantially between the regions

(shown) as they do within each region

(not shown). This information is also

summarized in Fig. 33; detailed results

for all thirteen climates are in

Appendix L. 

The bar chart shows results for all

three emissions scenarios, although the direct effects are so small for HC-600a and R-744 that they are

not visible. Groups of bars are shown for Japan, Europe, and North America; each group contains bars

for the TEWI of air conditioners using HFC-134a, HC-600a (isobutane), and R-744 (CO2). The
95

heavily shaded segment of each bar corresponds to the CO2 emissions from fuel use necessary to drive

the compressor. This segment is shortest for HFC-134a for each region. The second segment of each

bar corresponds to fuel use to transport the weight of the air conditioner. The three segments that

appear on the right of the bars for HFC-134a correspond to the three emissions scenarios; a minimum

value based on estimates of leakage for systems straight from the factory, and increment corresponding

to how much more refrigerant would be lost assuming 11⁄2 recharges, and a second increment assuming

2 recharges during the lifetime of the air conditioner. In other words, the bars for HFC-134a show the

TEWI assuming 2 system recharges during the lifetime; reducing emissions to 11⁄2 recharges eliminates

the rightmost segment; the lowest possible TEWI for HFC-134a is shown by eliminating the two

segments on the right (one air conditioner manufacturer believes they can cut emissions lower than the

“as manufactured” rates assumed in this analysis). It should be noted that different interpretation of the

assumptions used in this analysis and the incorporation of latent cooling loads can alter the results from

those presented here (Bhatti 1997).

7.5 CONCLUSIONS

The information in Table 17 and the detailed results in the appendices allow some general

conclusions:

! under most of the scenarios considered the TEWI of CO2 air conditioners is lower than

hydrocarbon or HFC-134a air conditioners; it is comparable to to higher than that of

HFC-134a in the regions of North America with high air-conditioning loads depending on

assumptions for the emission rates of HFC-134a and the evaporating temperatures (and

corresponding) discharge air temperatures for CO .2

! fuel consumption is lower for air conditioners using HFC-134a than for either of the two

alternative refrigerants under all of the climate and driving conditions considered. The

transcritical CO2 system requires 14 to 66% more energy, and fuel, for power and

transport the air conditioner than HFC-134a and hydrocarbons 35 to 45% more than

HFC-134a. It should be noted that fuel use for air conditioning has not been a major issue

in the past, and the assumptions used here are for an HFC-134a system that has not been

designed to provide maximum efficiency.

! estimated TEWI of the hydrocarbon system is generally lower than that of the HFC-134a

systems in cool climate regions; it is comparable to or higher than that of HFC-134a in

areas with high air conditioning loads.
96
 AFUE - Annual Fuel Utilization Efficiency - appliance heating efficiency calculated by 1

assuming 100% of the fuel is converted to thermal energy and then subtracting losses for exhausted

sensible and latent heat, cyclic effects, infiltration, and pilot losses over the whole year. AFUE does not

include electrical energy used for fans, pumps, ignition, exhaust, or blowers.

97

8. RESIDENTIAL GAS HEATING/COOLING OPTIONS

8.1 INTRODUCTION

In regions where natural gas is available as an energy source, additional technology options and

combinations are possible for unitary heating and air conditioning. Natural gas is used in commercial

and residential applications because of its convenience and affordability. The U.S. Department of

Energy (DOE) ranks natural gas as having the lowest average energy cost per therm when compared

with refined petroleum products, liquefied propane (LPG), and electricity. Natural gas forced air or

hot water furnaces in combination with an electrically powered air conditioner is a popular option for

residential comfort conditioning. Unitary air conditioning units and heat pumps that use gas as the

primary source of energy are currently available and under development.

8.2 UNITARY GAS TECHNOLOGIES

Gas furnaces in combination with a centralized, vapor-compression air conditioner is evaluated

to provide a baseline for TEWI comparisons. In appropriate locations an electric resistance heat

option is assessed.

8.2.1 Gas Engine Driven Heat Pumps

Engine-driven heat pumps rely on the same vapor compression cycle for heating and cooling as

conventional electric-powered systems except the electric motor is replaced by an internal combustion

engine which is usually powered by natural gas. Space heating capacity is supplemented by engine and

exhaust heat recovery. Local market conditions and factors such as the relative costs of gas and

electricity and local utility incentives rather than total TEWI are the main criteria used to choose one fuel

over the other. The seasonal heating performance of a gas engine heat pump available since 1994 is

listed at 126% AFUE by the Natural Gas Cooling Equipment Guide (AGCC 1996). TEWI values 1

were computed for this system using its published efficiencies.
98

8.2.2 Gas-Fired Absorption Heat Pumps

Absorption heat pumps using ammonia-water are under development to provide heating and

cooling for residential and light commercial applications. One product, based on the generator absorber

heat exchange (GAX) cycle, has reached the proof-of-concept stage under a project sponsored by the

U.S. Department of Energy and major American manufacturers have seriously considered hardware

introduction (AGCC 1996). A small number of ammonia-water GAX heat pump prototypes have been

built in the U.S. for testing and, optimistically, a product could enter the market in 1998; a more realistic

date for market entry is 2000 (Fiskum 1996, Erickson 1996). 

The GAX heating efficiency exceeds that of a gas furnace by 20% to 80% depending on the

furnace efficiency used for the comparison. Conservative target steady-state efficiencies are 1.4 COP

heating (gas-fired) at 8.3EC (47EF) and 0.9 COP cooling (gas-fired) at the DOE/B (82EF) test

condition. Parasitic power consumption targets for the outdoor package are in the range of 150 to 190

Watts per ton of refrigeration capacity (42 to 54 W/kW) (Marsala 1993). Similar efficiency data are

cited by Erickson and Rane (1992) with cycle modifications described that could boost the cooling

COP to 1.5. The next generation of GAX equipment will push the limits of single-stage efficiency and

will require the development of cost-effective heat and mass transfer surfaces for the absorber and

generator which are required to maintain the desired temperature and concentration profiles needed to

achieve the full performance potential from the sealed absorber system (Marsala 1993).

Actual performance tests performed on the prototype GAX heat pumps have measured gas

COPs of 1.4 heating (8.3EC, 47EF) and 0.75 cooling (28EC, 82EF), including flue losses but excluding

electric parasitics (DeVault 1994).

There are strong prospects for successful commercialization of the GAX concept for heat pump

applications if the remaining technical obstacles are overcome. Masala (1993) enumerated the technical

issues that must be resolved in bringing the GAX heat pump to market, these include:

! developing a solution pump that has a long life, low cost, and can operate without a lubricant in

a harsh environment,

! eliminating the production of non-condensable gases in the refrigerant/absorbent loop,

! reducing parasitic electric power consumption, and

! developing cost effective manufacturing techniques and quality control.

TEWIs are calculated for a GAX gas absorption heat pump in this section based on the same

projected heating and cooling COPs used in the previous AFEAS/DOE study which included parasitic

loads (Fischer 1994).

8.3.3 Desiccants
99

Desiccant materials which absorb or adsorb water vapor from the air can be used effectively in

building air-conditioning systems. Sprayed-liquid desiccants or solid desiccants that are bonded to a

porous substrate are used to dehumidify building ventilation air rather than cooling it below its dew point

which is the approach used by conventional air-conditioning equipment. This process produces dry, but

hot air since the water vapor gives up its latent heat of vaporization in the process. At this point of a

desiccant cycle, no useful work has been accomplished because the enthalpy of the “wet” and that of

the “dry” air are essentially the same (Collier 1996). So, some means for sensibly cooling the air after it

has been dehumidified is needed before it can be used in a conditioned space. The energy efficiency of

desiccant cooling applications is facilitated when this cooling can be accomplished via low energy input

processes like direct or indirect evaporative coolers or an air-to-air heat exchanger. Use of waste heat

to regenerate the desiccant also improves net efficiency.

Various hybrid cycles and combinations of HVAC equipment are currently used with

desiccants to provide low humidity air in hospital operating rooms, supermarkets, and hotels.

Dehumidification with desiccants allows building owners to install lower capacity air conditioning

systems (because the desiccant is handling the latent building load), exchange gas energy or waste heat

for electrical energy, and provide ventilation air with better indoor air quality. Because desiccant

systems tend to be designed on a custom basis and since they are usually used in combination with a

large variety of conventional air conditioning equipment, they were not evaluated in this report.

8.3 ASSUMPTIONS

National annual averages are used for the power plant emission rates in these calculations. The

averaged electrical power plant emission rates are 0.650 kg CO2/kWh for North America and 0.470

kg CO2/kWh for Europe (see Appendix A). These emission rates are compiled from the open literature

rather than calculated from the fundamental heat contents of fuels, fuel mix used in power production,

plant efficiencies, and transportation and distribution losses. Total CO2 production is divided by

delivered kWh so all of the power plant conversion and distribution losses are accounted for.

The heat content and carbon dioxide emission rate for natural gas which were used for the gas

powered technologies were 38,200 kJ/m and 51.1 g/MJ, respectively. A 96.5% distribution efficiency 3

was assumed for natural gas which raised the CO emission rate to 53.0 g/MJ (55.9 g CO /1000 Btu) 2 2

at its point of use (EIA 1997).

Fifteen year lifetimes are assumed for U.S. and European unitary equipment. Based on

information assembled from ARI member companies air conditioner annual leak rates of 4% of the

charge for 1996 equipment were used for the TEWI calculations. An end-of-life (E.O.L.) Charge loss

rate of 15% was rationalized for residential air conditioning units on the basis of recovering 90% of the

charge from 95% of the field units, but allowing for a 100% charge loss from about 5% of field stock

(Hourahan 1996a). The 1996 annual leak or make up rate and 15% E.O.L. loss were also used for the

European equipment calculations.
100

System

Efficiencies

Cooling/Heating

Electric A/C and Gas Furnace

 minimum efficiency

 high efficiency

SEER-10 / 80% furnace

SEER-12 / 92% furnace

Premium Technologies

 electric A/C and gas furnace

 engine driven heat pump (HCFC-22)

SEER-14 / 92% furnace

gCOP-1.30 / gCOP-1.28

2005 Technologies

 GAX absorption heat pump gCOP-1.0 / gCOP-1.5

Table 18. Residential gas options.

8.4 METHODOLOGY

8.4.1 Central Heating and Air Conditioning

Total equivalent warming impacts were calculated for baseline 10.5 kW (36,000 Btu/h) central

air conditioners with SEERs of 10 and 12 in combination with an 80% or 92% efficient gas furnace.

Seasonal energy use is computed based on a “typical” 1800 square foot residence with a 74.7x106

Btu/h (78.8x10 kJ/y) heating load and 16.1x10 Btu/y (17.0x10 kJ/y) cooling load in Pittsburgh, 6 6 6

Pennsylvania, USA; a 34.8x10 Btu/h (36.7x106 kJ/y) heating load and 33.8x10 Btu/y (35.7x10 6 6 6

kJ/y) cooling load in Atlanta, Georgia, USA; and 0 Btu/y (0 kJ/y) heating load and 82.2x10 Btu/y 6

(86.7x10 kJ/h) cooling load in Miami, Florida, USA (Ballou 1981). 6

For premium technology residential equipment, the baseline SEER of a central air conditioning

unit was increased to 14. The gas powered GAX absorption heat pump that is being developed by

DOE is shown as a 2005 option for Pittsburgh, Atlanta, and Europe.

Residential heating-only TEWI calculations were performed for Europe using averaged loads

and seasonal performance factors (SPFs), Appendix D. An average heating load for all European

countries was used. The European average of 0.470 kg CO2/kWh was used to convert electric use to

CO emissions. 2

8.5 RESULTS AND DISCUSSION

Total equivalent warming impacts for various residential heating/cooling options incorporating

natural gas were calculated for Pittsburgh, Pennsylvania; Atlanta, Georgia; and Miami, Florida in the

United States, and the results are shown in Figs. 33 to 35. These results are computed using the

efficiency data in Table 18. Each figure has two sections, the upper portion shows “standard

technologies,” or heating/cooling options that represent baseline costs for a residential system in each of

these cities, while the lower

indicates a “premium

technologies” section which is a

heating/cooling option that is

significantly more expensive

than the baseline technology as

in Chapter 4 of this report.

Each segment of the

bar graphs plotted in these

figures indicates TEWI

contributions from different

sources. The initial, darker gray

section of most of the bar

graphs is the indirect
Central A/C & Furnace

Minimum Efficiency

High Efficiency

Premium Technologies

 Central A/C & Gas

 Engine Driven HP

2005 Technologies

 GAX Absorption HP

0 25,000 50,000 75,000 100,000 125,000 150,000 175,000

TEWI (kg CO )

Indirect Effects

electricity

gas

secondary pump or

refrigerant emissions

Direct Effects

auxiliary electric

2

Central A/C & Gas Furnace

 minimum efficiency

 high efficiency

Premium Technologies

 central a/c & gas furnace

 engine driven HP

2005 Technologies

 GAX absorption HP

0 25,000 50,000 75,000 100,000 125,000

Indirect Effects

electricity

gas

secondary pump or

refrigerant emissions

Direct Effects

auxiliary electric

TEWI (kg CO ) 2

101

Figure 34. TEWI for residential gas heating/cooling options in

Pittsburgh, Pennsylvania, USA.

Figure 35. TEWI for residential gas heating/cooling options in

Atlanta, Georgia USA.

contribution from electric power used for the vapor compression cooling process. The lighter gray

section of bars shown in Figs. 34 to 36 with a lower density of fill dots, is the indirect contribution from

gas combustion. The lightest section on some bars show the indirect contribution to TEWI from

auxiliary electric parasitics such as pumps or fans that are not included in the SEER or HSPF ratings of

the equipment. The moderate gray section on the ends of most of the bar graphs is the direct

contribution to TEWI caused by refrigerant losses.

Figures 34 through 36 show

the advantages of increased

efficiency in both vapor-compression

air conditioning and natural gas

powered options.

In Pittsburgh and Atlanta

which have substantial heating loads,

newer gas-driven technologies like

the engine-driven heat pump and the

GAX absorption heat pump show a

significant decrease in TEWI when

compared to options using a

conventional gas furnace. The TEWI

advantage is also seen in the heating

only data for residential European

applications in Fig. 37.

In all the cases presented,

the direct contribution of refrigerant

to the TEWI was no larger than 6%

of the total. Results presented for

alternative refrigerants R-407C and

R-410A in Chapter 4 of this report

indicate they show efficiencies and

GWPs similar to refrigerant

HCFC-22 in the electric air

conditioning unit and, therefore, a

similar total TEWI.

As with heat pumps, nearly

80% of the direct TEWI effect for

vapor compression air conditioning

systems is due to the assumption on

annual emissions from leakage,

accidents, and maintenance practices. Procedures requiring conscientious maintenance and repairs of

leaks and strict adherence to refrigerant recovery are adopted and followed, the direct effect will

diminish in significance.
Air-to-Air A/C

HCFC-22: minimum efficiency

HCFC-22: high efficiency

Premium Technologies

high efficiency a/c

 engine driven a/c

0 25,000 50,000 75,000 100,000 125,000

Indirect Effects

electricity

gas

secondary pump or

refrigerant emissions

Direct Effects

auxiliary electric

TEWI (kg CO ) 2

102

Figure 36. TEWI for residential cooling options in Miami,

Florida USA.

Figure 37. TEWI for residential heating only options in Europe.

8.6 CONCLUSIONS

In climates with appreciable

heating loads, the gas-fired heat pump

options have a significantly smaller

TEWI than combined gas and electric

air conditioning options. This

advantage decreases as the balance

shifts to lower heating and higher

cooling loads. The comparisons

presented in Figs. 34 to 36 and Fig.

37 are based on laboratory

measurements for a GAX absorption

heat pump and ratings data for a

commercially available engine-driven

heat pump. First cost, climate, and

projected operating costs rather than

TEWI are likely to be the main

criteria for selection of a residential

heating/cooling system.
103

9. CONCLUSIONS

The TEWI of the technology options considered in this report have been estimated for several

end-use applications using 100 year ITH global warming potentials for each of the relevant greenhouse

gases. This choice of time horizon primarily affects alternative systems based on fluorocarbon

refrigerants and blowing agents and to a lesser extent those alternatives based on hydrocarbons. The

direct emissions effect of refrigerant and blowing agents is more heavily emphasized by using the 100

year GWP values in calculating the TEWI (the 100 year GWP values are approximately three times the

500 year values used in the 1991 TEWI report). Ultimately however, the choice of time horizon is a

political issue beyond the scope of this report.

Scientifically, there are arguments for using infinite time horizons or approximating them by using

a 500 year ITH as was done in the 1991 AFEAS/DOE study. Scientists have also stated that while

long term change reflects the cumulative effects, there is also a real concern that high release rates of

greenhouse gases could affect the rate of climate change in the next several decades. For these

reasons, policy makers typically use the 100 year ITH. Therefore this study, like the 1994 project,

presents the principal comparison results on the 100 year time horizon basis.

As in the previous studies, a consistent conclusion drawn from using either a 100 year or 500

year time horizon is that improving energy efficiency is a powerful tool to mitigate future potential

climate change since it is directly connected to energy-related CO emissions. Furthermore, emissions 2

of fluorocarbons need to be, and will be, minimized wherever practical to ensure that their direct

contribution to global warming does not outweigh the efficiency benefit derived from the use of

fluorocarbons as refrigerants and blowing agents when considering at TEWI for most applications.

9.1 GENERAL CONCLUSIONS

Several broad conclusions can be drawn from the study.

! TEWI evaluations emphasize the combined environmental effect of the direct emission of

greenhouse gases with the indirect effects of CO2 emissions from energy use by equipment using

these fluids as refrigerants or blowing agents. This is only one criterion in selecting between

technology options. System costs, operating costs, regional energy costs, ease of maintenance,

continuing technology improvements, etc., are equally important factors to consider in selecting the

most appropriate technology for any specific application. 

! Reductions in TEWI through the use of ammonia or hydrocarbons as refrigerants are insignificant

for refrigeration systems with low emissions and may lead to an increase in energy use when

applications of these fluids must meet the same safety design criteria currently defined as acceptable
104

for fluorocarbon refrigerants.

- Ammonia and some hydrocarbon refrigerants have thermophysical properties comparable to

(and for some applications superior to) those of HCFC or HFC refrigerants. They also have

system irreversibilities and system design features necessary for safe products (e.g., secondary

loops) which reduce their overall efficiency. Such changes often offset much of the TEWI

benefit claimed for non-fluorocarbon refrigerants.

- Insignificant TEWI differences for most applications occur when design and service

requirements, for low refrigerant emissions and safe operation of equipment using flammable or

toxic refrigerants, are applied to systems engineered for non-flammable or non-toxic

refrigerants.

! The direct contribution of HFC refrigerant and blowing agent emissions in refrigeration and

insulation applications is given greater emphasis by using shorter-term GWP values in calculating the

TEWI (the 100 year GWP values used herein are approximately three times the 500 year values

used in the 1991 TEWI report).

! Actions to reduce fluorocarbon system refrigerant losses will result in lower TEWIs for

supermarket refrigeration and automobile air-conditioning systems using these fluids.

! Non-fluorocarbon technologies may penetrate into mainstream refrigeration and air conditioning

application areas, but it is unlikely that they will significantly displace conventional fluorocarbon

technologies in the near future. Changing to HFC and non-fluorocarbon technologies both demand

some technician and servicing personnel retraining, but procedures associated with HFC

applications are more consistent with current practices associated with HCFC refrigerants and

blowing agents.

! Efficiencies of conventional technologies are likely to increase as electric and gas-driven equipment

and insulating foam formulations are further optimized for replacement refrigerants and blowing

agents.

! Innovative design and modifications of standard practice can lead to significant reductions in TEWI

for refrigeration systems using ammonia, fluorocarbon, or hydrocarbon refrigerants. These include

mandatory refrigerant recovery and recycling, distributed refrigeration systems, charge reduction,

elimination of flared fittings and reduced numbers of brazed connections, highly efficient purge units,

improved heat transfer surfaces, high-efficiency compressors, etc. 

! Average annual CO2 emissions from electricity generation vary widely for individual regions and

countries - from 0.0 to over 1.0 kg CO /kWh compared to the 1993 World average of 0.58. 2

Emission rates also vary with season and time of day depending on how the generation fuel mix

changes. Overall TEWI values in any particular location will be peculiar to the local electrical power
105

generating efficiency and seasonal and time of day generating characteristics. The direct contribution

can range from all (or nearly all) of total TEWI in areas with low CO2 emission rates [using mostly

nuclear or hydro power] to a minor fraction of TEWI for areas with high rates [using mostly coal].

9.2 Individual Applications

9.2.1 Household Refrigeration

The phase-out of HCFCs affects TEWI for household refrigerator/freezers through the choice of

refrigerant and insulating material. No clear conclusions are possible at this time for foam blowing

agents to replace HCFC-141b; the use of hydrocarbon and HFC blown foam insulations are under

active development in an effort to improve their thermal and mechanical properties. Hydrocarbon

blown foams continue to have higher thermal conductivity than HCFC-141b and HFC foams and,

consequently, exhibit higher energy use with increased impact on CO2 emissions. The increased energy

use must be balanced against any direct impact caused by the HFC blowing agent itself.

TEWI estimates from this analysis for household refrigerator-freezers using HC refrigerants and

foams are about 4-5% lower (in North America and Japan) and about 13% lower (in Europe) than

those of HFC-134a refrigerant/HFC foam units assuming refrigerant recovery at end-of-life disposal.

Energy use estimates for HC-based refrigerators are about 10% greater than that of the HFC units in all

regions. Use of HCs in refrigerators raises safety concerns and has resulted in higher unit costs.

Applying this cost differential to HFC designs (to incorporate vacuum panel insulation in the cabinet

walls, for instance) could yield a product with potentially superior TEWI characteristics.

9.2.2 Automobile Air Conditioners

The direct effect of refrigerant emissions for HFC based automobile air conditioners is a significant

part of the TEWI. The automobile manufacturers have responded aggressively with efforts to reduce

charge size and emissions. Research and laboratory development of air-conditioning systems based on

transcritical CO compression show a potential to reduce TEWI for this application. Estimated TEWIs 2

for CO and hydrocarbon systems are lower than those for HFC-134a systems in regions with cool 2

climates; TEWI are comparable to higher in climates with high cooling loads. Energy consumption

estimates for HFC systems are consistently lower than those of CO2 systems. The long term

performance, lifetimes, viability, and TEWIs of both the alternative systems must be proven through

extensive prototype and field trial testing. Energy consumption and TEWIs for HC-based systems are

negatively affected in all regions due to use of an indirect loop with attendant efficiency penalties to

keep the flammable refrigerant out of the passenger compartment.

9.2.3 Chillers

TEWI for this class of equipment has fallen significantly since the early 90's. New electric chillers
106

have 25% to 30% lower TEWIs than models of 4-5 years ago due to replacement of CFC refrigerants

with HCFC and HFC alternatives and to significant improvements in energy efficiency and reductions in

refrigerant loss rates. The choice of refrigerant makes only a minor difference in direct TEWI in new

equipment. Differences in chiller efficiencies for various refrigerant options can have a significant impact

on the indirect contribution to TEWI, however, which is dominant in this application.

Significant advances have also been made in gas-fired chiller technologies. Triple-effect absorption

chillers now under development show potential for 25-30% reductions in TEWI compared to existing

double-effect machines. Engine-driven chillers are now available with rated efficiencies more than 25%

higher than the value used in the TEWI-II report. Estimated TEWIs for these machines are around 25%

lower than those for the triple-effect absorption equipment.

TEWI estimates given herein for gas-fired engine driven and absorption chillers are not directly

comparable to those for electric driven chillers. It is difficult to make accurate interfuel comparisons

using TEWI as computed in this report. At the very least “local” CO2 emission factors need to be

used for electric power generation instead of the broad regional averages used in this study, and some

technologies may in fact require time of year or time of day factors to account for differences in

emissions due to intermediate and peak power. These differences are in part due to the types of

generating equipment brought on line for peak demand, changes in the local fuel mix, and lower

transmission and distribution efficiencies during peak generating periods.

9.2.4 Unitary Equipment

Transition away from HCFC-22 in this equipment appears to be achievable with either no change

or a slight reduction in the estimated TEWIs. The HFC-400 blends R-407C and R-410A are the

principal alternatives being considered as HCFC-22 substitutes. Laboratory and limited field testing

indicates that R-407C has equivalent performance compared to R-22 while R-410A-based equipment

has potential for slightly better efficiency, lowering the indirect contribution to TEWI. Geothermal heat

pumps and premium grade air-to-air heat pumps can significantly reduce TEWI for this application,

albeit with higher purchase costs.

Gas engine-driven and gas-fired absorption heat pumps for space heating and cooling show

potential to reduce TEWI in climates dominated by heating requirements. Long term performance and

reliability of the gas-driven technologies have not been demonstrated.

9.2.5 Commercial Refrigeration

Supermarket refrigeration systems have had high direct contributions because of historically high

refrigerant charges and leakage rates. Equipment manufacturers have worked to reduce refrigerant

leaks at the display cases and in the brazed and welded joints in refrigerant lines. The lower emission

rates have reduced TEWI significantly from the values reported previously. New system design

concepts (secondary loop and distributed compressor approaches) also dramatically reduce the direct

effect of refrigerant emissions and result in lower overall TEWI estimates for this application. The

differences in TEWI between the HFC mixtures that have been considered are due primarily to the
107

GWPs of the refrigerants; the differences in energy use are not considered significant. Ammonia with

secondary heat transfer loops has been shown to be a viable alternative for HFCs in this application,

but there can be an energy penalty associated with necessary secondary heat exchangers. In many

areas system designs will have to comply with regulation and permit requirements intended to ensure

safe use in retail and commercial areas. Some European and developing countries have fewer

regulations and are more open to using ammonia. Refrigerant containment measures necessary for

ammonia and hydrocarbons could also be used with HFCs, resulting in essentially identical TEWI for

these alternatives.
108
109

10. REFERENCES

ACH & RN., 1996, “New McQuay Chiller to Use AZ-20," Air Conditioning, Heating, and

Refrigeration News October 28, p. 5.

AGCC, 1996. Natural Gas Cooling Equipment Guide, American Gas Cooling Center, Fourth

Edition, Arlington, VA, April.

AHAM. 1996. 1996 AHAM Directory of Certified Refrigerators and Freezers, Edition No. 1,

Chicago, IL, January.

ANSI #Z21.40.4. 1994. “Performance Testing and Rating of Gas-Fired, Air-Conditioning and

Heat Pumping Appliances,” American Gas Association Laboratories, 8501 East Pleasant Valley

Road, Cleveland, Ohio, U.S.A.

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121

APPENDIX A: ELECTRIC POWER CO2 EMISSION RATE DATA
122
123

A.1 REGIONAL AVERAGES AND RANGES

In the first report (Fischer, et al 1991) CO emission rates were calculated based on data for 2

the carbon content of fossil fuels used for electricity generation (coal, oil, and natural gas) and the

average mix of generation capacity for a given region (Japan, Europe, or North America). For the

present study it was decided to use publicly available data on CO2 emissions from each major region

wherever possible. Accordingly much effort was devoted to soliciting data from utilities, utility

organizations, and other sources. Considerable data were received, primarily from European sources,

and are summarized in Tables 19 and 20. The data sources for these tables are listed as follows:

1. Jean-Yves Caneill, Electricite3 de France, 1996; data give “order of magnitude” CO emissions 2

for current plants of 0.99 kg CO /kWh for coal, 0.87 for oil, and 0.67 for gas; also gives 2

projections for future power plants (2010) of 0.73 for coal, 0.76 for oil, and 0.35 for gas.

Caneill also included a paper by W. Ruijgrok (KEMA), presented at a 1993 UNIPEDE/IEA

conference in Hamburg. This paper gives 1989 data on average emissions for EU-12 and

America of 0.39 and 0.54 kg CO /kWh, respectively. 2

2. S. Fischer, et al. 1994. Energy and Global Warming Impacts of Not-In-Kind and Next

Generation CFC and HCFC Alternatives, AFEAS and DOE.

3. Bert Stuij, IEA/Paris, 1996; gives average emissions for European OECD countries of 0.50

(range of 0.00 to 1.23 kg CO /kWh), a value of 0.71 for the U.S., and an OECD average of 2

0.58, all based on 1990 data.

4. P. Göricke, RWE Energie, Germany, 1996; provided a paper “Fuel Cycle Analysis” by H. J.

Laue giving 1990 average value for EU-12 of 0.50 (range of 0.08 to 0.96); also provided a

1995 value for Germany of 0.61 from VDEW (German association of electric power

generation and distribution companies).

5. K. Hara, Japan Industrial Conference for Ozone Layer Protection (JICOP), 1996; provided an

average value of 0.473 kg CO2/kWh for 1995 based on consolidated input from Japanese

electricity producers.

6. Jurgen Michorius, Dutch Electricity Generating Board, 1996; gives 1992 average emission data

for EU-12 of 0.47 kg CO2/kWh with country averages ranging from 0.00 to 1.08; major data

source (for all countries except Germany) was Annual Energy Review, Special Issue-June

1994, Directorate General for Energy (DGXVII); data for Germany include emissions from

plants in the former East Germany.

7. Energy Information Administration (EIA), U. S. Department of Energy, December, 1996a;

yields 1994 average rate for electric utilities of 0.65 (0.64 for 1995) which includes 6% 

transmission and distribution losses. Including nonutility generation, the annual average is about

8% higher.

CO2 emissions numbers given in the “Best” Value column in Tables 19 and 20 were generally

taken from the reference providing the most recent data. In most cases for Europe this was taken from

Michorius’ input. The emissions values for coal, oil, and gas plants (1.11, 0.77, and 0.55, respectively) 
124

Emissions

Source

CO2 Emission Rate

(kg / kWh)

Reference 1 2 3 4 5 7 Range “Best”

Value

Coal Plant 0.993 1.25 0.90 - 1.33 0.89 - 1.11 0.89 - 1.33 1.11

Oil Plant 0.866 0.96 0.65 - 0.89 0.70 - 0.88 0.65 - 0.96 0.77

Gas Plant 0.665 0.583 0.42 - 0.61 0.46 - 0.58 0.42 - 0.67 0.55

future coal plant 0.73 - 0.80 0.73 - 0.80

future oil plant 0.76 0.76

future gas plant 0.35 - 0.40 0.35 - 0.40

Australia 0.82 0.82 0.82

Canada 0.22 0.26 0.22 - 0.26 0.24

Japan 0.39 0.582 0.46 0.473 0.39 - 0.58 0.473

New Zealand 0.13 0.13 0.13

U.S. 0.54 0.650 0.71 0.64-0.65 0.54 - 0.71 0.65

Table 19. Carbon dioxide emissions from power plants.

are averages of the data provided by Caneill, Stuij, and Göricke. Hara’s input was used for the

Japanese average. In comparison, an average value of 0.45 was calculated for Japan for 1993 using

International Energy Agency (IEA) data (OECD/IEA, 1995b) and the average plant emissions factors

above. The average for North America was taken as 0.65 kg CO2/kWh. This is equal to the 1994 EIA

average and to that used in the 1994 AFEAS/DOE report (Fischer). It is approximately equal to the

average of the values provided by Caneill and Stuij and also agrees well with the 0.674 value

determined by Calm (1993). Calm also determined average values for six regions with the U.S., shown

in Table 21.

The 1995 U. S. CO2 emissions rate (0.64) could have been used for the North American

average. However, this data only became available during the final stage of this report’s preparation

and would have the effect of decreasing the indirect TEWI contributions for all electric technologies

considered in North America by only about 1.5%. A change of this magnitude would not cause a

significant change in relative total TEWI between technologies nor to any of the conclusions drawn from

the analyses and comparisons. For this reason it was decided to stay with the 0.65 value for this report.
125

Emissions Source

CO Emission Rate 2

(kg / kWh)

Reference 1 2 3 4 5 6 Range “Best” ValueAustria 0.22 0.00 0.37 0.220 0.00 - 0.37 0.22Belgium 0.30 0.47 0.37 0.35 0.288 0.29 - 0.47 0.29Denmark 1.08 1.01 0.80 0.96 0.844 0.80 - 1.08 0.84Finland 0.30 0.58 0.36 0.235 0.24 - 0.58 0.24France 0.11 0.27 0.13 0.08 0.093 0.08 - 0.27 0.09Germany 0.54 0.70 0.67 0.61 0.597 0.54 - 0.70 0.61Greece 0.92 0.77 1.23 0.86 0.976 0.77 - 1.23 0.98Iceland 0.15 0.00 0.00 - 0.15 0.00Ireland 0.78 0.80 0.91 0.80 0.700 0.70 - 0.91 0.70Italy 0.55 0.59 0.60 0.63 0.586 0.55 - 0.63 0.59Luxembourg 0.72 0.29 1.083 0.29 - 1.08 1.08Netherlands 0.58 0.67 0.66 0.637 0.58 - 0.67 0.64Norway 0.005 0.01 0.00 0.00 - 0.005 0.00Portugal 0.59 0.67 0.51 0.638 0.51 - 0.67 0.64Spain 0.43 0.53 0.40 0.483 0.40 - 0.53 0.48Sweden 0.04 0.23 0.04 0.044 0.04 - 0.23 0.04Switzerland 0.04 0.05 0.13 0.04 - 0.13 0.08Turkey 0.55 0.70 0.55 - 0.70 0.62U.K. 0.67 0.89 0.84 0.75 0.640 0.64 - 0.89 0.64European Average 0.513 0.50 0.427-0.469 0.43 - 0.51 0.47OECD Europe 0.50OECD 0.58

Table 20. Carbon dioxide emissions from power plants in Europe.
126

Region

CO2 Emissions

(kg/kWh)

East Central 0.939

South Central 0.737

West Central 0.672

Southeast 0.671

West 0.497

Northeast 0.489

Table 21. U.S. regional average

CO emissions from power plants 2

(Calm 1993).

Region

CO2 Emissions

(kg/kWh) Region

CO2 Emissions

(kg/kWh)

Africa Average

 Egypt

 South Africa

 Zaire

0.77

0.53

1.03

0.02

Former USSR Average

 Russia

 Tajikistan

 Ukraine

0.44

0.47

0.02

0.48

Asia Average (w/o China)

 Australia

 China

 India

 New Zealand

 Singapore

0.66

0.82

0.88

0.83

0.13

0.77

Non-OECD Europe

 Bulgaria

 Hungary

 Poland

 Romania

0.79

0.59

0.51

1.07

0.64

Latin America Average

 Argentina

 Brazil

 Columbia

0.14

0.24

0.04

0.22

Middle East Average

 Israel

 Saudi Arabia

 Syria

0.63

0.83

0.68

0.32

World Average 0.58

Table 22. 1993 average electric power plant CO emissions for other regions. 2

However, historical data from this source indicate that

carbon emissions from fossil-fueled power plants

in North America are falling gradually. As older, less

efficient plants are retired this trend should continue in all

regions.

The range of values for the North American (from

EIA, 1996a) and European annual averages suggest that an

uncertainty band of ±3-4% is appropriate for the regional

estimates used in the bulk of the comparisons in this study.

Representative values for other regions for 1993 are

given in Table 22. These data were determined by using the

average plant emissions factors from Table 18 and average

generation mixes based on data compiled by IEA

(OECD/IEA 1995a, 1995b). In general, these averages and

ranges are of the same order of magnitude as those for

Europe, Japan, and the U.S.
0.00

0.00

0.04

0.08

0.09

0.22

0.24

0.29

0.48

0.59

0.61

0.62

0.64

0.64

0.64

0.70

0.84

0.98

1.08

0%

5%

10%

15%

20%

25%

30%

Average CO Emission Rates (kg/kWh)

Percent of European Electrical Production

2

Iceland

Norway

Sweden

Switzerland

France

Austria

Finland

Belgium

Spain

Italy

Germany

Turkey

Netherlands

Portugal

United Kingdom

Ireland

Denmark

Greece

Luxemburg

127

Figure 38. Average power plant emission rates and electricity production for European nations.

A.2 SENSITIVITY OF TEWI TO CO2 EMISSION RANGES

As the preceding section indicates, individual country average emission rates vary widely

around the regional averages. Figure 38 illustrates the range of average country emission rates for

several western European nations and compares those with relative percentages of electricity produced

by each country. Indirect (and overall) TEWI values depend directly on the assumed CO2 emission rate

and conclusions drawn can vary. As an example, Fig. 39 shows TEWIs for two representative

European refrigerators, one installed in a country with no fossil-fuel electricity dependence (0.0 kg

CO /kWh) and the other in a country heavily dependent on coal-fired generation (0.98 kg CO /kWh). 2 2

In the latter case, improvements in electricity generation efficiency and/or a switch to different fuel

sources would be much more effective for TEWI reduction than a change in refrigerant or blowing

agent. The direct TEWI does not change in either case and is about an order of magnitude less than the

indirect TEWI for the coal-generation dominated situation.
0.00 0.98

0

1,000

2,000

3,000

4,000

5,000

6,000

Power Plant Emission Rate (kg CO /kWh)

TEWI (kg CO )

Indirect Effect

(energy)

Direct Effect

(blowing agent)

Direct Effect

(refrigerant)

2

2

128

Figure 39. TEWI for European refrigerator/freezer using high (0.98 kg

CO2/kWh) and low (0.0 kg CO2/kWh) emission rates for power plants

(HFC-134a, HCFC-141b, 15 year lifetime, 200 kWh/y).

A.3 IMPACT OF FUEL

PROCESSING AND

TRANSPORTATION

ENERGY

CONSUMPTION ON CO2

EMISSIONS

The previous two

studies did not consider this

effect in making the indirect

TEWI estimates because it

was felt that the magnitude of

the change to CO2 emission

rates would be small and

would, thus, not impact the

comparisons we were making.

However, if an absolute value

for the TEWIs of competing

alternatives was desired for a

particular application, this

impact would of course need

to be quantified. Data from a

paper by Beggs (1996) was

used to attempt to estimate the

impact of fuel processing and

transportation energy on the average CO2 emissions from electricity generation for the United Kingdom

(UK). Using the data presented by Beggs it is possible to come up with the following estimates for

increases in plant emission rates:

coal plants: add 4.5% to the plant rate;

gas plants: add 8.9% to the plant rate;

oil plants: add 10.9% to the plant rate; and

nuclear plants: use 0.005 kg CO /kWh. 2

Using these factors and applying them to the 1993 UK electricity supply mix from IEA (1995b) yields

an overall increase in the average CO2 emission rate of about 5.5%. This would raise the average value

for the UK in Table 20 from 0.64 to 0.675 kg CO /kWh. The absolute impact would be smaller for 2

countries with a heavier dependence on nuclear or hydro power and would be larger where use of coal,

oil, or gas is greater. Adding this effect to indirect contributions from electric power would slightly

increase TEWIs for all options which use electric power. However, the relative differences in TEWI
129

Fuel Source

Quarter

1st 2nd 3rd 4th Year

Coal 56.1% 54.2% 55.7% 59.6% 56.4%

Oil 2.9% 1.7% 2.2% 1.9% 2.2%

Natural Gas 5.8% 9.6% 11.6% 6.8% 8.5%

Nuclear 22.9% 21.9% 21.3% 21.6% 21.9%

Hydroelectric and Other 12.3% 12.6% 9.1% 10.1% 11.0%

Source: EIA 1997.

Table 23. U.S. electricity generation mix (1996).

between any two options would not change significantly. Therefore, omitting this effect from the TEWI

estimates does not affect the overall conclusions significantly.

A.4 IMPACT OF AVERAGE, SEASONAL, AND TIME-OF-DAY VARIATIONS IN CO2

EMISSION RATES

Electricity consumption for certain applications considered in this study tends to occur more

heavily at specific times of the year and during certain periods of each day. Residential and commercial

building air-conditioning and heating are the prime examples. Clearly, CO2 emission rates from electric

power production are not constant. They vary with region, time-of-day, and season as the generation

mix changes. In the UK for example, Beggs (1996) cites a 30% to 47% reduction in CO2 emissions in

summer nighttime hours relative to afternoon hours and an 11% to 22% reduction in winter for the UK.

He also shows that electric generation needs and CO emissions are highest in winter and, further, his 2

time-of-day average emissions estimates range from 0.23 to 0.63 kg CO2/kWh which are lower than

the 1992 overall average in Table 20 above (0.64). For the U.S., data for 1996 electricity generation

indicated that the utility generation mix remained fairly constant with natural gas generation peaking in

the spring and summer, coal in fall and winter, hydro in winter and spring, and oil and nuclear staying

relatively flat, Table 23 (EIA 1997). Historical data for 1994/1995 from EIA (1996b) mirror these

trends.

These variations

notwithstanding, a yearly

average CO2 emission value

was sufficient for the primary

purposes of the present

analysis -- a relative

comparison between the

overall global warming

impacts of HCFC

refrigerants and blowing

agents and those of

alternative fluids and

technologies. As noted in the

Executive Summary,

absolute TEWI comparisons,

particularly between options that use different fuels, are difficult. If a detailed and rigorous evaluation of

TEWI for competing air-conditioning or refrigeration technology options is required for a particular

location, then an hourly simulation of the systems coupled with an hourly schedule for the local

electricity source generation mix and CO2 emissions would be needed. Conclusions from such an

analysis would be applicable only to the specific location considered. In order to meet one of the major

goals of the project, to identify opportunities for CO2 emissions reduction that are broadly applicable
0.54 kW/ton

triple effect

0

10,000,000

20,000,000

30,000,000

40,000,000

TEWI (kg CO )

0.60 kW/ton

double effect

0.54 kW/ton

triple effect

0.60 kW/ton

double effect

0.54 kW/ton

triple effect

0.60 kW/ton

double effect

0.54 kW/ton

triple effect

0.60 kW/ton

double effect

0.54 kW/ton

triple effect

0.60 kW/ton

double effect

0.54 kW/ton

triple effect

0.60 kW/ton

double effect

direct (refrigerant)

indirect (gas)

indirect (electricity)

Utility B: coal and

gas base, oil and

gas intermediate

Utility B: coal and

gas base, oil and

gas intermediate

Utility A: nuclear

base, coal

intermediate

Utility A: nuclear

base, coal

intermediate

Utility C: coal

base, gas

intermediate

Utility C: coal

base, gas

intermediate

Seasonal Average Emission Rates Peak and Intermediate Power

Emission Rates

2

130

Figure 40. Sensitivity of TEWI for chillers to power plant CO emission rates. 2

across regions, these detailed analyses would have to be done for several locations in each region and

for each technology option considered to fairly illustrate how the comparisons would vary throughout

the regions. Such analyses are well beyond the scope of and time available for this project. A less

rigorous case study was conducted to illustrate the sensitivity of TEWI comparisons between electric

and absorption chiller options to different assumptions about electric power CO emissions. 2

Figure 40 illustrates TEWI values for 475 ton (1670 kW) electric powered centrifugal and

direct-fired absorption chillers in three U.S. utility districts selected for the broad differences in their

overall summertime average CO2 emission rates and summertime average emission rates for just peak

and intermediate power generation (Reid 1997). “Utility A” was chosen because it has a low average

emission rate because of a high nuclear capacity for base power generation and a high intermediate and

peak rate because of a high use of coal for intermediate power generation. “Utility C” was selected

because it has a high summer average CO emission rate (large coal base power) and lower 2

intermediate and peak emission rate (predominantly natural gas). “Utility B” in Fig. 40 has summertime

average and intermediate and peak power emission rates that are approximately the same as the

regional average used in the body of this report for North America. A 475 ton (1670 kW) chiller was

used for these calculations because it is the weighted average nominal size for the U.S. (Hourahan

1996a). Operating hours were adjusted to coincide with the utility locations.
131

References

1. Reid 1997. Personal communication from E. Reid of the American Gas Cooling Center

(AGCC) to J. Sand of Oak Ridge National Laboratory.
133

APPENDIX B: ATMOSPHERIC LIFETIMES AND GWPs FOR REFRIGERANTS AND

BLOWING AGENTS
134
135

Notes for Tables 23 and 24:

a. CFC and HCFC lifetimes and 100 year GWP values are taken from the 1995 Radiative

Forcing of Climate Change Report to IPCC from the Scientific Assessment Working Group

(WB1), Table 5.2 CFCs are not listed in the 1995 IPCC report.

b. When listed, the lifetimes and 100 year GWP values for HFC refrigerants and blowing agents

and methane are taken from the 1995 Radiative Forcing of Climate Change Report to IPCC

from the Scientific Assessment Working Group, Table 2.9.

c. The atmospheric lifetimes and 100 year GWPs given in Tables B.1 and B.2 for HFCs and HCs

not specifically listed in the 1994 or 1995 reports are from the open literature or estimates

provided by AFEAS member companies. This includes values for the zeotropic and azeotropic

mixtures.

References

1. IPCC 1994. Radiative Forcing of Climate Change 1994: Report to IPCC from the

Scientific Assessment Working Group. IPCC-X/DOC3 Part 1. Intergovernmental Panel on

Climate Change p. SPM-21.

2. IPCC 1995. Climate Change 1995: The Science of Climate Change, Working Group I,

Cambridge University Press, 1966.

3. B. Orfeo. Private communications from Bob Orfeo of AlliedSignal Inc., Fluorine Products

Division to Van Baxter of Oak Ridge National Laboratory, July 17, 1996 and July 29, 1996.
136

Refrigerant

Atmospheric Lifetime

(years) ODP

GPW 

(100 year ITH) Refrigerant

Atmospheric Lifetime

(years) ODP

GWP

(100 year ITH)

CFCs HFCs (cont)CFC-11 50±5 1.0 4000 HFC-152a 1.5 0 140 

CFC-12 102 1.0 8500 HFC-227ea 36.5 0 2900 CFC-113 85 0.8 5000 HFC-236ea 8 0 710 

CFC-114 300 1.0 9300 HFC-236fa 209 0 6300 CFC-115 1700 0.6 9300 HFC-245ca 6.6 0 560 

HCFCs HFC-245ea c 5 0 320 

HCFC-22a 13.3 0.055 1700 HFC-245eb 3.9 0 380 

HCFC-123a 1.4 0.02 93 HFC-245fa 7.3 0 820 

HCFC-124a 5.9 0.022 480 HFC-356mcf 1.3 0 125 HCFC-141ba 9.4 0.11 630 HFC-356mffm 7.0 0 760 HCFC-142ba 19.5 0.065 2000 HFC-365mfc 10.8 0 840 

HCFC-225caa 2.5 0.025 170 HFC-43-10mee 17.1 0 1300 

HFCs HCsHFC-23b 264 0 11,700 HC-50 (methane) 12.2±3 0 21 

HFC-32b 5.6 0 650 HC-170 (ethane) -- 0 11 

HFC-125b 32.6 0 2800 HC-290 (propane) -- 0 11 

HFC-134ab 14.6 0 1300 HC-3(11)0 (butane) -- 0 11 HFC-143ab 48.3 0 3800 Cyclo-Pentane -- 0 11 

Table 24. Atmospheric lifetimes and 100 year GWPs for refrigerants and blowing agents.
137

Zeotropes and

Azeotropes

Atmospheric Lifetime

(years) ODP

GWP

(100 year ITH)

R-401Ac ---- 0.037 970 

R-401Bc ---- 0.04 1060 

R-401Cc ---- 0.03 760 

R-402Ac ---- 0.021 2250 

R-402Bc ---- 0.033 1960 

R-403Bc 0.031 3570 

R-404Ac ---- 0 3260 

R-405Ac ---- 0.028 4480 

R-406Ac ---- 0.057 1560 

R-407Ac ---- 0 1770 

R-407Bc ---- 0 2290 

R-407Cc ---- 0 1530 

R-408Ac ---- 0.026 2650 

R-409Ac ---- 0.048 1290 

R-410Ac ---- 0 1730 

R-411Ac ---- 0.048 1330 

R-411Bc ---- 0.052 1410 

R-500 ---- 0.74 6010 

R-501 ---- 0.29 3150 

R-502 ---- 0.33 5260 

R-503 ---- 0.6 11,350 

R-504 ---- 0.31 4890 

R-507c ---- 0 3300 

Table 25.. Atmospheric lifetimes and 100 year GWPs for zeotropic and azeotropic

mixtures.
138
139

APPENDIX C: SPREADSHEETS FOR REFRIGERATOR/FREEZER TEWI

CALCULATIONS
140
141

Baseline Information Internal Volume Foam Volume Refrigerant Charge HCFC-141b Blown Foam Annual Energy Use CO Emission Rate 2

510

0.2670.155

12.0%

646.5

0.65

litersm3kg HFC-134ablowing agentkWh/yearkg CO /kWhe 2

Blowing Agent

Refrigerant: HFC-134a

Refrigerant: 

HC-290/HC-600a

HCFC-141b

Vacuum

Panels

HCFC-22/

HCFC-142b HFC-134a HFC-365mfc HFC-356mffm HFC-245fa

HCFC-22/

HCFC-141b HFC-236ea

PentaneIsomers

Carbon

Dioxide Cyclopentane

Carbon

Dioxide

Foam Volume (m )3 Density (kg/m )3 Mass (kg)

0.26732.60

9.02

0.23532.60

7.94

0.26733.99

9.08

0.26737.80

7.85

0.26732.90

7.48

0.26737.10

8.98

0.26733.00

8.96

0.26732.83

8.77

0.26731.40

8.96

0.26736.90

8.96

0.26733.64

8.98

0.26736.90

7.16

0.26733.60

8.97

Blowing Agent Weight % Mass (kg) GWP

11.84%

1.031

630

11.84%

0.907

630

11.00%

0.998

1700

8.28%

0.848

1300

10.50%

0.922

840

10.68%

1.058

760

12.60%

1.110

820

10.80%

0.947

1165

13.08%

1.097

710

5.07%

0.500

3

2.00%

0.180

0

5.07%

0.500

3

2.00%

0.179

1

Refrigerant Mass (kg) GWP

0.155

1300

0.155

1300

0.155

1300

0.155

1300

0.155

1300

0.155

1300

0.155

1300

0.155

1300

0.155

1300

0.155

1300

0.155

1300

0.05

3

0.05

3

Direct Effect Refrigerant Blowing agent

884202649

802202571

1899

202

1697

1304

202

1086

861202775

1020

202804

1127

202910

1304

202

1102

1003

202779

203202

1

202202

0

20

1.4

00

0.2

Energy Use Daily (kWh) Annual (kWh)

1.77

646.5

1.672591.7

1.86

679.4

2.01

734.2

1.93

704.9

1.84

672.1

1.79

653.8

1.77

646.5

1.85

675.1

1.95

712.2

1.98

723.2

1.95

712.2

1.98

723.2

Conversion Factors Lifetime (years) CO factor (kg/kWh) 2

15

0.65

15

0.65

15

0.65

15

0.65

15

0.65

15

0.65

15

0.65

15

0.65

15

0.65

15

0.65

15

0.65

15

0.65

15

0.65

Indirect Effect 6303 5769 6624 7158 6873 6553 6375 6303 6588 6944 7051 6944 7051TEWI % Direct Effect

7154

11.89%

65.42

11.82%

8446

22.28%

8446

15.41%

7849

12.44%

7558

13.30%

7415

14.99%

7608

17.15%

7568

12.95%

7147

2.84%

7253

2.78%

6946

0.02%

7051

0.00%

Table 26.. TEWI calculations for a North American refrigerator/freezer.
142

Baseline Information Internal Volume Foam Volume Refrigerant Charge HCFC-141b Blown Foam Annual Energy Use CO Emission Rate 2

230

0.1539

0.045

5.4%

300

0.470

litersm3kg HC-600ablowing agentkWh/yearkg CO /kWhe 2

Blowing Agent

Refrigerant: HC-600a (isobutane) Refrigerant: HFC-134a

Cyclopentane HCFC-141b HFC-245fa HFC-134a Cyclopentane HCFC-141b HFC-245fa HFC-134a

Foam Volume (m )3 Density (kg/m )3 Mass (kg)

0.1539

37.5

5.163

0.1539

33.805.202

0.1539

33.545.162

0.1539

37.8

4.525

0.1539

37.5

5.163

0.1539

33.805.202

0.1539

33.545.162

0.1539

37.8

4.525

Blowing Agent Weight % Mass (kg) GWP

5.1%

0.279

11

10.0%

0.520

630

11.5%

0.594

820

8.3%

0.366

1300

5.1%

0.279

11

10.0%

0.520

630

11.5%

0.594

820

8.3%

0.366

1300

Refrigerant Mass (kg) GWP

0.045

11

0.045

11

0.045

11

0.045

11

0.127

1300

0.127

1300

0.127

1300

0.127

1300

Direct Effect Refrigerant Blowing agent

3.6

0.503.22

328.2

0.50

327.7

487.3

0.50

486.8

476/9

0.50

627.7

168.2

165.10

3.22

492.8

165.10

327.7

651.9

165.10

486.8

641.5

165.10

627.7

Energy Use Daily (kWh) Annual (kWh)

0.90

328

0.82

300

0.82

300

0.85

312

0.90

328

0.82

300

0.82

300

0.85

312

Conversion Factors Lifetime (years) CO Emissions (kg/kWh) 2

15

0.470

15

0.470

15

0.470

15

0.470

15

0.470

15

0.470

15

0.470

15

0.470

Indirect Effect 2312.4 2115 2115 2200 2312.4 2115 2115 2200TEWI % Direct Effect

2316

0.15%

2443

15.52%

2602

23.04%

2828

28.56%

2481

7.29%

2608

23.30%

2767

30.82%

2992

36.04%

Table 27. TEWI calculations for a European refrigerator/freezer.
143

Baseline Information Internal Volume Foam Volume Refrigerant Charge HCFC-141b Blown Foam Annual Energy Use CO Emission Rate 2

355

0.1794

0.140

11.0%

522

0.473

litersm3kg HFC-134ablowing agentkWh/yearkg CO /kWhe 2

Blowing Agent

Refrigerant: HFC-134a Refrigerant: HC-600a

HCFC-141b HCFC-22/

HCFC-142b

HFC-152a HFC-236ea HFC-245ca HFC-245fa HFC-356mffm Pentane

Isomers

Pentane Blowing Agents

Foam Volume (m )3 Density (kg/m )3 Mass (kg)

0.1794

36

6.46

0.1794

34

6.10

0.1794

29.45.27

0.1794

33.54

6.02

0.1794

33.54

6.02

0.1794

32

5.74

0.1794

33.63

6.03

0.1794

37

6.64

0.1794

37

6.64

Blowing Agent Weight % Mass (kg) GWP

11.00%

0.710

630

11.00%

0.671

1820

8.10%

0.427

140

11.90%

0.716

710

11.90%

0.716

560

12.60%

0.723

820

11.30%

0.682

760

5.45%

0.362

3

5.45%

0.362

3

Refrigerant Mass (kg) GWP

0.140

1300

0.140

1300

0.140

1300

0.140

1300

0.140

1300

0.140

1300

0.140

1300

0.140

1300

0.080

3

Direct Effect Refrigerant Blowing agent

630182448

1403

182

1221

242182

60

690182508

583182401

775182593

700182518

183

1

182

101

Energy Use Daily (kWh) Annual (kWh)

1.43

522

1.43

522

1.67

611

1.49

543

1.50

548

1.43

522

1.48

540

1.57

572

1.57

572

Conversion Factors Lifetime (years) CO Factor (kg/kWh) 2

15

0.473

15

0.473

15

0.473

15

0.473

15

0.473

15

0.473

15

0.473

15

0.473

15

0.473

Indirect Effect 3704 3704 4335 3853 3888 3704 3831 4058 4058TEWI % Direct Effect

4333

14.5%

5107

27.5%

4577

5.3%

4543

15.2%

4471

13.0%

4479

17.3%

4531

15.5%

4241

4.3%

4060

0.0%

Table 28. TEWI calculations for a Japanese refrigerator/freezer.
144
145

APPENDIX D: AVERAGED HEATING AND COOLING LOADS AND HEAT PUMP

PERFORMANCE FACTORS FOR NEW SINGLE-FAMILY HOMES AND

COMMERCIAL BUILDINGS IN EUROPE
146
147

Housing Unit

1991 new stock

(units)

average load

(kWh/y)

Single-family homes -- heating

 France

 Germany

 Greece

 Italy

 Netherlands

 Norway

 Sweden

 United Kingdom

 Total

170,200 

122,100 

42,800 

182,000 

65,700 

7,000 

28,800 

210,100 

828,800 

14,700 

24,000 

6,300 

16,000 

8,700 

19,000 

19,000 

17,500 

 (avg) 16,400 

Single-family homes -- cooling

 Southern Europe (Greece)

 Northern Europe (Germany)

42,800 

122,100 

7,500 

100 

Source: International Heat Pump Status and Policy Review, Report

HPC-AR3, IEA Heat Pump Centre, September 1994 (Part 1, Tables 2.5 and

2.6; supplemented with information from Part 2A, National Position Papers --

Italy and The United Kingdom).

Table 29. European residential building data.
148

Country

Electric Absorption

Air-Source Geothermal Air-Source Geothermal

Residential Building Heat Pumps

France

Germany

Greece

Italy

Netherlands

Norway

Sweden

United Kingdom

2.0

2.1 - 2.3

3.2

2.7

2.5 - 3.0

2.0 - 2.5

2.1

2.0

2.5 - 3.0

2.2 - 3.0

3.4

3.0 - 3.5

2.5 - 3.5

2.6

1.1 - 1.3

1.1

1.1 - 1.3

Average 2.5 3.0 1.2 1.2

Commercial Building Heat Pumps

Germany

Greece

Italy

Netherlands

Norway

United Kingdom

2.1 - 2.3

3.1

2.9

2.0 - 2.5

2.5 - 3.5

2.4

Average 2.45

Source: International Heat Pump Status and Policy Review, Report HPC-AR3, IEA Heat Pump Centre,

September, 1994 (Part 1, Tables 4.1 and 4.2).

Note: The reference listed above made no distinction between cooling and heating seasonal performance factors

(SFP). For purposes of the analyses undertaken for this study, the SPFs for the electric systems above

are assumed to be the same whether heating or cooling. The absorption system SPFs are considered to be

for heating only. A value of 0.70 is used for a single-stage absorption cooling system to evaluate this

technology for unitary cooling in the European climates.

Table 30. Average seasonal performance factors for HCFC-22 based heat pumps in Europe.
149

Country

New Building

Stock: 1991

(units)

Average Building

Heating/Cooling Load

(kWh/y)

Commercial Building 

Heating Data

Germany

Greece

Italy

Netherlands

Norway

United Kingdom

35,200

1,100

20,000

900

200

55,900

80,000

22,100

145,000

8,800

1,300,000

120,000

Total 113,200 112,500

Commercial Building

Cooling Data

Germany

Greece

Netherlands

Norway

United Kingdom

35,200

1,100

900

200

55,900

200,000

30,100

6,500

10,000

50,000

Total 93,200 105,800

Table 31. European commercial building data.
150
151

APPENDIX E: UNITARY AIR CONDITIONING TEWI RESULTS
152
153

Equipment Lifetime 15 yearsBuilding Description 1800 

74,700,00016,100,000

ft2Btu/y heatingBtu/y cooling

CO Conversions: 2 0.0559

0.650.19

kg CO / 1000 Btu gas input 2

kg CO /kWh electricity 2

kg CO / 1000 Btu I R 2 2

Equipment Description Refrigerant

Seasonal Efficiency Indirect Effect Direct Effect

TEWI Cooling Heating Electricity Gas Parasitics Refrigerant

1996 Equipment refrigerant emissions: make-up rate: 4% of charge per year, end-of-life loss: 15% of chargeAir-to-Air Heat Pumps HCFC-22

HCFC-22

SEER - 10SEER - 12

HSPF - 7HSPF - 8

122,376107,742

00

00

35703570

125,946111,312

Premium Heat Pump Options Electric Air-to-Air  Geothermal 

HCFC-22HCFC-22

SEER - 14SEER -16.8

HSPF - 9HSPF -12

96,52670,038

00

00

35701785

100,096

71,823

2005 Equipment refrigerant emissions: make-up rate: 2% of charge per year, end-of-life loss: 15% of chargeAir-to-Air Heat Pumps Direct Expansion Secondary Loops 5.0EC )T 10EC )T 

HCFC-22R-407CR-410AHC-290

SEER - 12.0SEER - 12.0SEER - 12.6SEER - 12

HSPF -8.0HSPF - 8.0HSPF - 8.4HSPF - 8

107,742107,742103,113121,514137,953

00000

000

45834583

214219281791

22

111,312109,670104,904117,880132,398

Premium Heat Pump Options Electric Air-to-Air  Geothermal 

HCFC-22HCFC-22

SEER - 16SEER - 17.2

HSPF - 9.5HSPF -12.8

88,95766,027

00

00

21421071

91,09967,098

Table 32. TEWI calculations for unitary space conditioning equipment: Pittsburgh, Pennsylvania USA.
154

Equipment Lifetime 15 yearsBuilding Description 1800 

34,800,00033,800,000

ft2Btu/y heatingBtu/y cooling

CO Conversions 2 0.0559

0.650.19

kg CO / 1000 Btu gas input 2

kg CO /kWh electricity 2

kg CO / 1000 Btu I R 2 2

Equipment Description Refrigerant

Seasonal Efficiency Indirect Effect Direct Effect

TEWI Cooling Heating Electricity Gas Parasitics Refrigerant

1996 Equipment refrigerant emissions: make-up rate: 4% of charge per year, end-of-life loss: 15% of chargeAir-to-Air Heat PumpsCentral A/C and Electric Furnace

HCFC-22HCFC-22HCFC-22

SEER - 10SEER - 12SEER - 12

HSPF - 7HSPF - 8

100%

81,42669,87531,036

000

00

99,440

357035703570

84,99673,445

130,476

Premium Heat Pump Options Electric Air-to-Air  Geothermal 

HCFC-22HCFC-22

SEER - 14SEER - 16.8

HSPF - 9HSPF - 12

61,23947,891

00

00

35701785

64,80949,676

2005 Equipment refrigerant emissions: make-up rate: 2% of charge per year, end-of-life loss: 15% of chargeAir-to-Air Heat Pumps Direct Expansion Secondary Loops 5.0EC )T 10EC )T 

HCFC-22R-407CR-410AHC-290

SEER - 12SEER - 12SEER - 12.6SEER - 12

HSPF - 8HSPF - 8HSPF - 8.4HSPF - 8

69,87569,87566,54875,14987,102

00000

000

3,4663,466

214219281791

22

72,01771,8036,833978,61590,568

Premium Heat Pump Options Electric Air-to-Air  Geothermal 

HCFC-22HCFC-22

SEER - 16SEER -17.2

HSPF - 9.5HSPF - 12.8

56,31345,666

00

00

21421071

58,45546,739

Table 33. TEWI calculations for unitary space conditioning equipment: Atlanta, Georgia USA.
155

Equipment Lifetime 15 yearsBuilding Description 1800 

0

82,200,000

ft2Btu/y heatingBtu/y cooling

0.0559

0.650.19

kg CO / 1000 Btu gas input 2

kg CO /kWh electricity 2

kg CO / 1000 Btu I R 2 2

Equipment Description Refrigerant

Seasonal

Cooling

Efficiency

Indirect Effect Direct Effect

TEWI Electricity Gas Parasitics Refrigerant

1996 Equipment refrigerant emissions: make-up rate: 4% of charge per year, end-of-life loss: 15% of chargeAir-to-Air Heat Pumps HCFC-22

HCFC-22

SEER - 10SEER - 12

80,14566,788

00

00

35703570

83,71570,358

Premium Heat Pump Options Electric Air-to-Air Geothermal 

HCFC-22HCFC-22

SEER - 14SEER - 16.8

57,24647,705

00

00

35701785

60,81649,490

2005 Equipment refrigerant emissions: make-up rate: 2% of charge per year, end-of-life loss: 15% of chargeElectric Air Conditioning Direct Expansion Secondary Loops 5.0EC )T 10EC )T 

HCFC-22R-407CR-410AHC-290HC-290

SEER - 12SEER - 12SEER - 12.6SEER - 12SEER - 12

66,78866,78863,60778,575

92,762 

00000

000

4,1544,154

214219281791

22

68,93068,71665,39882,72996,916

Premium Air-Conditioner Options Electric Air-to-Air  Geothermal 

HCFC-22HCFC-22

SEER - 16 56,313

46,596

00

00

21421071

58,45547,667

Table 34. TEWI calculations for unitary space conditioning equipment: Miami, Florida USA.
156

Equipment Lifetime 15 years

Refrigerant Emissions:

 1996

 2005

 End-of-Life

1.50% of charge per year

1.00% of charge per year

15% of charge 

Refrigerant Charge:

 HCFC-22

 R-407C

 R-410A

6.90 kg

6.90

5.70

Equivalent Full Load

Cooling Hours

 Pittsburgh

 Atlanta

 Miami

600

1400

2700

CO Conversion Factor: 2

0.65 kg CO2/kWh

Cooling Climate

Technology

Base

Equipment

Efficiency

(EER)

Indirect

Effect

(kg CO )2

Direct

Effect

(kg CO )2

TEWI

(kg CO )2

Pittsburgh, Pennsylvania

 HCFC-22 RTU

 HCFC-22

 R-407C

 R-410A

1996

2005

2005

2005

10

11

11

11.6

52,650

47,834

47,834

45,584

4399

3519

3167

2958

57,049

51,383

50,801

48,542

Atlanta, Georgia

 HCFC-22 RTU

 HCFC-22

 R-407C

 R-410A

1996

2005

2005

2005

10

11

11

11.6

122,850

111,682

111,682

106,364

4399

3519

3167

2958

127,249

115,201

114,849

109,322

Miami, Florida

 HCFC-22 RTU

 HCFC-22

 R-407C

 R-410A

1996

2005

2005

2005

10

11

11

11.6

236,925

215,386

215,386

205,130

4399

3519

3167

2958

241,324

218,905

218,553

208,088

Table 35. TEWI for 7.5 ton roof top air conditioners (RTU): cooling only.
157

Equipment Specification: Lifetime Heat pump capacity Annual heating load

1510

16,400

yearskWkW/year

Refrigerant Loss Data: Make-up rate End-of-Life Loss

5% of charge/year15% of charge

CO Conversion Factors: 2 Electricity Generation Natural Gas Combustion

0.4690.190

kg/kWhkg/kWh

Technology

Efficiency

(SPF)

Electricity Refrigerant Emissions

TEWI(kg CO )2

Heating

(kWh/y)

Auxiliary

(kWh/y)

Lifetime CO2

(kg)

Charge

(kg)

Leakage

(g/y)

End-of-Life

(g)

Lifetime

Loss

(kg)

Refrigerant

GWP

(kg CO /kg) 2

Direct

Effect

(kg CO )2

Air-to-Air Heat Pump HCFC-22 HFC-134a R-407C (95% HCFC-22) R-407C (100% HCFC-22) R-410A (105% HCFC-22)

2.502.502.382.502.63

65606560689165606560

00000

4615046150484764615046150

3.12.73.13.12.2

155135155155155

465405465465330

2.792.432.792.791.98

17001300153015301730

47433159426942603425

50,89349,30952,74550,41847,377

Geothermal Heat Pumps HCFC-22 HFC-134a R-407C (95% HCFC-22) R-407C (100% HCFC-22) R-410A (105% HCFC-22)

3.003.002.853.00

3.15 

54675467574554675206

00000

3845838458404823845836627

3.12.73.13.12.2

155135155155155

465405465465330

2.792.432.792.791.98

17001300153015301730

47433159426942693425

43,02141,61744,75142,72740,052

Electric Resistance Heat 1.00 16,400 0 115,374 0 0 0 0 0 0 115,374

Table 36. TEWI for European residential space heating (data from the International Heat Pump Status and Policy Review Report,PHC-AR3, IEA Heat Pump Centre, September 1994, Part 1.
158

Equipment Specification: Lifetime Heat pump capacity Annual cooling load

1510

7,500

yearskWkW/year

Refrigerant EmissionData: Make-up rate End-of-Life Loss

5% of charge/year15% of charge

CO Conversion Factors: 2 Electricity Generation Natural Gas Combustion

0.9760.190

kg/kWhkg/kWh

 Technology

Efficiency

(SPF)

Electricity Refrigerant Emissions

TEWI(kg CO )2

Cooling

(kWh/y)

Lifetime

CO2(kg)

Charge

(kg)

Leakage

(g/y)

End-of-Life

(g)

Lifetime

Loss

(kg)

Refrigerant

GWP

(kg CO /kg) 2

Direct

Effect

(kg CO )2

Central Air Conditioner

 H FC-22 HFC-134a R-407C (95% HCFC-22) R-407C (100% HCFC-22) R-410A (105% HCFC-22)

3.203.203.043.203.36

23442344246723442232

34,31334,31336,11834,31332,679

3.12.73.13.12.2

155135155155155

465405465465330

2.792.432.792.791.98

17001300153015301730

47433159426942603425

39,05637,47240,38738,58134,182

Geothermal Heat Pumps

 HCFC-22 HFC-134a R-407C (95% HCFC-22) R-407C (100% HCFC-22) R-410A (105% HCFC-22)

3.403.403.233.40

 3.57

22062206232222062101

32,29432,29433,99432,29430,756

3.12.73.13.12.2

155135155155155

465405465465330

2.792.432.792.791.98

17001300153015301730

47433159426942693425

37,03735,45338,26336,56334,182

Table 37. TEWI for European residential cooling only (Greece). Source for loads and Seasonal Performance Factors (SPF): InternationalHeat Pump Status and Policy Review Report, HPC-AR3, IEA Heat Pump Centre, September 1994.
159

Refrigerant Emission Scenarios

1⁄2% annual; 15% end-of-life 1⁄2% annual; 10% end-of-life 1% annual; 15% end-of-life 1% annual; 5% end-of-life 1⁄2% annual; 5% end-of-life

R-22 R-407C R-410A R-22 R-407C R-410A R-22 R-407C R-410A R-22 R-407C R-410A R-22 R-407C R-410A

Direct Effectrefrigerant charge (kg)leak rate (kg/y)end-of-life loss (kg)lifetime emissions (kg)GWPDirect Effect (kg CO )2

1.00 

 0.005 

0.15 0.21 

1700 

357 

0.93 

0.005 

0.1395 

0.195 

1530 

299 

1.00 

0.005 

0.15 0.21 

1730 

363 

1.00 

0.005 

0.10 0.16 

1700 

272 

0.93 

0.005 0.093 0.149 

1530 

228 

1.00 

0.005 0.100 

0.16 

1730 

277 

1.00 

0.010 0.150 0.270 

1700 

459 

0.93 

0.009 

0.1395 

0.251 

1530 

384 

1.00 

0.010 

0.15 0.27 

1730 

467 

1.00 

0.010 

0.05 0.17 

1700 

289 

0.93 

0.009 

0.0465 

0.158 

1530 

242 

1.00 

0.010 

0.05 0.17 

1730 

294 

1.00 

0.005 

0.05 0.11 

1700 

187 

0.93 

0.005 

0.0465 

0.102 

1530 

157 

1.00 

0.005 

0.05 0.11 

1730 

190 

Indirect Effect (cooling)COPoperating time (hours/y)system capacity (kW)power input (kWh/y)lifetime input (kWh)Indirect Effect (kg CO )2

2.67 

350 

2.80 

367.0 

4405 

2083 

2.67 

350 2.80

367.0 

4405 2083 

2.67 

350 

2.80 

367.0 

4405 2083 

2.67 

350 

2.80 

367.0 

4405 2083 

2.67 

350 2.80

367.0 

4405 

2083

2.67 

350 

2.80 

367.0 

4405 2083 

2.67 

350 

2.80 

367.0 

4405 2083 

2.67 

350 2.80

367.0 

4405 

2083

2.67 

350 

2.80 

367.0 

4405 2083 

2.67 

350 

2.80 

367.0 

4405 2083 

2.67 

350 2.80

367.0 

4405 

2083

2.67 

350 

2.80 

367.0 

4405 2083 

2.67 

350 

2.80 

367.0 

4405 2083 

2.67 

350 2.80

367.0 

4405 

2083

2.67 

350 

2.80 

367.0 

4405 2083 

Indirect Effects (heating)COPoperating time (hours/y)system capacity (kW)power input (kWh/y)lifetime input (kWh)Indirect Effect (kg CO )2

3.20 

450 

2.80 

393.8 

4725 2235 

3.20 

450 

2.80 

393.8 

4725 2235 

3.20 

450 

2.80 

393.8 

4725 2235 

3.20 

450 

2.80 

393.8 

4725 2235 

3.20 

450 

2.80 

393.8 

4725 2235 

3.20 

450 

2.80 

393.8 

4725 2235 

3.20 

450 

2.80 

393.8 

4725 2235 

3.20 

450 

2.80 

393.8 

4725 2235 

3.20 

450 

2.80 

393.8 

4725 2235 

3.20 

450 

2.80 

393.8 

4725 2235 

3.20 

450 

2.80 

393.8 

4725 2235 

3.20 

450 

2.80 

393.8 

4725 2235 

3.20 

450 

2.80 

393.8 

4725 2235 

3.20 

450 

2.80 

393.8 

4725 2235 

3.20 

450 

2.80 

393.8 

4725 2235 

TEWI (kg CO )2 4675 4617 4681 4590 4546 4595 4777 4702 4785 4607 4560 4612 4505 4475 4508 Note: 2.8 kW split system, 12 year equipment lifetime; 0.473 kg CO / kWh, heating and cooling hours for Tokyo 2

Table 38. TEWI for unitary heating and cooling equipment in Japan.
160

Emission Rates: 0.20% per year, 15% end-of-life Emission Rates: 0.00% per year, 5% end-of-life

HCFC-22 R-407C R-410A HCFC-22 R-407C R-410A

Energy Use (Cooling) relative efficiency COP cooling capacity (kW) power input (kW) operating time (hours/y) power input (kWh/y)

100% 

2.50 14.0 

5.6 

600 

3360 

97% 

2.43 14.0 5.76 

600 

3464 

100% 

2.50 14.0 

5.6 

600 

3360 

100% 

2.50 14.0 

5.6 

600 

3360 

105% 

2.63 14.0 5.32 

600 

3200 

100% 

2.50 14.0 

5.6 

600 

3360 

97% 

2.43 14.0 5.76 

600 

3464 

100% 

2.50 14.0 

5.6 

600 

3360 

100% 

2.50 14.0 

5.6 

600 

3360 

105% 

2.63 14.0 5.32 

600 

3200 

Energy Use(Heating) relative efficiency COP heating capacity (kW) power input operating time (hours/y) power input (kWh/y)

100% 

3.00 16.0 5.33 

900 

4800 

97% 

2.91 16.0 5.49 

900 

4948 

100% 

3.00 16.0 5.33 

900 

4800 

100% 

3.00 16.0 5.33 

900 

4800 

105% 

3.15 16.0 5.08 

900 

4571 

100% 

3.00 16.0 5.33 

900 

4800 

97% 

2.91 16.0 5.49 

900 

4948 

100% 

3.00 16.0 5.33 

900 

4800 

100% 

3.00 16.0 5.33 

900 

4800 

105% 

3.15 16.0 5.08 

900 

4571 

Indirect Effect annual energy use (kWh/y) lifetime energy use (kWh) CO emissions 2

8160 

81,600 38,597 

8412 

84,120 39,791 

8160 

81,600 38,597 

8160 

81,600 38,597 

7771 

77,710 36,759 

8160 

81,600 38,597 

8412 

84,120 39,791 

8160 

81,600 38,597 

8160 

81,600 38,597 

7771 

77,710 36,759 

Direct Effect charge size (kg) refrigerant emissions (g/y) end-of-life emissions (g) lifetime emissions (kg) GWP (kg CO /kg) 2 Equivalent CO emissions 2

4.45 

8.9 

668 

0.757 

1500 1135 

 4.45 

8.9 

668 

0.757 

1530 1157 

4.45 

8.9 

668 

0.757 

1500 1157 

4.45 

8.9 

668 

0.757 

1730 1309 

4.45 

8.9 

668 

0.757 

1730 1309 

4.45 

0 

223 

0.223 

1500 

334 

4.45 

0 

223 

0.223 

1530 

340 

4.45 

0 

223 

0.223 

1530 

340 

4.45 

0 

223 

0.223 

1730 

385 

4.45 

0 

223 

0.223 

1730 

385 

TEWI (kg CO )2 39,732 40,948 39,754 39,906 38,068 38,931 40,131 38,937 38,982 37,144 Note: equipment lifetime 10 years, power generation CO emission factor 0.473 kg CO /kWh. 2 2

Table 39. TEWI for packaged air conditioners (PACs) for Japanese unitary applications.
161

APPENDIX F: CALCULATION OF COOLING TOWER ELECTRICAL BURDEN FOR

LARGE CHILLERS
162
163

THR ' Qcool@

1

COP

% 1

THR ' Qcool@

1

COP

% 1

ABSORPTION CHILLER

Qexhaust

Qfuel

Qaux

Qcool

Qrej

Figure 41. Heat flows for an absorption chiller.

F.1 CHILLER HEAT REJECTION

The total heat rejection for an electric chiller is the sum of the work into the compressor plus the

refrigeration effect. The total heat of rejection (THR) can be expressed in terms of the coefficient of

performance as follows:

If the chiller is hermetic, the COP is defined based on the electrical power supplied to the prime mover

(the electric motor). If the chiller is an open-drive electric, the COP would be defined based on the

shaft work into the compressor (THR is slightly lower since the electric motor heat is not rejected to the

tower).

For absorption chillers, Fig.

41 shows the relevant energy flows

where: Qfuel is the thermal energy

associated with the energy source

“firing” the chiller, Qaux is the energy

associated with providing power to

the chiller’s auxiliaries (solution

pumps, purge units, etc), Qexhaust

is the thermal energy leaving the

unit’s exhaust via combustion

byproducts (or exhaust steam for

indirect fired), Qcool is the heat

absorbed in the evaporator, Qrej is

the total heat needed to be rejected

to a tower. If Qaux.Qexhaust (or if

Qaux and Qexhaust are small, relative to the other energy flows), the relation for the total heat

rejection, in terms of the COP and Qcool, is identical to the electric chiller case.

 

The identical relation given above is true for an engine-driven chiller that rejects both the heat from the
164

1.465 GPM ' Qcool@

1

COP

% 1

kWe pump '

GPMHead

5308@ç p@ç m

'

GPM×40

(5308)(0.6)(0.8)

'

GPM

64

or

GPM ' 64×kWe pump

chiller as well as the engine jacket to the ambient environment through the cooling tower (assuming the

COP is defined based on the fuel input).

Thus given the unit’s COP and cooling capacity, the total heat rejection can be determined.

F.2 ELECTRICAL AUXILIARIES FOR HEAT REJECTION

The simplest approach is to determine the electrical energy (kWe) associated with condenser

water pumps and cooling tower fans on a per unit heat rejection basis (kWt). The approach requires

some encompassing assumptions but is more defensible than the previous approach which was based

on a “rule-of-thumb”.

F.2.1 Condenser Water Pumping

Assuming a 10E F range on the tower (aTcondenser=10EF), the total heat rejection, on a kWt

basis, will be a function of the condenser water flowrate (GPM) as given by:

THR (kWt) = 1.465 GPM

Equating the above equation with the relation previously found for total heat rejection gives:

where: Qcool is given on a kW thermal basis.

The kWe for condenser water pumping can be determined, based on the water-side flow rate by

assuming a condenser water pump efficiency, çp, of 60%; motor efficiency, çm, of 80%; and a waterside pressure drop, head, of 40 ft of H20.
6

 The heat rejection factor is a dimensionless ratio of the total heat rejected to the cooling effect.

165

kWe

pump

'

Qcool

1.465×64

1

COP

% 1

Fan Type

Tower Fan

Ratio

(kWe/kWt)

Axial 0.0128 - 0.0168

Centrifugal 0.0250 - 0.358

Table 40. Cooling tower fan

power.

kWefan

Qcool

'0.0142 HRF

1.25

'0.0114 1

COP

% 1

kWe'

1

93.76

% 0.0114 Qcool@

1

COP

% 1 ' 0.022@ Qcool 1

COP

% 1

Substituting the above expression into the heat rejection balance gives the pumping kWe as a function

of the unit COP and cooling capacity.

 

F.2.2 Cooling Tower Fans

Several different configurations of cooling tower fans are available. The most common type for chiller

applications is an induced draft type with either an axial or centrifugal fan. The data assumes a heat

rejection factor of 1.256 and a 10 F temperature rise on the condenser. 

The range of tower fan power is given in Table 40.

Treating all technologies equally, the tower fan power calculation

can be simplified by assuming:

Thus, the total (pumping and tower fan) electrical energy associated with heat rejection equipment is

given by,

where Qcool is the unit cooling capacity in kW.

Table 41 uses the proposed method to calculate an appropriate cooling tower electrical burden
166

Electric

Chiller

Engine

Driven

Chiller

Direct-Fired Double

Effect Absorption

Chiller

Direct-Fired Triple

Effect Absorption

Chiller

COP 6.40 2.00 1.05 1.45 

Power

 Pump (kW)

 Fan (kW)

 Total

4.33 

4.64 

8.97 

5.62 

6.02 

11.64 

7.31 

7.83 

15.15 

6.34 

6.78 

13.12 

Acknowledgment: this method of assessing the additional loads associated with chiller

tower operation was conceived and derived by D. Reindl at the University of WisconsinMadison.

Table 41. Cooling tower electrical burden.

for each of the technologies assuming a 100 ton (352 kW Qcool) chiller.
167

APPENDIX G: SPREADSHEETS FOR CHILLER TEWI CALCULATIONS
168
169

IPLV

(kW/RT

Cooling

Tower

Burden

(kWe/RT)

Effective

Efficiency

(kW/RT)

Annual

Energy Use

(kWh/y/RT)

Lifetime

Energy

Use

(kWh/RT)

Indirect

Effect

(kg CO )2

0.45 0.087 0.54 1142 34,251 22,263

0.47 0.088 0.56 1185 35,554 23,110

0.49 0.088 0.58 1229 36,857 23,957

0.51 0.089 0.60 1272 38,160 24,804

0.53 0.089 0.62 1315 39,463 25,651

0.55 0.089 0.64 1359 40,766 26,498

0.57 0.090 0.66 1402 42,069 27,345

0.58 0.090 0.67 1424 42,721 27,788

0.59 0.090 0.68 1446 43,372 28,192

0.60 0.091 0.69 1467 44,024 28,615

Note: 2125 annual operating hours, 30 year equipment lifetime, CO2 conversion

factor 0.65 kg CO2/kWhe

Table 42. Parametric analysis of the indirect effects for chillers 
170

Refrigerant

Refrigerant Emission Rate (% of charge/year/RT)

0.0 0.50 1.00 2.00 4.00 10.0

CFC-12

 lifetime refrigerant loss (kg/RT)

 direct effect

 indirect effect

 TEWI per RT

 % direct effect

0.062

527

26,498

27,025

2.0%

0.248

2108

26,498

28,606

7.4%

0.434

3689

26,498

30,187

12.2%

0.806

6851

26,498

33,349

20.5%

1.55

13,175

26,498

39,673

33.2%

3.782

32,147

26,498

58,645

54.8%

CFC-11

 lifetime refrigerant loss (kg/RT)

 direct effect

 indirect effect

 TEWI per RT

 % direct effect

0.050

200

32,252

32,452

0.6%

0.20

800

32,252

33,052

2.4%

0.35

1400

32,252

33,552

4.2%

0.65

2600

32,252

34,852

7.5%

1.25

5000

32,252

37,252

13.4%

3.05

12,200

32,252

44,452

27.4%

HFC-134a

 lifetime refrigerant loss (kg/RT)

 direct effect

 indirect effect

 TEWI per RT

 % direct effect

0.041

52.7

26,498

26,551

0.20%

0.162

211

26,498

26,709

0.79%

0.284

369

26,498

26,870

1.37

0.527

685

26,498

27,180

2.52%

1.013

1316

26,498

27,814

4.73%

2.471

3212

26,498

29,710

10.81%

HCFC-123

 lifetime refrigerant loss (kg/RT)

 direct effect

 indirect effect

 TEWI per RT

 % direct effect

0.0525

5

26,498

26,503

0.02%

0.210

20

26,498

26,518

0.07%

0.368

34

26,498

26,532

0.13%

0.683

63

26,498

26,561

0.24%

1.313

122

26,498

26,620

0.46%

3.203

298

26,498

26,796

1.11%

HCFC-22

 lifetime refrigerant loss (kg/RT)

 direct effect

 indirect effect

 TEWI per RT

 % direct effect

0.041

69

26,498

26,667

0.26%

0.16

275

26,498

26,773

1.03%

0.28

482

26,498

26,980

1.79%

0.53

895

26,498

27,393

3.27%

1.01

1721

26,498

28,219

6.10%

2.47

4,200

26,498

30,980

13.68%

HFC-245ca

 lifetime refrigerant loss (kg/RT)

 direct effect

 indirect effect

 TEWI per RT

 % direct effect

0.038

23

26,498

26,521

0.09%

0.15

92

26,498

26.590

0.35%

0.26

160

26,498

26,658

0.60%

0.49

298

26,498

26,796

1.11%

0.94

572

26,498

27,070

2.11%

2.29

1396

26,498

27,894

5.00%

R-717

 direct effect

 indirect effect

 TEWI per RT

 % direct effect

0.00

26,498

26,498

0.00%

0.00

26,498

26,498

0.00%

0.00

26,498

26,498

0.00%

0.00

26,498

26,498

0.00%

0.00

26,498

26,498

0.00%

0.00

26,498

26,498

0.00%

Notes: IPLV = 0.55 kW/ton, 2125 annual operating hours, 0.65 kg CO2/kWhe

Table 42. Sensitivity of TEWI in chillers to refrigerant make-up rate.
171

1996 Technology IPLV*Indirect Effect

Direct Effect

Refrigerant Make-Up Rate (% charge/year)

0% 0.5% 1.0% 2.0% 4.0%

Centrifugal Chillers:

 CFC-11

 CFC-12

 HCFC-123

 HFC-134a

 HCFC-22

0.58

0.59

0.47

0.54

0.54

27,768,000

28,192,000

23,110,000

26,074,500

26,074,500

200,000

527,000

5,000

52,700

69,000

800,000

2,108,000

20,000

211,000

275,000

1,400,000

3,689,000

34,000

369,000

482,000

2,600,000

6,851,000

63,000

685,000

895,000

5,000,000

13,175,000

122,000

1,316,000

1,721,000

Screw Chillers:

 HCFC-22

 R-717

0.60

0.60

28,615,000

28,615,000

69,000

0

275,000

0

482,000

0

895,000

0

1,721,000

0

HFC-134a Engine Driven Chillers 2.30 23,193,000 52,700 211,000 396,000 685,000 1,316,000

Direct-Fired Double Effect Absorption 1.07 47,577,000 0 0 0 0 0

2005 Technology

Centrifugal Chillers:

 HCFC-123

 HFC-134a

 HCFC-22

0.45

0.48

0.48

22,263,000

23,533,500

23,533,500

5,000

52,700

69,000

20,000

211,000

275,000

34,000

369,000

482,000

63,000

685,000

895,000

122,000

1,316,000

1,721,000

Screw Chillers:

 HCFC-22

 R-717

0.58

0.57

27,788,000

27,345,000

69,000

0

275,000

0

482,000

0

895,000

0

1,721,000

0

Engine Driven Chillers:

 HCFC-22

 HFC-134a

2.40

2.40

22,360,000

22,360,000

69,000

52,700

275,000

211,000

482,000

369,000

895,000

685,000

1,721,000

1,316,000

Direct-Fired Absorption Chillers:

 Double Effect

 Triple Effect

1.15

1.50

44,592,000

35,712,000

0

0

0

0

0

0

0

0

0

0

IPLVs given as kW/RT for electric powered chillers and gas COPs for thermally powered chillers, 3500 kW (1000 RT) cooling capacity, 30 year *

equipment lifetime

Table 44. TEWI for 3500 kW (1000 RT) chillers using data from Tables 42 and 43.
172

Double Effect Chiller IPL (gCOP) Triple Effect Chiller IPL

(gCOP)

1.00 1.05 1.07 1.15 1.45 1.50

Gas:

 g CO2/1000 Btu

 kg CO2/12,000 Btu

 gas COP

 kg CO2/hton

 kg CO2/year/ton

 kg CO2/ton

55.9

0.671

1.00

0.671

1,424

42,713

55.9

0.671

1.05

0.638

1,356

40,679

55.9

0.671

1.07

0.626

1,331

39,918

55.9

0.671

1.15

0.583

1,238

37,141

55.9 

0.671 

1.45 

0.462 

982 

29,457 

55.9 

0.671 

1.50 

0.447 

949 

28,475 

Electricity:

 kW/ton cooling

 gas COP

 peripherals (kW/ton)

 cooling tower burden (kW/ton)

 kW/ton - total

 kWh/year/ton

 kWh/ton lifetime

 kg CO2/ton

3.516

1.00

0.035

0.155

0.1899

403.5

12,106

7,869

3.516

1.05

0.035

0.151

0.1862

395.7

11,871

7,716

3.516

1.07

0.035

0.150

0.1848

392.8

11,783

7,659

3.516

1.15

0.035

0.145

0.1798

382.1

11,463

7,451

3.516 

1.45 

0.046 

0.131 

0.1764 

374.9 

11,248 

7,311 

3.516 

1.50 

0.046 

0.129 

0.1747 

371.1 

11,134 

7,237 

TEWI (per Rton) 50,581 48,395 47,577 44,592 36,768 35,712 

Note: 2125 annual operating hours (1200 EFLH adjusted for IPLV) 30 year lifetime, 0.65 kg CO2/kWhe, 55.9 g

CO /1000 Btu (assumes 96.5% distribution efficiency). 2

Table 45. TEWI for direct-gas fired absorption chillers.
173

Source of Emissions

Primary Fuel IPLVs (COP)

1.90 1.95 2.10 2.20 2.30 2.40

Indirect Effect (gas):

 g CO2/1000 Btu

 kg CO2/12,000 Btu

 kg CO2/hr/RT w/ gCOP

 kg CO2/year/ton

 Indirect Effect (kg CO /Rton) 2

55.9

0.671

0.353

750

22,507

55.9

0.671

0.344

731

21,930

55.9

0.671

0.319

679

20,364

55.9

0.671

0.305

648

19,438

55.9

0.671

0.292

620

18,593

55.9

0.671

0.280

594

17,818

Indirect Effect (electric) :

*

 kW/RT

 kWh/year/RT

 kWh/ton-lifetime

 kg CO2/kWh

 Indirect Effect (kg CO /Rton) 2

0.118

251

7528

0.65

4893

0.117

249

7462

0.65

4850

0.114

243

7281

0.65

4733

0.113

239

7174

0.65

4663

0.111

236

7077

0.65

4600

0.110

233

6987

0.65

4542

Total Indirect Effect (kg CO /Rton) 2 27,401 26,780 25,096 24,101 23,193 22,360

Note: 2125 annual operating hours, 30 year lifetime, 55.9 g CO2/1000 Btu natural gas (96.5%

distribution efficiency)

cooling tower contribution only *

Table 46. TEWI for natural gas engine driven chillers using HCFC-22, HFC-134a, or R-717

(screw and centrifugal compressors, HFC-134a).

Refrigerant Loss Data HCFC-22 HFC-134a R-717

Charge (kg/ton) 0.81 0.81 0.46 

Make-Up Rate (kg/ton/year) 0.0162 0.0162 0.0092 

Lifetime Emissions (kg/ton)

 make-up

 end-of-life

 total

 

0.486 

0.041 

0.527 

 0.486 

0.041 

0.527 

0.276 

0.023 

0.299 

GWP 1700 1300 0 

Direct Effect (kg CO /ton) 2 895 685 0 

Note: 2.0% of charge per year make-up rate, 30 year lifetime, end-of-life

loss of 5% of charge

Table 47. Direct effects taken from Table 43.
174

HCFC-123 HFC-134a

Double Effect Absorption

Chillers

Direct

Fired

IndirectFired

Direct Effect

 refrigerant charge (kg)

 loss rate (% charge/year)

 annual losses (kg)

 lifetime losses (kg)

 end-of-life losses (kg)

 lifetime emissions (kg)

 GWP

 direct effect

255 

1.0% 

2.55 

63.75 

12.8 

76.5 

93 

7115 

255

0.50%

1.28

31.87

12.8

44.6

1300

58,013

0 

0 

0 

0 

0 

0 

0 

0 

0 

0 

0 

0 

0 

0 

0 

0 

Indirect Effect: Electricity

 a. power consumption (kW/RT)

 1055 kW (300 RT) chiller power (kW)

 annual power consumption (kWh)

 b. cooling tower power per ton (kW/RT)

 cooling tower for 1055 kW (300 RT) (kW) 

 annual cooling tower power (kWh)

 c. absorption auxiliaries 1055 kW chiller (kW)

 annual absorption auxiliary (kWh)

 d. lifetime electrical energy use (kWh)

 e. lifetime CO2 emissions

0.70 

210 

147,000 

0.14 

42 

29,400 

0 

0 

4,410,000 

2,085,930 

0.70

210

147,000

0.14

42

29,400

0

0

4,410,000

2,085,930

0.035 

0 

0 

0.22 

66 

45,465 

10.5 

7350 

1,320,375 

624,537 

0.025 

0 

0 

0.22 

66 

45,465 

7.5 

5,250 

1,267,875 

599,705 

Indirect Effect: Natural Gas

 system COP 

 boiler efficiency

 gas consumption 1055 kW (300 RT) chiller (kW)

 annual gas consumption (kWh)

 CO emissions (kg/year) 2

 lifetime CO emissions (kg) 2

NA 

NA 

NA 

NA 

NA 

NA 

NA 

NA 

NA 

NA 

NA 

NA 

1.00 

----- 

1055 

738,500 

140,868 

3,521,700 

1.00 

80% 

1319 

923,125 

176,085 

4,402,125 

TEWI (kg CO )2 2,093,045 2,143,943 4,146,237 5,001,183 

Notes: equipment lifetime 25 years, CO2 conversion factor 0.473 kg CO2/kWhe, typical chiller capacity 1055 kW

(300 RT), 0.671 kg CO2/12,000 Btu gas input (0.190 kg CO2/kWh gas input), 700 annual full load operating

hours, 20% auxiliary power consumption

Table 48. TEWI for centralized chiller applications in Japan.
175

APPENDIX H: SECONDARY HEAT TRANSFER LOOP CALCULATIONS FOR

COMMERCIAL REFRIGERATION
176
177

Secondary Loop Parameters

North America Europe Japan

low

temp

medium

temp

low

temp

medium

temp

low

temp

medium

temp

Refrigeration Load (kW) 88 264 44 132 24 127 

Secondary Loop )T (EC) 2.8 2.8 2.8 2.8 2.8 2.8 

Pipe Parameters

 equivalent length (m)

 e/D

 ID (mm)

122 

0.0012 

53 

122 

0.0012 

53 

61 

0.0012 

53 

61 

0.0012 

53 

46 

0.0012 

41 

46 

0.0012 

53 

Fluid Properties

 density (kg/m )

3

 velocity (m/s)

 dynamic viscosity (mPa-s)

 kinematic viscosity (mm /sec) 2

 thermal conductivity (W/mEC)

 specific heat (kJ/kgEC)

1218 

1.3 

0.942 

23.3 

0.439 

2.976 

1207 

1.0 

5.87 

7.10 

0.455 

3.036 

1218 

2.0 

19.2 

23.3 

0.439 

2.976 

1207 

2.0 

5.87 

7.10 

0.455 

3.036 

1218 

1.8 

19.2 

23.3 

0.439 

2.976 

1207 

1.2 

5.87 

7.10 

0.455 

3.036 

Fluid Characteristics

 Reynolds number

 Prandtl number

 Nusselt number

 friction factor

3027 

6.4 

2.6 

0.0206 

7369 

39.2 

9.6 

0.0206 

4541 

130.0 

9.7 

0.0206 

14,700 

39.2 

16.7 

0.0250 

3162 

130.0 

7.2 

0.0206 

8507 

39.2 

10.7 

0.0206 

Pressure Drop (kPa) 105.3 57.3 118.3 114.5 91.1 28.6 

Pumping Power (kW) 1.2 2.0 0.7 2.0 0.3 0.5 

Annual Energy Input (kWh) 10,700 17,300 6035 17,300 2520 4160 

Table 49. Secondary loop pumping power for commercial refrigeration applications.
178
179

APPENDIX I: COMPLETE RESULTS FOR COMMERCIAL REFRIGERATION
180
181

R-404A

R-507

R-407A

R-407C

R-404A

R-507

R-407A

R-407C

R-717

R-404A

R-507

R-407A

R-407C

0 200,000 400,000 600,000 800,000 1,000,000

TEWI (kg CO )2

Indirect (Energy)

Direct (minimum)

Direct (incremental)

DIRECTEXPANSION

SECONDARYLOOP

DISTRIBUTEDSYSTEMS

Figure 43. Low temperature refrigeration in Japan.

R-404A

R-507

R-407A

R-407C

R-404A

R-507

R-407A

R-407C

R-717

R-404A

R-507

R-407A

R-407C

0 500,000 1,000,000 1,500,000 2,000,000

Indirect (Energy)

Direct (minimum)

Direct (incremental)

DIRECTEXPANSION

SECONDARYLOOP

DISTRIBUTEDSYSTEMS

TEWI (kg CO )2

Figure 42. Low temperature refrigeration in Europe.
R-404A

R-507

R-407A

R-407C

R-404A

R-507

R-407A

R-407C

R-717

R-404A

R-507

R-407A

R-407C

0 1,000,000 2,000,000 3,000,000 4,000,000 5,000,000

Indirect (Energy)

Direct (minimum)

Direct (incremental)

DIRECTEXPANSION

SECONDARYLOOP

DISTRIBUTED

SYSTEMS

TEWI (kg CO )2

R-404A

R-507

R-134a

R-410A

R-404A

R-507

R-134a

R-410A

R-717

R-404A

R-507

R-134a

R-410A

0 500,000 1,000,000 1,500,000 2,000,000 2,500,000

Indirect (Energy)

Direct (minimum)

Direct (incremental)

TEWI (kg CO ) 2

DIRECT

EXPANSION

SECONDARYLOOP

DISTRIBUTED

SYSTEMS

182

Figure 44. Low temperature refrigeration in North America.

Figure 45. Medium temperature refrigeration in Europe.
R-404A

R-507

R-134a

R-410A

R-404A

R-507

R-134a

R-410A

R-717

R-404A

R-507

R-134a

R-410A

0 1,000,000 2,000,000 3,000,000 4,000,000 5,000,000 6,000,000

Indirect (Energy)

Direct (minimum)

Direct (incremental)

TEWI (kg CO ) 2

DIRECT

EXPANSION

SECONDARYLOOP

DISTRIBUTED

SYSTEMS

R-404A

R-507

R-407A

R-407C

R-404A

R-507

R-407A

R-407C

R-717

R-404A

R-507

R-407A

R-407C

0 200,000 400,000 600,000 800,000 1,000,000

TEWI (kg CO )2

Indirect (Energy)

Direct (minimum)

Direct (incremental)

DIRECT

EXPANSION

SECONDARY

LOOP

DISTRIBUTED

SYSTEMS

183

Figure 47. Medium temperature refrigeration in North America.

Figure 46. Medium temperature refrigeration in Japan.
184

 Component of TEWI

Low Temperature Refrigeration Medium Temperature Refrigeration

R-404A R-507 R-407A R-407C R-404A R-507 R-134a R-410A

 1. Compressor Calculations a. design load (kW) b. fraction on-time c. cooling delivered (kWh/y) d. pumping power (kWh/y) e. total cooling required (kWh/y) f. COP g. compressor input (kWh/y)

87.9

40%

307,999

0

307,999

1.762

174.797

87.9

40%

307,999

0

307,999

1.762

174,797

87.9

40%

307,999

0

307,999

1.708

180,317

87.9

40%

307,999

0

307,999

1.654

186,197

263.7

40%

923,966

0

923,996

3.425

269,800

263.7

40%

923,966

0

923,996

3.515

262,882

263.7

40%

923,966

0

923,996

3.317

278,598

263.7

40%

923,966

0

923,996

3.569

258,899

 2. Condenser Calculations a. refrigeration load (kWh/y) b. pumping power (kWh/y) c. compressor power (kWh/y) d. total heat rejected (kWh/y) e. required fan power (kWh/y)

307,999

0

174,797482,796

8,835

307,999

0

174,797482,796

8,835

307,999

0

180,317488,315

8,936

307,999

0

186,197494,195

9,044

923,996

0

269,800

1,193,796

21,846

923,996

0

262,882

1,186,878

21,720

923,996

0

278,598

1,202,594

22,007

923,996

0

258,899

1,182,895

21,647

 3. Total Energy Consumption a. compressor (kWh/y) b. secondary fluid pump (kWh/y) c. condenser fan (kWh/y) d. total annual energy (kWh/y) e. lifetime energy (kWh) f. lifetime CO emissions (kg) 2

174,797

0

8,835

183,632

2,754,4781,790,411

174,797

0

8,835

183,632

2,754,4781,790,411

180,317

0

8,936

189,253

2,838,7921,845,215

186,197

0

9,044

195,240

2,928,6041,903,593

269,800

0

21,846

291,646

4,374,6922,843,550

262,882

0

21,720

284,602

4,269,0242,774,866

278,598

0

22,007

300,605

4,509,0752,930,899

258,899

0

21,6647280,546

4,208,1852,735,320

 4. Refrigerant Emissions a. refrigerant GWP b. refrigerant charge (kg) c. current practice (13.5% loss/year) (1) lifetime emissions (kg) (2) direct effect d. near term practice (6% loss/year) (1) lifetime emissions (kg) (2) direct effect

3260

323654

2,133,177

291

948,079

3300

319646

2,132,317

287

947,697

1770

353716

1,266,949

318

563,088

1530

347704

1,076,358

313

478,381

3260

402814

2,653,710

362

1,179,427

3300

397804

2,652,640

357

1,178,951

1300

467947

1,230,459

421

546,871

1730

397804

1,390,626

357

618,056

 TEWI a. current practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI b. near term practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI

1,790,4112,133,2003,923,6001,790,411

948,100

2,738,500

1,790,4112,132,3003,922,7111,790,411

947,700

2,738,100

1,845,2001,266,9003,112,1001,845,200

563,100

2,408,300

1,903,6001,076,4002,980,0001,903,600

478,400

2,382,000

2,843,6002,653,7005,497,3002,843,6001,179,4004,023,000

2,774,9002,652,6005,427,5002,774,9001,179,0003,953,900

2,930,9001,230,5004,161,4002,930,900

546,900

3,477,800

2,735,3001,390,6004,125,9002,735,300

618,100

3,353,400

Table 50. TEWI for low and medium temperature direct expansion refrigeration systems in North America.
185

Component of TEWI

Low Temperature Refrigeration Medium Temperature Refrigeration

R-404A R-507 R-407A R-407C R-404A R-507 R-134a R-410A

1. Compressor Calculations a. design load (kW) b. fraction on-time c. cooling delivered (kWh/y) d. pumping power (kWh/y) e. total cooling required (kWh/y) f. COP g. compressor input (kWh/y)

87.9

40%

307,999

4,380

312,379

1.692

184,669

87.9

40%

307,999

4,380

312,379

1.692

184,669

87.9

40%

307,999

4,380

312,379

1.640

190,501

87.9

40%

307,999

4,380

312,379

1.588

196,713

263.7

40%

923,996

12,895

936,891

3.219

291,026

263.7

40%

923,996

12,895

936,891

3.304

283,564

263.7

40%

923,996

12,895

936,891

3.118

300,516

263.7

40%

923,996

12,895

936,891

3.355

279,268

2. Condenser Calculations a. refrigeration load (kWh/y) b. pumping power (kWh/y) c. compressor power (kWh/y) d. total heat rejected (kWh/y) e. required fan power (kWh/y)

312,379

4,380

184,669501,428

9,176

312,379

4,380

184,669501,428

9,176

312,379

4,380

190,501507,260

9,283

312,379

4,380

196,713513,472

9,397

936,891

12,895

291,026

1,240,812

22,707

936,891

12,895

283,564

1,233,350

22,570

936,891

12,895

300,516

1,250,302

22,881

936,891

12,895

279,268

1,229,054

22,492

3. Total Energy Consumption a. compressor (kWh/y) b. secondary fluid pump (kWh/y) c. condenser fan (kWh/y) d. total annual energy (kWh/y) e. lifetime energy (kWh) f. lifetime CO emissions (kg) 2

184,669

4,3809,176

198,225

2,973,3811,932,698

184,669

4,3809,176

198,225

2,973,3811,932,698

190,501

4,3809,283

204,164

3,062,4571,990,597

196,713

4,3809,397

210,489

3,157,3422,052,272

291,026

12,89522,707

326,628

4,899,4203,184,623

283,564

12,89522,570

319,029

4,785,4393,110,535

300,516

12,89522,881

336,292

5,044,3753,278,844

279,268

12,89522,492

314,654

4,719,8133,067,878

4. Refrigerant Emissions a. refrigerant GWP b. refrigerant charge (kg) c. current practice (5% loss/year) (1) lifetime emissions (kg) (2) direct effect d. near term practice (2% loss/year) (1) lifetime emissions (kg) (2) direct effect

3260

8161

197,516

24

79,007

3300

8060

197,437

24

78,975

1770

8866

117,310

27

46,924

1530

8765

99,663

26

39,865

3260

100

75

245,714

30

98,286

3300

9974

245,615

30

98,246

1300

117

88

113,931

35

45,573

1730

9974

128,762

30

51,505

TEWI a. current practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI b. near term practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI

1,932,700

197,500

2,130,2001,932,700

79,000

2,011,700

1,932,700

197,400

2,130,1001,932,700

79,000

2,011,700

1,990,600

117,300

2,107,9001,990,600

46,900

2,037,500

2,052,300

99,700

2,152,0002,052,300

39,900

2,092,200

3,184,600

245,700

3,430,3003,184,600

98,300

3,282,900

3,110,500

245,600

3,356,1003,110,500

98,200

3,208,700

3,278,800

113,900

3,392,7003,278,800

45,600

3,324,400

3,067,900

128,800

3,196,7003,067,900

51,500

3,119,400

Table 51. TEWI for low and medium temperature distributed refrigeration systems in North America.
186

Component of TEWI

Low Temperature Refrigeration Medium Temperature Refrigeration

R-404A R-507 R-407A R-407C R-404A R-507 R-134a R-410A

1. Compressor Calculations a. design load (kW) b. fraction on-time c. cooling delivered (kWh/y) d. pumping power (kWh/y) e. total cooling required (kWh/y) f. COP g. compressor input (kWh/y)

43.9

40%

153,999

0

153,999

1.762

87,398

43.9

40%

153,999

0

153,999

1.762

87,398

43.9

40%

153,999

0

153,999

1.708

90,158

43.9

40%

153,999

0

153,999

1.654

93,098

131.8

40%

461,998

0

461,998

3.425

134,900

131.8

40%

461,998

0

461,998

3.515

131,441

131.8

40%

461,998

0

461,998

3.317

139,299

131.8

40%

461,998

0

461,998

3.569

129,449

2. Condenser Calculations a. refrigeration load (kWh/y) b. pumping power (kWh/y) c. compressor power (kWh/y) d. total heat rejected (kWh/y) e. required fan power (kWh/y)

153,999

0

87,398

241,398

4,418

153,999

0

87,398

241,398

4,418

153,999

0

90,158

244,158

4,468

153,099

0

93,098

247,098

4,522

461,998

0

134,900596,898

10,923

461,998

0

131,441593,439

10,860

461,998

0

139,299601,297

11,004

461,998

0

129,449591,448

10,823

3. Total Energy Consumption a. compressor (kWh/y) b. secondary fluid pump (kWh/y) c. condenser fan (kWh/y) d. total annual energy (kWh/y) e. lifetime energy (kWh) f. lifetime CO emissions (kg) 2

87,398

0

4,418

91,816

1,377,239

647,302

87,398

0

4,418

91,816

1,377,239

647,302

90,158

0

4,468

94,626

1,419,396

667,116

93,098

0

4,522

97,6620

1,464,302

688,222

134,900

0

10,923

145,823

2,187,3461,028,053

131,441

0

10,860

142,301

2,134,5121,003,221

139,299

0

11,004

150,302

2,254,5371,059,633

129,449

0

10,823

140,273

2,104,093

988,923

4. Refrigerant Emissions a. refrigerant GWP b. refrigerant charge (kg) c. current practice (13.5% loss/year) (1) lifetime emissions (kg) (2) direct effect d. near term practice (6% loss/year) (1) lifetime emissions (kg) (2) direct effect

3260

162327

1,066,589

145

474,039

3300

160323

1,066,159

144

473,848

1770

177358

633,474

159

281,544

1530

174352

538,179

156

239,191

3260

201407

1,326,855

181

589,713

3300

198402

1,326,320

179

589,476

1300

234473

615,230

210

273,435

1730

198402

695,313

179

309,028

TEWI a. current practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI b. near term practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI

647,300

1,066,6001,713,900

647,300474,000

1,121,300

647,300

1,066,2001,713,500

647,300473,800

1,121,100

667,100633,500

1,300,600

667,100281,500948,600

688,200538,200

1,226,400

688,200239,200927,400

1,028,1001,326,9002,355,0001,028,100

589,700

1,617,800

1,003,2001,326,3002,329,5001,003,200

589,500

1,592,700

1,059,600

615,200

1,674,8001,059,600

273,400

1,333,000

988,900695,300

1,684,200

988,900309,000

1,297,900

Table 52. TEWI for low and medium temperature direct expansion refrigeration systems in Europe.
187

Component of TEWI

Low Temperature Refrigeration Medium Temperature Refrigeration

R-404A R-507 R-407A R-407C R-717 R-404A R-507 R-134a R-410A R-717

1. Compressor Calculations a. design load (kW) b. fraction on-time c. cooling delivered (kWh/y) d. pumping power (kWh/y) e. total cooling required (kWh/y) f. COP g. compressor input (kWh/y)

87.9

40%

307,999

10,775

318,774

1.603

198,804

87.9

40%

307,999

10,775

318,774

1.603

198,804

87.9

40%

307,999

10,775

318,774

1.554

205,082

87.9

40%

307,999

10,775

318,774

1.505

211,770

87.9

40%

307,999

10,775

318,774

1.638

194,633

263.7

40%

923,996

17,345

941,341

3.117

302,049

263.7

40%

923,996

17,345

941,341

3.199

294,304

263.7

40%

923,996

17,345

941,341

3.018

311,898

263.7

40%

923,996

17,345

941,341

3.248

289,845

263.7

40%

923,996

17,345

941,341

3.205

293,701

2. Condenser Calculations a. refrigeration load (kWh/y) b. pumping power (kWh/y) c. compressor power (kWh/y) d. total heat rejected (kWh/y) e. required fan power (kWh/y)

318,774

10,775

198,804528,353

9,669

318,774

10,775

198,804528,352

9,669

318,774

10,775

205,082534,631

9,784

318,774

10,775

211,770541,318

9,906

318,774

10,775

194,633524,182

9,592

923,996

17,345

302,049

1,234,390

22,754

923,996

17,345

294,304

1,235,645

22,612

923,996

17,345

311,898

1,253,239

22,934

923,996

17,345

289,845

1,231,186

22,531

923,996

17,345

293,701

1,235,043

22,601

3. Total Energy Consumption a. compressor (kWh/y) b. secondary fluid pump (kWh/y) c. condenser fan (kWh/y) d. total annual energy (kWh/y) e. lifetime energy (kWh) f. lifetime CO emissions (kg) 2

198,804

10,775

9,669

219,248

3,288,7202,137,688

198,804

10,775

9,669

219,248

3,288,7162,137,665

205,082

10,775

9,784

225,641

3,384,6092,199,669

211,700

10,775

9,906

232,450

3,486,7572,266,392

194,633

10,775

9,592

215,000

3,225,0032,096,252

302,049

17,34522,754

342,147

5,132,2123,335,938

294,304

17,34522,612

334,261

5,013,9133,259,044

311,898

17,64522,934

352,177

5,282,6563,433,727

289,845

17,34522,531

329,720

4,945,8023,214,771

293,701

17,34522,601

333,647

5,004,7113,253,062

4. Refrigerant Emissions a. refrigerant GWP b. refrigerant charge (kg) c. current practice (4% loss/year) (1) lifetime emissions (kg) (2) direct effect d. near term practice (2% loss/year) (1) lifetime emissions (kg) (2) direct effect

3260

3621

69,526

11

34,763

3300

3521

69,498

11

34,749

1770

3923

41,293

12

20,647

1530

3823

35,081

11

17,541

0

1811

050

3260

4427

86,491

13

43,246

3300

4426

86,456

13

43,228

1300

5131

40,104

15

20,052

1730

4426

45,324

13

22,662

0

2314

070

TEWI a. current practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI b. near term practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI

2,137,700

69,500

2,207,2002,137,700

34,800

2,172,500

2,137,700

69,500

2,207,2002,137,700

34,700

2,172,400

2,200,000

41,300

2,241,3002,200,000

20,600

2,220,600

2,266,400

35,100

2,301,5002,266,400

17,500

2,283,900

2,096,252

0

2,096,2522,096,252

0

2,096,252

3,335,900

86,500

3,422,4003,335,900

43,200

3,379,100

3,259,000

86,500

3,345,5003,259,000

43,200

3,302,200

3,433,700

40,100

3,473,8003,433,700

20,100

3,453,800

3,214,800

45,300

3,260,1003,214,800

22,700

3,327,500

3,253,100

0

3,253,1003,253,100

0

3,253,100

Table 53. TEWI for low and medium temperature secondary loop refrigeration systems in North America.
188

Component of TEWI

Low Temperature Refrigeration Medium Temperature Refrigeration

R-404A R-507 R-407A R-407C R-717 R-404A R-507 R-134a R-410A R-717

1. Compressor Calculations a. design load (kW) b. fraction on-time c. cooling delivered (kWh/y) d. pumping power (kWh/y) e. total cooling required (kWh/y) f. COP g. compressor input (kWh/y)

43.9

40%

153,999

6,044

160,044

1.603

99,812

43.9

40%

153,999

6,044

160,044

1.603

99,812

43.9

40%

153,999

6,044

160,044

1.554

102,964

43.9

40%

153,999

6,044

160,044

1.505

106,321

43.9

40%

153,999

6,044

160,044

1.638

97,718

131.8

40%

461,998

17,345

479,343

3.117

153,807

131.8

40%

461,998

17,345

479,343

3.199

149,863

131.8

40%

461,998

17,345

479,343

3.018

158,822

131.8

40%

461,998

17,345

479,343

3.248

147,593

131.8

40%

461,998

17,345

479,343

3.205

149,556

2. Condenser Calculations a. refrigeration load (kWh/y) b. pumping power (kWh/y) c. compressor power (kWh/y) d. total heat rejected (kWh/y) e. required fan power (kWh/y)

160,044

6,044

99,812

265,900

4,866

160,044

6,044

99,812

265,900

4,866

160,044

6,044

102,964269,052

4,924

160,044

6,044

106,321272,409

4,985

160,044

6,044

97,718

263,806

4,828

479,343

17,345

153,807650,495

11,904

479,343

17,345

149,863646,551

11,832

479,343

17,345

158,822655,510

11,996

479,343

17,345

147,593644,280

11,790

479,343

17,345

149,556646,244

11,826

3. Total Energy Consumption a. compressor (kWh/y) b. secondary fluid pump (kWh/y) c. condenser fan (kWh/y) d. total annual energy (kWh/y) e. lifetime energy (kWh) f. lifetime CO emissions (kg) 2

99,812

6,0444,866

110,722

1,660,832

780,591

99,812

6,0444,866

110,722

1,660,832

780,591

102,964

6,0444,924

113,932

1,708,976

803,219

106,321

6,0444,985

117,351

1,760,261

827,323

97,718

6,0444,828

108,590

1,628,848

765,559

153,807

17,34511,904

183,056

2,745,8381,290,544

149,863

17,34511,832

179,040

2,685,5991,262,232

158,822

17,34511,996

188,163

2,822,4471,326,550

147,593

17,34511,790

176,728

2,650,9161,245,930

149,556

17,34511,826

178,728

2,680,9131,260,029

4. Refrigerant Emissions a. refrigerant GWP b. refrigerant charge (kg) c. current practice (4% loss/year) (1) lifetime emissions (kg) (2) direct effect d. near term practice (2% loss/year) (1) lifetime emissions (kg) (2) direct effect

3260

1811

34,763

5

17,381

3300

1811

34,749

5

17,374

1770

1912

20,647

6

10,323

1530

1911

17,541

6

8,770

095030

3260

2213

43,246

7

21,623

3300

2213

43,228

7

21,614

1300

2615

20,052

8

10,026

1730

2213

22,662

7

11,331

0

11

7030

TEWI a. current practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI b. near term practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI

780,600

34,800

815,400780,600

17,400

798,000

780,600

34,700

815,300780,600

17,400

798,000

803,200

20,600

823,800803,200

10,300

813,500

827,300

17,500

844,800827,300

8,800

836,100

765,600

0

765,600765,600

0

765,600

1,290,500

43,200

1,333,7001,290,500

21,600

1,312,100

1,262,200

43,200

1,305,4001,262,200

21,600

1,283,800

1,326,500

20,100

1,346,6001,326,500

10,000

1,336,500

1,245,900

22,700

1,268,6001,245,900

11,300

1,257,200

1,260,000

0

1,260,0001,260,000

0

1,260,000

Table 54. TEWI for low and medium temperature secondary loop refrigeration systems in Europe.
189

Component of TEWI

Low Temperature Refrigeration Medium Temperature Refrigeration

R-404A R-507 R-407A R-407C R-404A R-507 R-134a R-410A

1. Compressor Calculations a. design load (kW) b. fraction on-time c. cooling delivered (kWh/y) d. pumping power (kWh/y) e. total cooling required (kWh/y) f. COP g. compressor input (kWh/y)

43.9

40%

153,999

2,873

156,873

1.692

92,739

43.9

40%

153,999

2,873

156,873

1.692

92,739

43.9

40%

153,999

2,873

156,873

1.640

95,667

43.9

40%

153,999

2,873

156,873

1.588

98,787

131.8

40%

461,998

8,585

470,583

3.219

146,177

131.8

40%

461,998

8,585

470,583

3.304

142,429

131.8

40%

461,998

8,585

470,583

3.118

150,944

131.8

40%

461,998

8,585

470,583

3.355

140,271

2. Condenser Calculations a. refrigeration load (kWh/y) b. pumping power (kWh/y) c. compressor power (kWh/y) d. total heat rejected (kWh/y) e. required fan power (kWh/y)

156,873

2,873

92,739

252,485

4,620

156,873

2,873

92,739

252,485

4,620

156,873

2,873

95,667

255,413

4,674

156,873

2,873

98,787

258,533

4,731

470,583

8,585

146,177625,345

11,444

470,583

8,585

142,429621,597

11,375

470,583

8,585

150,944630,112

11,531

470,583

8,585

140,271619,439

11,336

3. Total Energy Consumption a. compressor (kWh/y) b. secondary fluid pump (kWh/y) c. condenser fan (kWh/y) d. total annual energy (kWh/y) e. lifetime energy (kWh) f. lifetime CO emissions (kg) 2

92,739

2,8734,620

100,232

1,503,485

706,638

92,739

2,8734,620

100,232

1,503,485

706,638

95,667

2,8734,674

103,214

1,548,217

727,662

98,787

2,8734,731

106,391

1,595,867

750,058

146,177

8,585

11,444

166,206

2,493,0871,171,751

142,429

8,585

11,375

162,389

2,435,8361,144,843

150,944

8,585

11,531

171,060

2,565,8951,205,971

140,271

8,585

11,336

160,192

2,402,8731,129,350

4. Refrigerant Emissions a. refrigerant GWP b. refrigerant charge (kg) c. current practice (5% loss/year) (1) lifetime emissions (kg) (2) direct effect d. near term practice (2% loss/year) (1) lifetime emissions (kg) (2) direct effect

3260

4030

98,758

12

39,503

3300

4030

98,718

12

39,487

1770

4433

58,655

13

23,462

1530

4333

49,831

13

19,933

3260

5038

122,857

15

49,143

3300

5037

122,807

15

49,123

1300

5844

56,966

18

22,786

1730

5037

64,381

15

25,752

TEWI a. current practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI b. near term practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI

706,600

98,800

805,400706,600

39,500

746,100

706,600

98,700

805,300706,600

39,500

746,100

727,700

58,700

786,400727,700

23,500

751,200

750,100

49,800

799,900750,100

19,900

770,000

1,171,800

122,900

1,294,7001,171,800

49,100

1,220,900

1,144,800

122,800

1,267,6001,144,800

49,100

1,93,900

1,206,000

57,000

1,263,0001,206,000

22,800

1,228,800

1,129,400

64,400

1,193,8001,129,400

25,800

1,155,200

Table 55. TEWI for low and medium temperature distributed refrigeration systems in Europe.
190

Component of TEWI

Low Temperature Refrigeration Medium Temperature Refrigeration

R-404A R-507 R-407A R-407C R-404A R-507 R-134a R-410A

1. Compressor Calculations a. design load (kW) b. fraction on-time c. cooling delivered (kWh/y) d. pumping power (kWh/y) e. total cooling required (kWh/y) f. COP g. compressor input (kWh/y)

23.8

40%

83,519

0

83,519

1.762

47,399

23.8

40%

83,519

0

83,519

1.762

47,399

23.8

40%

83,519

0

83,519

1.708

48,896

23.8

40%

83,519

0

83,519

1.654

50,490

126.8

40%

444,442

0

444,442

3.425

129,774

126.8

40%

444,442

0

444,442

3.515

126,446

126.8

40%

444,442

0

444,442

3.317

134,005

126.8

40%

444,442

0

444,442

3.569

124,530

2. Condenser Calculations a. refrigeration load (kWh/y) b. pumping power (kWh/y) c. compressor power (kWh/y) d. total heat rejected (kWh/y) e. required fan power (kWh/y)

83,519

0

47,399

130,918

2,396

83,519

0

47,399

130,918

2,396

83,519

0

48,896

132,415

2,423

83,519

0

50,490

134,009

2,452

444,442

0

129,774574,216

10,508

444,442

0

126,446570,888

10,447

444,442

0

134,005578,448

10,586

444,442

0

124,530568,973

10,412

3. Total Energy Consumption a. compressor (kWh/y) b. secondary fluid pump (kWh/y) c. condenser fan (kWh/y) d. total annual energy (kWh/y) e. lifetime energy (kWh) f. lifetime CO emissions (kg) 2

47,399

0

2,396

49,795

746,923353,294

47,399

0

2,396

49,795

746,923353,294

48,896

0

2,423

51,319

769,786364,109

50,490

0

2,452

52,943

794,140375,628

129,774

0

10,508

140,282

2,104,227

995,299

126,446

0

10,447

136,893

2,053,401

971,259

134,005

0

10,586

144,591

2,168,8651,025,873

124,530

0

10,412

134,942

2,024,137

957,417

4. Refrigerant Emissions a. refrigerant GWP b. refrigerant charge (kg) c. current practice (13.5% loss/year) (1) lifetime emissions (kg) (2) direct effect d. near term practice (6% loss/year) (1) lifetime emissions (kg) (2) direct effect

3260

64

129

421,303

57

187,246

3300

63

128

421,133

57

187,170

1770

70

141

250,222

63

111,210

1530

69

139

212,581

62

94,480

3260

240486

1,585,462

216

704,650

3300

237480

1,584,822

213

704,366

1300

279565

735,139

251

326,728

1730

237480

830,831

213

369,258

TEWI a. current practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI b. near term practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI

353,300421,300774,600353,300187,200540,500

353,300421,100774,400353,300187,200540,500

364,100250,200614,300364,100111,200475,300

375,600212,600588,200375,600

94,500

470,100

995,300

1,585,5002,580,800

995,300704,600

1,699,900

971,300

1,584,8002,556,100

971,300704,400

1,675,700

1,025,900

735,100

1,761,0001,025,900

326,700

1,352,600

957,400830,800

1,788,200

957,400369,300

1,326,700

Table 56. TEWI for low and medium temperature direct expansion refrigeration systems in Japan.
191

Component of TEWI

Low Temperature Refrigeration Medium Temperature Refrigeration

R-404A R-507 R-407A R-407C R-717 R-404A R-507 R-134a R-410A R-717

1. Compressor Calculations a. design load (kW) b. fraction on-time c. cooling delivered (kWh/y) d. pumping power (kWh/y) e. total cooling required (kWh/y) f. COP g. compressor input (kWh/y)

23.8

40%

83,519

2,540

86,059

1.603

53,671

23.8

40%

83,519

2,540

86,059

1.603

53,671

23.8

40%

83,519

2,540

86,059

1.554

55,366

23.8

40%

83,519

2,540

86,059

1.505

57,171

23.8

40%

83,519

2,540

86,059

1.638

52,545

126.8

40%

444,442

4,205

448,647

3.117

143,958

126.8

40%

444,442

4,205

448,647

3.199

140,266

126.8

40%

444,442

4,205

448,647

3.018

148,652

126.8

40%

444,442

4,205

448,647

3.248

138,141

126.8

40%

444,442

4,205

448,647

3.205

139,979

2. Condenser Calculations a. refrigeration load (kWh/y) b. pumping power (kWh/y) c. compressor power (kWh/y) d. total heat rejected (kWh/y) e. required fan power (kWh/y)

86,059

2,540

53,671

142,271

2,604

86,059

2,540

53,671

142,271

2,604

86,059

2,540

55,366

143,966

2,635

86,059

2,540

57,171

145,771

2,668

86,059

2,540

52,545

141,145

2,583

448,647

4,205

143,958596,810

10,922

448,647

4,205

140,266593,118

10,854

448,647

4,205

148,652601,504

11,008

448,647

4,205

138,141590,993

10,815

448,647

4,205

139,979592,831

10,849

3. Total Energy Consumption a. compressor (kWh/y) b. secondary fluid pump (kWh/y) c. condenser fan (kWh/y) d. total annual energy (kWh/y) e. lifetime energy (kWh) f. lifetime CO emissions (kg) 2

53,671

2,5402,604

58,815

882,227417,293

53,671

2,5402,604

58,815

882,227417,293

55,366

2,5402,635

60,541

908,116429,539

57,171

2,5402,668

62,380

935,693442,583

52,545

2,5402,583

57,669

865,029409,159

143,958

4,205

10,922

159,084

2,386,2601,128,701

140,266

4,205

10,854

155,325

2,329,8791,102,033

148,652

4,205

11,008

163,864

2,457,9631,162,616

138,141

4,205

153,161

2,297,4171,086,678

139,979

4,205

10,849

155,033

2,325,4931,099,958

4. Refrigerant Emissions a. refrigerant GWP b. refrigerant charge (kg) c. current practice (4% loss/year) (1) lifetime emissions (kg) (2) direct effect d. near term practice (1) lifetime emissions (2% loss/year) (2) direct effect

3260

74

13,731

2

6,866

3300

74

13,726

2

6,863

1770

85

8,155

2

4,078

1530

85

6,929

2

3,464

042010

3260

2616

51,674

8

25,837

3300

2616

51,653

8

25,827

1300

3118

23,960

9

11,980

1730

2616

27,079

8

13,539

0

14

8040

TEWI a. current practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI b. near term practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI

417,300

13,700

431,000417,300

6,900

424,200

417,300

13,700

431,000417,300

6,900

424,200

429,500

8,200

437,700429,500

4,100

433,600

442,600

6,900

449,500442,600

3,500

446,100

409,200

0

409,200409,200

0

409,200

1,128,700

51,700

1,180,4001,128,700

25,800

1,154,500

1,102,000

51,700

1,153,7001,102,000

25,800

1,127,800

1,162,600

24,000

1,186,6001,162,600

12,000

1,174,600

1,086,700

27,100

1,113,8001,086,700

13,500

1,100,200

1,100,000

0

1,100,0001,100,000

0

1,100,000

Table 57. TEWI for low and medium temperature secondary loop refrigeration systems in Japan.
192

Component of TEWI

Low Temperature Refrigeration Medium Temperature Refrigeration

R-404A R-507 R-407A R-407C R-404A R-507 R-134a R-410A

1. Compressor Calculations a. design load (kW) b. fraction on-time c. cooling delivered (kWh/y) d. pumping power (kWh/y) e. total cooling required (kWh/y) f. COP g. compressor input (kWh/y)

23.8

40%

83,519

1,752

85,271

1.692

50,410

23.8

40%

83,519

1,752

85,271

1.692

50,410

23.8

40%

83,519

1,752

85,271

1.640

52,002

23.8

40%

83,519

1,752

85,271

1.588

53,697

126.8

40%

444,442

8,760

453,202

 3.219

140,778

126.8

40%

444,442

8,760

453,202

3.304

137,168

126.8

40%

444,442

8,760

453,202

3.118

145,369

126.8

40%

444,442

8,760

453,202

3.355

135,090

2. Condenser Calculations a. refrigeration load (kWh/y) b. pumping power (kWh/y) c. compressor power (kWh/y) d. total heat rejected (kWh/y) e. required fan power (kWh/y)

85,271

1,752

50,410

137,433

2,515

85,271

1,752

50,410

137,433

2,515

85,271

1,752

52,002

139,025

2,544

85,271

1,752

53,697

140,720

2,575

453,202

8,760

140,778602,741

11,030

453,202

8,760

137,168599,131

10,964

453,202

8,760

145,369

6607,331

11,114

453,202

8,760

135,090597,052

10,926

3. Total Energy Consumption a. compressor (kWh/y) b. secondary fluid pump (kWh/y) c. condenser fan (kWh/y) d. total annual energy (kWh/y) e. lifetime energy (kWh) f. lifetime CO emissions (kg) 2

50,410

1,7522,515

54,677

820,151387,932

50,410

1,7522,515

54,677

820,151387,932

52,002

1,7522,544

56,298

844,467399,433

53,697

1,7522,575

58,025

870,368411,684

140,778

8,760

11,030

160,568

2,408,5251,139,232

137,168

8,760

10,964

156,893

2,353,3891,113,153

145,369

8,760

11,114

165,243

2,478,6441,172,399

135,090

8,760

10,926

154,776

2,321,6441,098,137

4. Refrigerant Emissions a. refrigerant GWP b. refrigerant charge (kg) c. current practice (5%/year loss) (1) lifetime emissions (kg) (2) direct effect d. near term practice (2%/year loss) (1) lifetime emissions (kg) (2) direct effect

3260

1612

39,009

5

15,604

3300

1612

38,994

5

15,598

1770

1713

23,169

5

9,267

1530

1713

19,683

5

7,873

3260

6045

146,802

18

58,721

3300

5944

146,743

18

58,697

1300

7052

68,068

21

27,227

1730

5944

76,929

18

30,772

TEWI a. current practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI b. near term practice (1) indirect effect (energy use) (2) direct effect (emissions) (3) TEWI

387,900

39,000

426,900

87,900

15,6600403,500

387,900

39,000

426,900

87,900

15,6600403,500

399,400

23,200

422,600399,400

9,300

408,700

411,700

19,700

431,400411,700

7,900

419,600

1,139,200

146,800

1,286,0001,139,200

58,700

1,197,900

1,113,200

146,700

1,259,9001,113,200

58,700

1,171,900

1,172,400

68,100

1,240,5001,172,400

27,200

1,199,600

1,098,100

76,900

1,175,0001,098,100

30,800

1,128,900

Table 58. TEWI for low and medium temperature distributed refrigeration systems in Japan
193

APPENDIX J: AUTOMOBILE AIR-CONDITIONING REFRIGERANT EMISSION

SCENARIOS
194
0 2 4 6 8 10

0

200

400

600

800

1,000

1,200

Time (years)

System Charge (g)

0 5 10 15 20

0

200

400

600

800

1,000

1,200

Time (years)

System Charge (g)

195

Figure 48. Air-conditioner charge with “as manufactured”

emission rate of 35 g/year.

Figure 49. Air-conditioner charge with 11⁄2 recharges during

operational lifetime (3 recharges in 22 years).

Three scenarios are considered for

HFC-134a emissions from automobile air

conditioners. The first uses an estimate for

leakage from hoses, gaskets, and seals

based on manufacturers data for the design

and operation of new systems (Zietlow

1997, Hara 1996, Fernqvist 1997). This

estimate of 35 g/y (1.23 oz) represents a

lower limit on HFC-134a emissions and it

is very unlikely that emissions would be

lower than this. The second and third

scenarios are based on 11⁄2 and 2 service

visits during the operating lifetime of the air

conditioner requiring refrigerant additions

of 40% of the original equipment charge (Baker 1996). These three scenarios are shown in Figs. 48 to

50. The horizontal axis in Fig. 48 has been extended to 22 years to show that three recharges would

indeed be required under this scenario in twice the assumed system lifetime.

The lifetime system refrigerant emissions under these assumptions are summarized in Table 59.

This shows lifetime losses of 415 g (14.6 oz) under the minimal emissions scenario, 860 g (30.3 oz)

assuming 11⁄2 recharges, and 1080 g (38.1 oz) assuming 2 recharges. 
0 5 10 15 20

0

200

400

600

800

1,000

1,200

Time (years)

System Charge (g)

196

Refrigerant Requirement “As Manufactured” 11⁄2 Recharges 2 Recharges

Original Equipment Charge 1000 g 1000 g 1000 g 

Service Additions 0 g 400 g 800 g 

Total Refrigerant Usage 1000 g 1400 g 1800 g 

End-of-Life Charge 650 g 600 g 800 g 

End-of-Life Refrigerant

Recovery 585 g 540 g 720 g 

Net Lifetime Refrigerant

Usage: Emissions (Total

Usage - Recovery)

415 g 860 g 1080 g 

Table 59. Refrigerant emissions for automobile air conditioning (North

America).

Figure 50. Air-conditioner charge with two recharges during

operational lifetime (4 recharges in 22 years).
197

APPENDIX K: AMBIENT TEMPERATURE DISTRIBUTIONS
198
0 3 6 9 11 14 17 20 22 25 28 31 34 36 39

0

200

400

600

800

1,000

1,200

1,400

Temperature (°C)

Temperature Incidence (h/y)

6 9 11 14 17 20 22 25 28 31 34

0

500

1,000

1,500

2,000

Temperature (°C)

Temperature Incidence (h/y)

199

Figure 51. Outdoor ambient air temperature distribution for

Greece (average).

Figure 52. Outdoor ambient temperature distribution for Okinawa,

Japan.
200

-2.5 3.1 8.6 14.2 19.7 25.3 30.8 36.4 41.9 47.5

0

200

400

600

800

1,000

Temperature (°C)

Temperature Incidence (h/y)

Figure 53. Outdoor ambient temperature distribution for the

southwestern U.S.
201

APPENDIX L: TEWI FOR AUTOMOBILE AIR CONDITIONING
202
203

Air-Conditioning System Midwest Southeast Northeast Southwest

HFC-134a: as manufactured

Direct Effect

Weight

Air Conditioner Energy

TEWI

540

363

1088

1991

540

363

3093

3996 

540

363

1405

2308

540

363

2889

3792

HFC-134a: 11⁄2 recharges

Direct Effect

Weight

Air Conditioner Energy

TEWI

1118

363

1088

2569

1118

363

3093

4574

1118

363

1405

2886

1118

363

2889

4370

HFC-134a: 2 recharges

Direct Effect

Weight

Air Conditioner Energy

TEWI

1404

363

1088

2855

1404

363

3093

4860

1404

363

1405

3172

1404

363

2889

4656

Hydrocarbon

Direct Effect

Weight

Air Conditioner Energy

TEWI

4

491

1598

2093

4

491

4482

4977

4

491

2062

2557

4

491

4160

4655

CO2: prototype

Direct Effect

Weight

Air Conditioner Energy

TEWI

1

491

1339

1831

1

491

3798

4290

1

491

1730

2221

1

491

3759

4251

CO2: equivalent air-side

Direct Effect

Weight

Air Conditioner Energy

TEWI

1

491

1773

2264

1

491

5035

5526

1

491

2291

2782

1

491

4902

5393

Table 60.. TEWI for automobile air conditioners in the U.S.
204

Air-Conditioning System

United

Kingdom Germany Greece Italy Spain

HFC-134a: as manufactured

Direct Effect

Weight

Air Conditioner Energy

TEWI

540

245

454

1239

540

245

480

1265

540

245

1255

2040

540

245

927 

1712

540

245

974

1759

HFC-134a: 11⁄2 recharges

Direct Effect

Weight

Air Conditioner Energy

TEWI

1118

245

454

1817

1118

245

480

1843

1118

245

1255

2618

1118

245

927 

2290

1118

245

974 

2337

HFC-134a: 2 recharges

Direct Effect

Weight

Air Conditioner Energy

TEWI

1404 

245

454

2103

1404

245

480

2129

1404

245

1255

2904

1404

245

927 

2576

1404

245

974 

2623

Hydrocarbon

Direct Effect

Weight

Air Conditioner Energy

TEWI

4

331

681

1016

4

331

715 

1050

4

331

1838

2173

4

331

1362

1697

4

331

1431

1766

CO2: prototype

Direct Effect

Weight

Air Conditioner Energy

TEWI

1

331

576

908

1

331

601

933

1

331

1547

1879

1

331

1152

1484

1

331

1210

1542

CO2:equivalent air-side

Direct Effect

Weight

Air Conditioner Energy

TEWI

1

331

757

1089

1

331

792

1124

1

331

2048

2380

1

331

1521

1853

1

331

1598

1930

Table 61. TEWI for automobile air conditioners in Europe.
205

Air-Conditioning System Itazuke Okinawa Misawa Yokota

HFC-134a: as manufactured

Direct Effect

Weight

Air Conditioner Energy

TEWI

507 

154 

610 

1271 

507 

154 

972 

1633 

507 

154 

355 

1016 

507 

154 

521 

1182 

HFC-134a: 11⁄2 recharges

Direct Effect

Weight

Air Conditioner Energy

TEWI

783 

154 

610 

1547 

783 

154 

972 

1909 

783 

154 

355 

1292 

783 

154 

521 

1458 

HFC-134a: 2 recharges

Direct Effect

Weight

Air Conditioner Energy

TEWI

983 

154 

610 

1747 

983 

154 

972 

2109 

983 

154 

355 

1492 

983 

154 

521 

1658 

Hydrocarbon

Direct Effect

Weight

Air Conditioner Energy

TEWI

3 

167 

888 

1058 

3 

167 

1408 

1578 

3 

167 

521 

691 

3 

167 

760 

930 

CO2: prototype

Direct Effect

Weight

Air Conditioner Energy

TEWI

0 

167 

719 

886 

0 

167 

1147 

1304 

0 

167 

412 

579 

0 

167 

608 

775 

CO2: equivalent air-side

Direct Effect

Weight

Air Conditioner Energy

TEWI

0

167

1065

1232

0

167

1694

1861

0

167

612

780

0 

167 

902 

1069 

Table 62. TEWI for automobile air conditioners in Japan.
206
207

APPENDIX M: AUTOMOBILE AIR CONDITIONING (EQUATIONS)
208
209

Pcomp(Tamb) '

ÄH(Tamb)

COP(Tamb)

Q0(Tamb,rpm) ' çvol(rpm)@ñ@D@rpm@ÄH(Tamb)

COPsys(Tamb,rpm) '

Q0(Tamb,rpm)

Pcomp%Pblower%Ppump

Time

sys

(Tamb) '

Hours(T amb ,city)

8760 hours/year

@

1210 vehicle hours

50,000 miles

@(miles per year for the region)@(% compressor on&time)

P(rpm) ' jTamb

Q0(Tamb,rpm)@Timesys(Tamb)

COPsys(Tamb,rpm)

1. Compressor power:

2. Energy output (cooling rate):

3. System COP:

4. Hours of air conditioner operation:

5. Energy input (system power consumption):
210
211

APPENDIX N: RESIDENTIAL GAS HEATING/COOLING TEWI RESULTS
212
213

Equipment Lifetime 15 yearsBuilding Description 1800 

74,700,00016,100,000

ft 2Btu/yheatingBtu/ycooling

Carbon Dioxide Conversions 0.0559

0.650.19

kg CO2 / 1000 Btu gas

input

kg CO2/kWh electricity

kg CO2 / 1000 Btu I2R

Equipment Description Refrigerant

Seasonal Efficiency Indirect Effect Direct

Effect

TEWI

Cooling Heating Electricity Gas Parasitics Refrigeran

t

1996 Equipment refrigerant emissions: make-up rate: 4% of charge per year, end-of-life loss: 15% of chargeCentral A/C and Gas Furnace HCFC-22

HCFC-22

SEER - 10SEER - 12

80% flue92% flue

15,69813,081

78,29568,083

00

35703570

97,56284,734

Premium Options High Efficiency A/C and Gas Furnace Gas Engine Driven 

HCFC-22HCFC-22

SEER - 14gCOP1.30

92% fluegCOP1.26

11,213

0

68,08360,080

0

8010

35703570

82,86671,660

2005 Equipment refrigerant emissions: make-up rate: 2% of charge per year, end-of-life loss: 15% of chargePremium Heat Pump Options Gas Engine Driven GAX Absorption Heat Pump

HCFC-22

R-717

gCOP1.30gCOP -1.0

gCOP1.26gCOP- 1.5

00

60,08055,274

8,010

15,578

2142

0

70,23370,851

Table 63. TEWI calculations for unitary space conditioning equipment: Pittsburgh, Pennsylvania USA.
214

Equipment Lifetime 15 yearsBuilding Description 1800 

34,800,00033,800,000

ft 2Btu/y heatingBtu/y cooling

Carbon Dioxide Conversions 0.0559

0.650.19

kg CO2 / 1000 Btu gas input

kg CO2/kWh electricity

kg CO2 / 1000 Btu I2R

Equipment Description Refrigerant

Seasonal Efficiency Indirect Effect Direct

Effect

TEWI Cooling Heating Electricity Gas Parasitics Refrigerant

1996 Equipment refrigerant emissions: make-up rate: 4% of charge per year, end-of-life loss: 15% of chargeCentral A/C and Gas FurnaceCentral A/C and Electric Furnace

HCFC-22HCFC-22HCFC-22

SEER - 10SEER - 12SEER - 12

80% flue92% flue

100%

32,95527,46327,463

36,47531,717

0

00

99,440

357035703570

73,00062,750

130,476

Premium Options High Efficiency A/C and Gas Furnace Gas Engine Driven

HCFC-22HCFC-22

SEER - 14gCOP - 1.30

92% fluegCOP-1.26

23,539

0

31,71744,973

0

7,780

35703570

58,82656,323

2005 Equipment refrigerant emissions: make-up rate: 2% of charge per year, end-of-life loss: 15% of chargePremium Options Gas Engine Driven GAX Absorption Heat Pump

HCFC-22R-717

gCOP - 1.30gCOP - 1.0

gCOP-1.26gCOP- 1.5

00

44,97347,809

7,780

13,668

2142

0

54,89561,476

Table 64. TEWI calculations for unitary space conditioning equipment: Atlanta, Georgia USA.
215

EquipmentSpecification: Lifetime Heat pump capacity Annual heating load

1510

16,400

yearskWkW/year

Refrigerant Emission Data: Make-up rate End-of-Life Loss

5% of charge peryear15% of charge

CO2 Conversion Factors: Electricity Generation Natural Gas Combustion

0.4690.190

kg/kWhkg/kWh

Technology

Efficiency

(SPF)

Electricity Natural Gas Refrigerant Emissions

TEWI(kg CO2)

Heating(kWh/y

)

Auxiliary

(kWh/y)

Lifetime

CO2(kg)

Heating(kWh/y

)

Lifetime

CO2(kg)

Charge

(kg)

Leakage

(g/y)

End-ofLife

(g)

Lifetime

Loss

(kg)

Refrigeran

t

GWP

(kg

CO2/kg)

DirectEffect

(kg

CO2)

Electric ResistanceHeat

1.00 16,400 0 115,374 0 0 0 0 0 0 0 0 115,374

Gas Boiler or Furnace 0.80

0.92

00

00

00

42033654

58,54850,911

00

00

00

00

00

00

58,54850,911

Thermal Sorption 1.27 0 895 6299 2647 36881 0 0 0 0 0 0 43,180

Table 65. TEWI for European residential space heating (data from the International Heat Pump Status and Policy Review Report,PHC-AR3, IEA Heat Pump Centre, September 1994, Part 1.