INVESTIGATION OF 

SECONDARY LOOP SUPERMARKET 

REFRIGERATION SYSTEMS

Prepared For: 

California Energy Commission 

Public Interest Energy Research Program 

Prepared By: 

Southern California Edison 

Foster-Miller, Inc. 

CONSULTANT 

REPORT

March 2004 

500-04-013
Prepared By: 

Southern California Edison, RTTC 

P.O. Box 6400 

Rancho Cucamonga, CA 91729 

Ramin T. Faramarzi, P.E. 

Principal Investigator 

Foster-Miller, Inc. 

350 Second Avenue 

Waltham, MA 02451 

David H. Walker, P.E. 

Principal Investigator 

Contract No. 500-98-039 

Prepared For: 

California Energy Commission 

Chris Scruton, 

Project Manager 

Nancy Jenkins, 

Program Area Lead 

Buildings End-Use Energy Efficiency

Terry Surles, 

PIER Program Manager 

Robert L. Therkelsen, 

Executive Director

DISCLAIMER 

This report was prepared as the result of work sponsored by the 

California Energy Commission. It does not necessarily represent 

the views of the Energy Commission, its employees or the State 

of California. The Energy Commission, the State of California, its 

employees, contractors and subcontractors make no warrant, 

express or implied, and assume no legal liability for the 

information in this report; nor does any party represent that the 

uses of this information will not infringe upon privately owned 

rights. This report has not been approved or disapproved by the 

California Energy Commission nor has the California Energy

Commission passed upon the accuracy or adequacy of the 

information in this report. 
i 
Acknowledgements 

The authors would like to thank Chris Scruton and Nancy Jenkins of the California 

Energy Commission for their help and guidance throughout this project. We would also 

like to recognize the outstanding work and cooperation that we received from personnel 

at several organizations: 

Safeway, Inc. Rob Uhl 

Hill PHOENIX, Inc Henry Pellerin 

Scott Martin 

Georgi Kazachki 

David Hinde 

Yakov Arshansky 

Ken Mongrain 

Southern California Edison’s RTTC Bruce Coburn 

Rafik Sarhadian 

Authors: 

Ramin T. Faramarzi, P.E. - Southern California Edison, RTTC 

David H. Walker, P.E. - Foster-Miller, Inc. 

ii 
iii 
Preface 

The Public Interest Energy Research (PIER) Program supports public interest energy 

research and development that will help improve the quality of life in California by 

bringing environmentally safe, affordable, and reliable energy services and products to 

the marketplace. 

The PIER Program, managed by the California Energy Commission (Commission), 

annually awards up to $62 million to conduct the most promising public interest energy 

research by partnering with Research, Development, and Demonstration (RD&D) 

organizations, including individuals, businesses, utilities, and public or private research 

institutions. 

PIER funding efforts are focused on the following six RD&D program areas: 

• Buildings End-Use Energy Efficiency 

• Industrial/Agricultural/Water End-Use Energy Efficiency 

• Renewable Energy 

• Environmentally-Preferred Advanced Generation 

• Energy-Related Environmental Research 

• Strategic Energy Research. 

What follows is the final report for the Investigation of Secondary Loop Supermarket 

Refrigeration Systems, Contract No. 500-98-039, conducted by Southern California 

Edison RTTC. The report is entitled Investigation of Secondary Loop Supermarket 

Refrigeration Systems. This project contributes to the Buildings End-Use Energy 

Efficiency program. 

For more information on the PIER Program, please visit the Commission's Web site at: 

http://www.energy.ca.gov/research/index.html or contact the Commission's 

Publications Unit at 916-654-5200. 

iv 
v 
Table of Contents 

Preface ............................................................................................................................................ iv

Executive Summary ........................................................................................................................ 1

Abstract ........................................................................................................................................... 5

1.0 Introduction ......................................................................................................................... 7

1.1 Present Supermarket Refrigeration Systems ................................................................... 7

1.2 Low-charge Refrigeration System Options................................................................... 10

1.3 Project Objectives.......................................................................................................... 12

1.4 Project Organization and Approach .............................................................................. 12

1.5 Report Organization ...................................................................................................... 13

2.0 Project Approach............................................................................................................... 15

2.1 Design of the Secondary Loop Refrigeration System ................................................... 15

2.1.1 Description of the Secondary Loop Refrigeration System.................................... 15

2.1.2 High-Efficiency Features for Secondary Loop Refrigeration ............................... 17

2.1.3 Display Cases Designed for Use with Secondary Fluid ........................................ 18

2.1.4 High-efficiency Refrigeration Compressors.......................................................... 19

2.1.5 Close-Coupling of Compressors and Evaporator.................................................. 19

2.1.6 Multiple Parallel Pumps........................................................................................ 19

2.1.7 Evaporative Heat Rejection................................................................................... 20

2.1.8 Low Viscosity Secondary Fluid ............................................................................ 20

2.1.9 Refrigerant Subcooling from Warm Brine Defrost ............................................... 22

2.2 Analysis of the Secondary Loop and Multiplex Refrigeration Systems........................ 23

2.2.1 Multiplex Refrigeration Model.............................................................................. 23

2.2.2 Modeling of Secondary Loop Refrigeration.......................................................... 27

2.3 Field Testing of the Secondary Loop Refrigeration System ......................................... 30

2.3.1 Description of the Secondary Loop Refrigeration Test Site.................................. 30

2.3.2 Measurement Plan to Monitor Secondary Loop Refrigeration.............................. 35

2.3.2.1 Refrigeration Energy Consumption................................................................... 35

2.3.2.2 Refrigeration Supplied....................................................................................... 36

2.3.2.3 Refrigeration System Operating State Points.................................................... 37

2.3.2.4 Brine Pump Operation....................................................................................... 37

2.3.2.5 Brine Defrost Operation .................................................................................... 38

2.3.2.6 Display Case Operation..................................................................................... 38

2.3.2.7 Ambient Conditions........................................................................................... 38

2.3.2.8 Data Collection at the Secondary Loop Store.................................................... 39

2.3.3 Field Testing of the Multiplex Refrigeration System............................................ 39

2.3.3.1 Description of the Multiplex Refrigeration Test Site ........................................ 39

2.3.4 Measurement Plan for the Multiplex Refrigeration System.................................. 43

2.3.4.1 Refrigeration Energy Consumption................................................................... 43

2.3.4.2 Refrigeration System Operating State Points.................................................... 43

2.3.4.3 Display Case Operation..................................................................................... 44

2.3.4.4 Ambient Conditions........................................................................................... 44

2.3.5 Data Collection at the Multiplex Refrigeration Test Store.................................... 44

3.0 Project Outcomes............................................................................................................... 45

vi 
3.1 Analysis Results............................................................................................................ 45

3.2 Field Test Results.......................................................................................................... 46

3.2.1 Energy Comparison ............................................................................................... 46

3.2.2 Power Demand ...................................................................................................... 50

3.2.3 Refrigeration Efficiency ........................................................................................ 52

3.2.4 Product Storage Temperatures............................................................................... 57

3.2.5 Refrigerant Leakage Comparison.......................................................................... 60

4.0 Conclusion & Recommendations...................................................................................... 63

5.0 References......................................................................................................................... 65

List of Tables 

Table 1. Comparison of Operating Temperatures for Direct-Expansion and Secondary-Fluid 

Refrigerated Display Cases ................................................................................................... 19

Table 2. Transport Properties of Propylene Glycol-Water and Organic Salt-Water (Dynalene) at 

Operating Temperatures of Secondary Loop Refrigeration .................................................. 21

Table 3. Low Temperature Circuits for the Secondary Loop Refrigeration Test Store ................ 31

Table 4. Medium Temperature Circuits for the Secondary Loop Refrigeration Store (Loop B1) 32

Table 5. Medium Temperature Circuits for the Secondary Loop Refrigeration Store .................. 33

Table 6. Low Temperature Circuits at the Multiplex Refrigeration Store (Rack A)..................... 40

Table 7. Medium Temperature Circuits at the Multiplex Refrigeration Store (Rack B).............. 40

Table 8. Medium Temperature Circuits at the Multiplex Refrigeration Store (Rack C).............. 41

Table 9. Medium Temperature Circuits at the Multiplex Refrigeration Store (Rack D).............. 41

Table 10. Secondary Loop Modeling Results ............................................................................... 45

Table 11. Multiplex Modeling Results – Air-Cooled Condensing................................................ 45

Table 12. Multiplex Modeling Results – Evaporative Condensing............................................... 46

Table 13. Multiplex Refrigeration Average Daily Energy Consumption...................................... 47

Table 14. Secondary Loop Refrigeration Average Daily Energy Consumption ........................... 47

Table 15. Performance of Multiple Pump Arrangement for Secondary Fluid Pumping............... 50

Table 16. Peak Power Demand of the Multiplex Store ................................................................. 51

Table 17. Peak Power Demand for the Secondary Loop Store ..................................................... 51

Table 18. Multiplex Refrigeration................................................................................................. 53

Table 19. Secondary Loop Refrigeration ...................................................................................... 53

Table 20. Comparison of Secondary Loop and Multiplex Refrigeration Using Normalized EER 

Values, Multiplex Store Refrigeration Loads and Ambient Wet-bulb Temperatures........... 54

Table 21. Comparison of Secondary Loop and Multiplex Refrigeration Using Normalized EER 

Values, Secondary Loop Store Refrigeration Loads and Ambient Wet-bulb Temperatures. 55

Table 22. Comparison of Display Case Performance for the Multiplex and Secondary Loop 

Refrigeration Systems ........................................................................................................... 57

Table 23. Refrigerant Additions at the Secondary Loop and Multiplex Stores............................. 60

Table 24. Estimated Annual Refrigerant Leak Rates for the Secondary Loop and Multiplex Test 

Stores..................................................................................................................................... 61

vii 
List of Figures 

Figure 1. Layout of Refrigerated Display Cases and Walk-in Coolers in a Typical Supermarket.. 7

Figure 2. The Multiplex System is the Most Common Type of Refrigeration Used in 

Supermarkets........................................................................................................................... 8

Figure 3. Flow Diagram for a Secondary Loop Refrigeration System.......................................... 16

Figure 4. Layout of the Secondary Loop Refrigeration Test Store ............................................... 30

Figure 5. Piping Diagram for the Low Temperature Secondary Loop Refrigeration System

(Patent by Hill PHOENIX, Inc.)............................................................................................ 34

Figure 6. Piping for the Medium Temperature Secondary Loop Refrigeration System (Patent by

Hill PHOENIX, Inc.)............................................................................................................. 35

Figure 7. Layout of the Multiplex Refrigeration Test Store.......................................................... 39

Figure 8. Piping Diagram for the Low Temperature Multiplex Refrigeration .............................. 42

Figure 9. Piping Diagram for the Medium Temperature Multiplex Refrigeration........................ 42

Figure 10. Store and Total Refrigeration Energy use for the Multiplex and Secondary Loop Test 

Sites....................................................................................................................................... 48

Figure 11. Multiplex Refrigeration System Energy Consumption................................................ 49

Figure 12. Secondary Loop Refrigeration Energy Consumption .................................................. 49

Figure 13. Peak Power Demand for the Multiplex and Secondary Loop Stores........................... 51

Figure 14. Relation between Ambient Wet-Bulb Temperature and Low Pressure Refrigeration 

EER ....................................................................................................................................... 55

Figure 15. Relation between Ambient Wet-Bulb Temperature and Medium Temperature 

Refrigeration EER for the Multiplex System ........................................................................ 56

Figure 16. Relation between Ambient Wet-Bulb Temperature and Medium Temperature 

Refrigeration EER for the Secondary Loop System.............................................................. 56

Figure 17. Discharge Air and Product Temperature Profiles for the Single-Deck Meat Case - 

Multiplex Refrigeration......................................................................................................... 58

Figure 18. Discharge Air and Product Temperature Profiles for the Single-Deck Meat Case - 

Secondary Loop Refrigeration .............................................................................................. 58

Figure 19. Discharge Air and Product Temperature Profiles for the Multi-Deck Product Case – 

Multiplex Refrigeration......................................................................................................... 59

Figure 20. Discharge Air and Product Temperature Profiles for the Multi-Deck Product Case – 

Secondary Loop Refrigeration .............................................................................................. 59

viii 
ix 
Executive Summary 

Background 

The largest single use of energy in a supermarket is refrigeration, accounting for half of 

total energy use, on the order of 2-3 million kWh annually for a 35,000 ft2

 store. 

Refrigeration is typically provided by direct expansion air-refrigerant coils that are 

located in the display cases and walk-in coolers. The compressors are located in a 

machine room in a remote part of the store, such as in the back room area or on the roof. 

As a result of using this layout, the amount of refrigerant needed to charge a 

supermarket refrigeration system is very large. A typical store will require 3,000- 5,000 

lb of refrigerant. The large amount of piping and fittings used in supermarket 

refrigeration leads to significant refrigerant leakage. On average, a supermarket 

refrigeration system can be expected to lose 30 – 50% of the total charge annually. 

With increased concern about the impact of refrigerant leakage on global warming, new 

supermarket system types requiring significantly less refrigerant charge are now being 

considered. Of these advanced systems, the secondary loop refrigeration system 

employs central mechanical systems and can reduce the total refrigerant charge to 

approximately 300 – 500 lb. The anticipated refrigerant leak rate for the secondary loop 

refrigeration system is much less than that of the direct-expansion systems, because all 

refrigerant is contained within the chiller system. Refrigeration is provided to display 

cases by a secondary fluid pumped between the cases and the chiller system. Secondary 

loop systems, however, have tended to use more energy than multiplex systems. 

The overall goal of this project was to determine if an advanced, high-efficiency 

secondary loop refrigeration system could have a lower operating cost than a state-ofthe-art multiplex refrigeration system, considering both energy and refrigerant use. This 

would make the environmentally benign secondary loop system economically attractive. 

Project Approach 

Southern California Edison and its subcontractor, Foster-Miller, Inc., obtained support 

from Safeway, Inc. to provide a supermarket where an advanced secondary loop 

refrigeration system could be installed and tested. After design and analysis work was 

performed, a specification for the advanced secondary loop system was prepared and 

several refrigeration manufacturers were asked to propose and bid on its construction. 

Hill PHOENIX, Inc. of Conyers, GA was selected as the system manufacturer and 

installing contractor based on their response to this solicitation. 

The advanced secondary loop refrigeration system was installed and instrumented for 

evaluation of its performance. At that same time, a second supermarket operated by 

Safeway was identified that employed a state-of-the-art multiplex refrigeration system. 

The multiplex system was also instrumented for performance measurement and both 

sites were monitored for approximately 9 months. 

1 
Project Objectives and Results 

The specific technical objectives of the project were to build and test a secondary loop 

refrigeration system, which: 

• Consumes approximately 14% less electricity than a state-of-the-art multiplex refrigeration 

system (baseline system) installed in a comparable store. 

The modeling results for the secondary loop refrigeration system designed for this 

project predicted annual energy savings of 145,894 kWh, or 14.9%, when compared with 

a state-of-the-art multiplex refrigeration system with air-cooled condensing. The 

multiplex system with an air cooled condenser is the most common configuration found 

in supermarkets. 

In the field test, the secondary loop system was compared to a multiplex system with 

evaporative condensing, which is more efficient than air cooled condensing. This was 

appropriate because the secondary loop system also used evaporative condensing. 

Modeling results predicted annual savings of 6,130 kWh, or just under 1% for this 

comparison, but the actual savings achieved by the secondary loop refrigeration system 

were 37,266 kWh/yr, or 4.9%. This suggests the benefit of the secondary loop system as 

tested over an air cooled multiplex system would have been greater than 14.9%. 

The savings achieved by the secondary loop refrigeration system can be attributed to 

energy-saving features incorporated in its design. The annual savings achieved by each 

of these features compared to a more conventional secondary loop refrigeration system 

were: 

• Multiple, parallel brine pumps – Estimated annual savings versus single large 

pumps are 99,718 kWh. 

• Subcooling from warm brine defrost – Estimated annual savings versus no 

subcooling are 49,570 kWh. 

Additional savings were also achieved by the use of a minimum difference between the 

display case discharge air and refrigeration saturated suction temperatures, the use of a 

low-viscosity secondary fluid, and evaporative condensing as noted previously. 

• Has a refrigerant charge that is ten times less (less than 500 lbs.) than the baseline system. 

The advanced secondary loop refrigeration system as tested had an initial refrigerant 

charge of 1,400 lb., which is considerably larger than the stated goal. The reason for the 

larger refrigerant charge is that the secondary loop system also had heat reclaim 

capabilities for both hot water and space heating. The added charge was needed for 

piping and heat exchangers associated with heat reclaim. Heat reclaim is of great value 

to the operation of the supermarket since it is capable of displacing all energy use 

associated with hot water heating and space heating for the store (Estimated to be 

approximately 7% of the total store energy use). Without the heat reclaim equipment 

the refrigerant charge would have been approximately 500 lbs. Further design 

2 
enhancements such as including the heat recovery equipment within the factory-made 

chiller equipment room could result in a reduced overall refrigerant charge. 

 

• Loses annually no more than 15% of its refrigerant charge due to leakage. 

Service records allowed refrigerant leak rates for the 2 stores to be calculated, using the 

methods provided by the USEPA. These calculations showed that the refrigerant leak 

rate for the secondary loop store was on the order of 14.8%/yr, or 206 lb/yr. The 

multiplex system was estimated to have a leak rate of only 12.4%/yr, but because of its 

much larger charge this amounted to 370 lb/yr. The refrigerant data for the multiplex

store were very limited stemming from a change in maintenance contractors. It is likely 

that a full year of data would show a larger loss of refrigerant for the multiplex 

refrigeration system. Over time, and as equipment is serviced, the more scattered 

refrigeration apparatus of the multiplex system will probably begin to leak more, which 

would increase the relative benefit of the more centralized secondary loop system. 

The ability of the 2 refrigeration systems to maintain product storage temperature was 

also assessed. The comparison for single-deck meat cases showed that the multiplex and 

secondary loop systems maintained the product at acceptable temperature levels. The 

case associated with the multiplex system had a lower and more uniform product 

temperature than was seen for the case operating in the secondary loop system. The 

multiplex display case had to operate a much lower rack SST (2.3°F for multiplex vs. 

14.1°F for secondary loop) in order to achieve this condition. For the multi-deck 

produce cases, the multiplex and secondary loop systems operated at similar rack SST 

values. The resulting average product temperature was approximately the same for 

both systems, but the product temperature of the multi-deck case in the secondary loop 

system was more uniform in value. 

Conclusion 

Based on this research, we can conclude that secondary loop refrigeration systems are a 

viable option for supermarket refrigeration because: 

• With efficient design the energy consumption of a secondary loop system can be 

less than a multiplex system. The project system consumed 4.9% less electricity 

than the baseline multiplex system. 

• Refrigerant use is reduced in proportion to the quantities of refrigerant, piping 

and fittings employed. Refrigerant leaks are difficult to find and repair, and after 

years of repairs and modifications the multiplex system will probably leak more 

than the compact and centralized secondary loop system. 

• The somewhat higher first cost for the secondary loop system will probably be 

mitigated over time by lower operating and maintenance costs. 

Benefits to California 

This project contributed to the PIER program objective of reducing the environmental 

costs of California’s electrical system, by developing an alternative refrigeration system 

which uses significantly less refrigerant than conventional systems. It also contributed 

3 
to the PIER program objective of improving energy value of California’s electricity by 

lowering electrical consumption of supermarket secondary loop refrigeration systems. 

4 
Abstract 

Present multiplex supermarket refrigeration systems consume approximately 1 to 1.5 

million kWh/yr. and may lose annually as much as half of the 3,000 to 5,000 lb system 

refrigerant charge. The secondary loop refrigeration system employs fluid loops and a 

central chiller to provide refrigeration to the supermarket display cases requiring just 

300 to 500 lb of refrigerant for operation. An advanced secondary loop system was 

investigated, taking advantage of energy-saving features to reduce energy consumption. 

Modeling of the advanced system indicated energy savings of 14.9 and 0.3% versus 

multiplex systems with air-cooled and evaporative condensing, respectively. Results of 

a field test comparison of the secondary loop system, a multiplex refrigeration system 

with evaporative condensing developed energy savings of 4.9% for the secondary loop 

refrigeration system. Savings produced by the secondary loop refrigeration can be 

attributed to: minimum temperature differences between display case discharge air and 

saturated suction temperatures; the use of multiple parallel brine pumps; low-viscosity 

secondary fluid; and subcooling produced by the warm fluid defrost system. An 

evaluation of product temperature for similar display cases at each site was conducted. 

The secondary loop refrigeration system maintained the same or lower product 

temperatures than the product temperatures achieved by the multiplex refrigeration 

system. 

5 
6 
1.0 Introduction 

1.1 Present Supermarket Refrigeration Systems 

One of the largest uses of energy in supermarkets is for refrigeration. Most of the 

product sold is perishable and must be kept refrigerated during storage and display. 

Typical energy consumption for supermarket refrigeration is on the order of half of the 

store’s total (Total store energy use is on the order of 2-3 million kWh annually for a 

35,000 ft2

 store). Compressors and condensers account for 30-35%. The remainder is 

consumed by the display and storage cooler fans, display case lighting, and for antisweat heaters used to prevent condensate from forming on doors and outside surfaces of 

display cases. 

Figure 1 shows the layout of a typical US supermarket refrigeration system. Direct 

expansion air/refrigerant coils, located in the display cases and walk in coolers, provide 

refrigeration. The compressors are located in a machine room in a remote part of the 

store, either in the back room area or on the roof. The system condensers are located 

either in the machine room, or more likely, on the roof above the machine room. Piping 

is provided for refrigerant supply and return between the machine room and the 

refrigerated fixtures. 

Figure 1. Layout of Refrigerated Display Cases and Walk-in Coolers in a Typical 

Supermarket 

7 
Figure 2 is a diagram of the most commonly used compressor arrangement in 

supermarkets, which is multiplex refrigeration. The term “multiplex” refers to the use 

of multiple compressors piped to common suction and discharge manifolds, all installed 

on a skid. The skid contains all necessary piping, control valves, and electrical wiring to 

control the compressors and the refrigeration provided to the display cases and walk-in 

coolers. The hot discharge gas from the compressors is piped to a remotely located 

condenser, which condenses the gas to a liquid. Liquid refrigerant returning from the 

condenser is piped back to the compressor rack, where a receiver, liquid manifold and 

associated control valves are located, distributing the liquid to the cases and coolers. 

Each case or cooler circuit is piped with liquid and suction return lines connected to the 

liquid and suction manifolds located on the compressor skid. Valves used for control of 

these circuits are located on the manifolds. The control valves employed consist of 

regulators to control suction pressure and solenoid valves used to control gas routing 

during defrost. The compressor rack is normally equipped with a number of pressure 

regulators used to control system head pressure, heat reclaim, and defrost. The rack will 

contain an oil separator in the discharge piping and an oil distribution system to return 

oil to the compressors. 

Figure 2. The Multiplex System is the Most Common Type of Refrigeration Used in 

Supermarkets 

Typically, three to five compressor racks are employed to provide all refrigeration in the 

supermarket. The display cases and coolers are grouped, and attached to the 

compressor racks based on required saturated suction temperature (SST) to maintain the 

desired case temperature. A supermarket will have 1 or 2 low temperature racks to 

address all frozen food refrigeration requirements. The low temperature racks will 

typically operate at a –20°F SST. Refrigeration loads as low as –30°F, or as high as –10°F, 

will also be provided by the low temperature racks. In these situations, the suction 

8 
manifold will be divided, and one or two compressors will provide the off-temperature 

refrigeration. The discharge of these “satellite” compressors will be piped to the 

common discharge manifold with the other low temperature loads so that a common 

condenser and liquid manifold can be used for all circuits on the rack. The remaining 

refrigeration circuits in the store are referred to as medium temperature and normally 

require a 20°F SST. Two or more compressor racks are needed to meet all medium 

temperature refrigeration requirements. Satellite compressors are also used for medium 

temperature loads requiring an SST significantly higher or lower than 20°F. 

Multiplex systems commonly consist of three or four compressors that are sized such 

that operation of all compressors simultaneously can provide adequate capacity to meet 

the design refrigeration load. During off-design operation, the refrigeration load can be 

considerably less than the design value; at the same time, the refrigeration capacity of 

the compressors can increase with a decrease in ambient temperature and the condenser 

operating pressure. In this situation, the compressors can be selected so that the 

combined capacities of the compressors closely match the refrigeration load. The 

selection of, and the on-off cycling of the compressors is done based on suction pressure 

value measured at the compressor rack. The use of microprocessor-based controls 

allows more sophisticated control algorithms to be employed, so that very close 

matching of the suction pressure and the set point value can be maintained with 

multiplex compressor systems. 

The most common type of condenser used in supermarket refrigeration is air-cooled. 

The reason is air-cooled condensers require the least maintenance and have been shown 

to operate reliably in the non-operator environment of supermarket refrigeration. Aircooled condensers employ finned coil construction with 8-10 fins/inch and multiple 

fans. On/off fan cycling is used as a means to control condensing temperature and to 

reduce fan energy at ambient temperatures below design. 

Evaporative condensers are used in some supermarkets, primarily in states like 

California with a drier climate, where a substantial difference in dry-bulb and wet-bulb 

temperatures exists. The evaporative condenser consists of a tube bundle, a fan for air 

flow, and a water sump and pump system used to spray water over the tube bundle. 

The refrigerant vapor is passed through the tube bundle where heat is removed and the 

refrigerant is condensed. The resulting condensing temperature can be close to the 

ambient wet-bulb. Evaporative condensers require less air flow than air-cooled 

condensers of equivalent rejection capability and can, therefore, be operated at a lower 

minimum condenser temperature without a fan energy penalty. 

Water treatment and consumption are major issues that prevent more use of evaporative 

condensers in supermarkets. Water treatment is needed because of the evaporation of 

the water, which concentrates dissolved minerals. The minerals will precipitate and 

form deposits on tube surfaces. Exposure of the water to air causes biological growth 

within the evaporative condensers in the form of algae and slime. Treatment of 

evaporative condenser water consists primarily of “blow down” in which a fraction of 

the water is discharged to drain and replaced with fresh water. The amount of blow 

down is controlled by a conductivity controller. The discharged water will carry away 

excess minerals and solids, which prevents precipitation. Biocides, such as chlorine, are 

added to the water by automatic feeders to prevent biological growth. 

9 
The use of evaporative condensers is strongly recommended when energy-efficiency is 

an issue, because of the lower average condensing temperatures achieved with this type 

of heat rejection. 

1.2 Low-charge Refrigeration System Options 

As a result of using the layout described above in supermarkets, the amount of 

refrigerant needed to charge a supermarket refrigeration system is very large. A typical 

store will require 3,000 to 5,000 lb of refrigerant. The amount of piping and pipe joints 

present in supermarket refrigeration systems leads to significant refrigerant leakage. 

This leakage can amount to a loss of up to 30 to 50% of the total charge annually (1). 

With increased concern about the impact of refrigerant leakage on global warming, new 

supermarket system configurations requiring significantly less refrigerant charge are 

now being considered. Advanced systems of this type include: 

• Low-charge multiplex - Several refrigeration system manufacturers now offer 

control systems for condensers that limit the amount of refrigerant charge needed for 

the operation of multiplex refrigeration resulting in reducing the charge by 

approximately 1/3. 

• Distributed – Multiple scroll compressors are located in cabinets placed on or near 

the sales floor. Scroll compressors are employed to minimize system noise in the 

sales area. The cabinets are close-coupled to the display cases and heat rejection 

from the cabinets can be done through the use of a glycol loop that connects the 

cabinets to a fluid cooler, in order to minimize refrigerant charge. 

• Advanced self-contained – Self-contained refrigeration consists of compressors and 

condensers built into the display cases. An advanced version of this concept would 

use horizontal scroll compressors with capacity control, such as unloading, and 

water-cooled condensers with a water loop for heat rejection. The advanced selfcontained refrigeration system would employ the smallest refrigerant charge. 

• Secondary loop – A secondary fluid loop is run between the display cases and a 

central chiller system. The secondary fluid is refrigerated at the chiller and is then 

circulated through coils in the display cases where it is used to chill the air in the 

case. 

While all of these systems use less refrigerant, energy use varies and can be greater than 

the energy usage with centralized multiplex racks. This is particularly true for the 

advanced self-contained system where only small, inefficient compressors are available 

for use in individual display cases. Both the distributed and self-contained systems 

involve the use of compressors and other mechanical hardware on or near the sales area, 

which is an arrangement not favored by supermarket operators. 

In contrast, the secondary loop refrigeration system employs central mechanical systems 

and reduces the total refrigerant charge to approximately 300 to 500 lb. The anticipated 

refrigerant leak rate for the secondary loop refrigeration system is much less than that of 

the direct-expansion systems, because all refrigerant is contained within the chiller 

system. 

10 
Another significant benefit derived from the operation of secondary loop refrigeration is 

improved product temperature of stored perishables. Information provided in (2) 

suggests that with properly designed coils, display cases using secondary fluid for 

refrigeration produce a more uniform air temperature throughout the case than seen in 

direct-expansion display cases. When warm brine is employed for display case defrost, 

the time needed for defrost is significantly shorter, which prevents warming of the 

product. The resulting product temperatures are lower and more consistent than the 

product temperature recorded in direct expansion cases operating at the same air 

temperature. 

Secondary loop refrigeration has been used extensively in Europe, particularly for 

medium temperature applications (3). These present systems, in general, have used 

more energy than comparable multiplex refrigeration. The reasons for the increased 

energy use include: 

• The use of display cases with evaporators designed for direct-expansion operation –

heat transfer characteristics of these coils are not suited for the use of the secondary 

fluid. The coil must be operated at a lower temperature to achieve the desired 

refrigeration load and case temperature. 

• Operation of the refrigeration system at lower saturated suction temperature –

caused by the added temperature difference of the secondary fluid that must be 

provided to maintain the required refrigeration temperature. The lower saturated 

suction temperature reduces the efficiency of the refrigeration compressors, 

increasing refrigeration energy use. 

• Pumping energy associated with the circulation of the secondary fluid. Many of the 

secondary loop systems employ propylene glycol and water as the secondary fluid, 

because it is non-toxic and acceptable for use near foods. At low temperature, the 

viscosity of the glycol-water mixture is much higher than that of water alone, greatly 

increasing the fluid friction and pumping power. 

Presently, the installed cost of the secondary loop refrigeration system may be higher 

than that of multiplex refrigeration. The reason for this increased cost is the multiplex

system is established and there are many qualified contractors, which leads to cost 

competition. It is very likely that increasing the number of installations of secondary 

loop refrigeration could also cause the first cost of these systems to drop. 

 Supermarket operators can justify the higher first cost of the secondary loop system, if 

the benefits derived from its use provide an acceptable payback and return on 

investment. The environmental and food storage benefits help offset the added cost, but 

the added operating cost of the energy penalty reduces the payback potential. 

The purpose of this project is to address the energy consumption of a secondary loop 

refrigeration system and find ways to reduce it. Since the secondary loop system 

employs a chiller system similar to the multiplex system, many of the energy-saving 

features now used with multiplex systems, such as floating head pressure, mechanical 

subcooling, and evaporative heat rejection can be implemented to reduce energy use. 

Certain unique characteristics of the secondary loop system can also be exploited for 

11 
energy savings. These characteristics include: close-coupling between the chiller 

evaporator and the compressor suction which reduces pressure drop and refrigerant gas 

superheating; and liquid refrigerant subcooling by brine warming for defrost. The use 

of a low viscosity fluid and the secondary fluid along with better pump capacity control 

can greatly reduce the energy associated with brine pump. Finally, the use of display 

cases and fixtures with heat transfer surfaces designed specifically for secondary loop 

systems can greatly reduce heat transfer penalties. 

Initial estimates for an advanced secondary loop system employing all of these design 

and energy-saving features suggest that the annual energy use could be significantly less 

than that of the multiplex refrigeration system now employed in many supermarkets. 

1.3 Project Objectives 

The overall goal of this project was to determine if a high-efficiency secondary loop 

refrigeration system could be economically attractive, due to reduced operating costs, 

than a state-of-the-art multiplex refrigeration system, which contains significantly more 

refrigerant. 

This project contributes to the PIER program objective of reducing the environmental 

costs of California’s electrical system, by developing an alternative refrigeration system, 

which uses significantly less refrigerant than conventional refrigeration systems. It also 

contributes to the PIER program objective of improving energy cost/value of 

California’s electricity by lowering electrical consumption of supermarket secondary 

loop refrigeration systems. 

The project identified system improvements to achieve maximum energy efficiency for 

secondary loop refrigeration systems in supermarkets. 

 The specific technical objectives of the project were to design and test a secondary loop 

refrigeration system, which: 

• Consumes approximately 14% less electricity than a state-of-the-art multiplex

refrigeration system (baseline system) installed in a comparable store. 

• Has a refrigerant charge that is one tenth of (less than 500 lbs.) the baseline 

system. 

• Loses annually no more than 15% of its refrigerant charge due to leakage. 

The specific economic objective of the project was to determine if the prototype, highefficiency secondary loop refrigeration system has lower total operating costs 

(operations, maintenance and repair) than the baseline system. 

1.4 Project Organization and Approach 

Southern California Edison RTTC and its subcontractor, Foster-Miller, Inc. obtained 

support from Safeway, Inc. to provide a supermarket where an advanced secondary 

loop refrigeration system could be installed and tested. After design and analysis work 

was performed, a specification for the advanced secondary loop system was prepared 

and several refrigeration manufacturers were asked to propose and bid on the 

12 
construction. Hill PHOENIX, Inc. of Conyers, GA was selected as the system 

manufacturer and installing contractor based on their response to the request for 

proposal. 

The advanced secondary loop refrigeration system was installed and instrumented for 

evaluation of its performance. Concurrently, a second supermarket operated by 

Safeway was identified in which a state-of-the-art multiplex refrigeration system was 

installed. The multiplex system was instrumented for performance measurement and 

both sites were monitored for approximately 9 months. Originally, the test plan called 

for a one-year field test, but problems with the data acquisition system at the multiplex 

store forced the testing to be truncated. 

The measurements were recorded from both sites and compared with respect to the 

following variables: 

• Refrigeration electric energy consumption 

• Power demand – The peak power demand for the refrigeration at each site was 

determined and the impact on the total store demand was developed. 

• Energy Efficiency Ratio (EER) –defined as: 

Power Required

Refrigeration Supplied EER = 

• Product temperature – Product sensors in similar cases were provided at each 

site so the average product temperature and temperature uniformity could be 

compared 

Refrigerant replenishment data for each site were obtained from Safeway so the leak 

rates and annual refrigerant cost could be found and compared. 

1.5 Report Organization 

This report is organized as follows: 

Section 1.0 Introduction 

Section 2.0 Project Approach 

Section3.0 Project Outcomes 

Section 4.0 Conclusions and Recommendations 

Section 5.0 References 

13 
14 
2.0 Project Approach 

2.1 Design of the Secondary Loop Refrigeration System 

2.1.1 Description of the Secondary Loop Refrigeration System 

Figure 3 shows the elements of the secondary loop refrigeration system. The difference 

between secondary loop and direct expansion refrigeration in supermarket applications 

is refrigeration of the display cases and walk-in coolers is provided by a chilled, 

secondary fluid, which is pumped between the refrigeration system and the refrigerated 

fixtures. In direct expansion systems, liquid refrigerant is piped directly to each fixture 

where it is flashed and evaporated to produce cooling. The resulting refrigerant vapor is 

piped back to the central compressor system, completing the refrigeration cycle. 

Secondary loop systems have all of the refrigeration associated with liquid chilling 

located within the central machine room of the supermarket. The chiller system 

resembles the multiplex refrigeration system since multiple parallel compressors are 

employed in both systems. The compressors are utilized based upon suction pressure 

which controls the supply temperature of the chilled fluid. The use of multiple 

compressors allows the refrigeration capacity to conform to changing operating 

conditions, resulting in better fluid temperature control and lower compressor energy 

use. The cooling of the secondary fluid takes place in a heat exchanger whereby the 

liquid refrigerant of the chiller system is evaporated to provide chilling to the secondary 

fluid. Heat rejection for the chiller system is accomplished utilizing a condenser cooled 

by ambient air. Either a dry, air-cooled condenser or an evaporative condenser may be 

employed. The evaporative condenser is favored because of the lower condensing 

temperature which may be obtained by rejecting heat to the lower ambient wet-bulb 

temperature. The lower condensing temperature provides greater compressor 

refrigeration capacity and lower power draw, both of which result in reduced 

refrigeration energy consumption. 

15 
Figure 3. Flow Diagram for a Secondary Loop Refrigeration System 

The secondary loop refrigeration can be configured to operate with two to four separate 

secondary fluid loops and chiller systems. In the two loop configuration, the secondary 

fluid temperatures are 20 and -20°F for medium and low temperature refrigeration. As a 

consequence of using only two temperatures for refrigeration, all display cases and 

storage coolers must operate with these two temperatures. For this and other reasons, 

the heat exchangers in the cases and coolers must be sized differently than those used 

for direct expansion operation. Additional loops may be employed if a secondary fluid 

temperature addresses a large fraction of the refrigeration load and this temperature is 

significantly different from 20 or –20°F. Possible examples of alternate loop 

temperatures are values of –10, 0, or 15°F, depending upon the air temperature required 

by these refrigeration loads. The use of multiple secondary fluid loops with 

temperatures closely matching the case air temperature requirements can result in more 

energy efficient operation of the secondary loop system. 

Secondary fluid loop piping consists of which convey all fluid flow to and from the 

chiller. The piping terminates with branches which feed chilled fluid to the display case 

lineups. The piping for the main and branches is typically sized at a fluid velocity of 4 –

6 ft/sec. In each lineup, further branching occurs to provide flow to each display case 

coil. A temperature control valve is used for each case coil to control case air 

temperature. The fluid flow is regulated to maintain the air temperature set point. The 

16 
branch piping is equipped with balancing valves, used to set the branch flow rate. This 

adjustment is normally done at installation. 

The piping for the secondary fluid loops may be steel, copper, or plastic. The 

recommended type of plastic piping is constructed of high-density polyethylene (4). 

Insulation is also required. For the medium temperature piping, the recommended 

insulation is closed-cell foam (4). The recommended thickness of the insulation is 1⁄2 - 3⁄4 

in. for piping underground or in air-conditioned space, 1 1⁄2 in. for pipe in nonconditioned space. For the low temperature piping, the insulation should be either 

Styrofoam or Polyisocyanurate foam. The recommended insulation thickness for low 

temperature piping is 1 in for piping in air-conditioned space and 1 1⁄2 to 3 in. in nonconditioned space. 

Secondary fluid flow rates are moderately high, since it is desirable to limit the 

temperature change of the fluid to 7 to 10°F while refrigerating the display cases. The 

resulting total flow rate for each loop is as high as 300 to 500 gpm. Because of the high 

viscosity of the secondary fluid at refrigerating temperatures, and the fact the fluid is 

circulated continuously, the energy associated with pumping is substantial and 

therefore, is a major component of the overall energy consumption of the secondary loop 

system. 

The central chiller systems employ multiple compressors that are piped in parallel. The 

compressors are multiplexed and are on/off cycled in response to the suction pressure 

of the chiller evaporator. 

Because of the location of the evaporator on the chiller skid, the compressors for the 

secondary loop system are considered close-coupled to the evaporator. The pressure 

drops and return gas heat gain are minimized in this configuration. Both these factors 

help reduce compressor energy consumption. 

The chiller system for the low temperature refrigeration may be equipped with 

mechanical subcooling. A direct expansion heat exchanger is used to subcool the liquid 

refrigerant used by the low temperature chiller system. 

Secondary loop refrigeration systems employ warm secondary fluid to defrost the 

display cases. In this system, a heat exchanger at the chiller is used to heat a portion of 

the fluid using discharge gas from the compressors. The warm fluid is piped to the 

cases using an additional pipe. The fluid leaving the defrosted case is returned to the 

chiller through the common return piping system. The use of warm fluid to defrost has 

been found to shorten defrost time substantially (2) and also reduce the amount of 

refrigeration associated with pull-down and recovery of the case back to operating 

temperature. It is estimated that the energy use for defrost is half that used by multiplex 

systems using hot gas defrost (4). 

2.1.2 High-Efficiency Features for Secondary Loop Refrigeration 

The purpose of the research project is to investigate methods which maximize the 

energy efficiency of the secondary loop refrigeration system. The features chosen for 

this purpose include the following: 

17 
• Display cases designed for use with secondary fluid 

• High-efficiency reciprocating compressors 

• Close-coupling of the evaporator to the compressor system 

• Multiple-parallel pumps 

• Evaporative heat rejection 

• Low viscosity secondary fluid 

• Warm-fluid defrost 

2.1.3 Display Cases Designed for Use with Secondary Fluid 

A significant factor affecting the energy consumption of the secondary loop system is 

the SST at which the refrigeration must operate to meet the temperature requirements of 

the display cases. Without proper display case design, the SST of the secondary loop 

system can be lower than that seen in a comparable direct expansion refrigeration 

system because of the added temperature difference associated with the secondary fluid. 

In previously installed secondary loop systems, the display cases were often the same as 

those used in direct expansion systems. The heat exchanger coils in these cases were 

designed specifically for refrigerant and evaporation and did not provide sufficient heat 

transfer surface for secondary fluid use. The heat exchangers used required a larger 

temperature difference between the secondary fluid and the case air to achieve the 

desired display case air temperature. 

Prior to this project, Hill PHOENIX, Inc. (manufacturer and contractor for the field test 

secondary loop system) had initiated an extensive redesign of the display case heat 

exchanger coils. As a result the coils are better suited for use with secondary fluid. The 

heat exchanger coils are larger than those used for direct refrigerant expansion, resulting 

in a lower temperature difference to achieve the desired air temperature. 

Table 1 shows a comparison of the required operating conditions for secondary loop and 

direct expansion refrigeration for similar display cases. For each display case type, the 

discharge air temperature of the case is listed. The expected saturated evaporator 

temperature and compressor rack SST needed to maintain this air temperature with 

direct expansion refrigeration is shown. The comparable secondary fluid supply 

temperature and SST of the chiller are also provided. For the cases listed, the SST values 

of the compressor rack and the chiller are the same, indicating the secondary loop 

system has no energy penalty because of the use of the fluid loop. There are two reasons 

for this result: the first is that the heat exchangers of the secondary loop display cases 

have been sized to utilize a fluid at approximately the same temperature as the directexpansion display cases; and secondly, the compressors of the direct expansion system 

are located remotely from the case evaporators and a pressure drop occurs in the 

refrigerant suction lines which forces the compressors to operate at an SST that is 

significantly lower than the saturated evaporator temperature at the display case. The 

secondary loop refrigeration system has the evaporator chiller and compressors close 

coupled (explained in detail below), which eliminates the pressure drop due to the 

piping runs necessary for the direct expansion system. 

18 
Table 1. Comparison of Operating Temperatures for Direct-Expansion and SecondaryFluid Refrigerated Display Cases 

Display Case & Discharge Air 

Temperature 

Direct-Expansion Refrigeration Secondary-Fluid 

Refrigeration 

Display Case Discharge Air 

Temp

(°F) 

Evaporator Sat. 

Temp

(°F) 

Compressor 

Rack SST

(°F) 

Secondary

Fluid Supply

Temp 

(°F) 

Chiller SST

(°F) 

Frozen Food 

Reach-in 

-12 -19 -25 -20 -25

Single-deck 

Meat 

28 21 15 20 15

Multi-deck Deli 29 21 15 20 15

Multi-deck 

Produce 

32 24 20 25 20

2.1.4 High-efficiency Refrigeration Compressors 

The chiller system may employ either reciprocating or screw compressors. Screw 

compressors can provide the large refrigeration capacities needed to address the 

refrigeration load associated with each fluid loop. Screw compressors do not exhibit as 

high an energy efficiency value as reciprocating units, but are less susceptible to liquid 

refrigerant damage and, in general, are considered to be less of a maintenance issue. 

Reciprocating compressors may be used in the chiller system for better energy use 

characteristics. 

Both reciprocating and screw compressors were evaluated for the advanced secondary 

loop refrigeration system. High-efficiency reciprocating compressors were chosen, 

because of the potential for energy savings. 

2.1.5 Close-Coupling of Compressors and Evaporator 

One of the energy-saving features of the secondary loop system is the close-coupling of 

the compressor system to the evaporator of the chiller. With this arrangement, the 

pressure drop between the evaporator and the compressors and the heat gain to the 

suction gas are minimized. Direct expansion systems incur substantial energy penalties 

because of the long pipe runs between the display cases and compressor racks, which 

result in pressure drops of 2-3 psi and as much as 40°F suction gas heating. Both of 

these penalties reduce the mass flow rate of the compressors, which results in loss of 

refrigeration capacity. Compressor run times are increased in direct expansion systems, 

which increase compressor energy consumption. 

The evaporator chillers were sized so the difference between the chiller evaporator and 

the brine supply temperatures was minimized to a delta T of 5°F. This small difference 

allowed the compressors to operate at a higher saturated suction temperature, which 

helps to maximize the EER. 

2.1.6 Multiple Parallel Pumps 

The design secondary fluid flow rates represent the maximum brine flow rate required 

to refrigerate the display cases. When the refrigeration load is at less than design, a flow 

control valve in each case throttles the brine flow to maintain discharge air temperature 

19 
set point. Display cases in a defrost cycle do not require chilled brine flow. For the 

above reasons, the total brine flow of each loop is often considerably less than the design 

value. The total loop flow cannot be reduced if a single pump is employed. Throttling 

of the pump to reduce flow reduces total pumping head is required at the display cases 

which are in a cooling mode to insure proper brine flow through the heat exchanger. 

The standard approach is to provide a fluid by-pass at the pump so that the excess flow 

is returned to the suction of the pump, while the full pressure head is maintained. 

Pump power is constant when a bypass is used to reduce system capacity when the 

system requires a constant head pressure. 

A variable-speed pump could be used, but the matching of the head-flow characteristics 

of the pump to the head and flow values needed in the brine loop can be difficult, if not 

impossible. Often, the pump must operate at full speed to meet the head required even 

though a lower flow rate is appropriate. The expense of a variable-speed drive for the 

flow rates associated with the secondary loop system for supermarkets is quite high and 

the energy savings often do not justify this added cost. 

The approach investigated is the use of multiple, parallel pumps with each pump sized 

at approximately 1/3 of the design brine flow rate with the pump head chosen to meet 

the full flow condition. Pumps are added or subtracted based upon the pressure 

difference between the brine supply and return flow. A rise in the pressure difference 

indicates flow valves have restricted flow or closed, reducing the total brine flow. When 

a pressure rise occurs, a pressure switch stops one of the pumps. A continuing rise in 

pressure causes the second pump to stop. The third pump continues to run to insure 

that brine flow to the display cases is continuous. A lowering of the pressure difference 

signifies more brine flow is required by the system. One of the two remaining pumps is 

started and the second added pressure continues to drop to the start set point of the 

third pump. 

2.1.7 Evaporative Heat Rejection 

Two types of condensers are available for supermarket refrigeration systems, air-cooled 

or evaporative. 

In dry climates, similar to the climate in California, the temperature difference between 

the dry-bulb and wet-bulb temperature is significant, and can be as high as 25°F. When 

an evaporative condenser is employed with a refrigeration system, the decrease in 

condensing temperature is substantial, resulting in significant energy savings in the 

operation of the refrigeration compressors (5). Ambient airflow required for an 

evaporative condenser is less than a comparably sized air-cooled unit; therefore, the fan 

power of an evaporative condenser is also less. 

The use of evaporative condensing for the secondary loop system is recommended to 

minimize energy use. 

2.1.8 Low Viscosity Secondary Fluid 

The pumping power needed for the secondary fluid loops is greatly influenced by the 

thermal capacitance (the product of the fluid density and specific heat) and viscosity of 

the fluid. Higher thermal capacitance means that less mass flow is required to meet a 

20 
refrigeration load, reducing the pumping power. Viscosity is a key parameter of fluid 

friction, and greatly influences pumping power. This is of particular importance for low 

temperature refrigeration applications, since fluid viscosity increases as temperature 

decreases. Secondary loop systems often employ propylene glycol-water solutions as 

the brine, wherein the concentration of the glycol is set to avoid ice formation. Glycols 

are preferred because they are inert to all common piping materials and most nonmetallic gaskets and seals. Propylene glycol is also non-toxic and non-flammable and is 

the most suitable glycol for use where food products are involved. A 30% propylene 

glycol / water mixture shows reasonable heat transfer and pumping properties for 

medium temperature (+20°F fluid temperature). In low temperature refrigeration, the 

brine temperature will fall in the range of -20 to -30°F. The required concentration of 

glycol for anti-freeze in this temperature range is 50% and the resulting viscosity is 156.1 

centipoise (for comparison, the viscosity of water at 68°F is 1.0 centipoise). 

Alternative secondary fluids to glycol-water solutions exist that will remain liquid at the 

required operating temperatures and also have better transport properties. These fluids 

have been evaluated by others (6). The results of these evaluations show the preferred 

secondary fluid is an organic-salt and water mixture, which has lower viscosity than 

propylene glycol-water and also has higher thermal capacitance. 

The organic-salt and water solution investigated in this project is marketed under the 

trade name of Dynalene. Table 2 gives a comparison of the relevant transport properties 

of propylene glycol-water and Dynalene at fluid temperatures of -20 and 20°F, which are 

standard values for low and medium temperature refrigeration applications. Two 

Dynalene compositions specified are HC-30 and HC-10 for low and medium 

temperature refrigeration, respectively. The differences in properties between 

propylene glycol-water and Dynalene are most notable at -20°F, particularly in terms of 

viscosity. 

Table 2. Transport Properties of Propylene Glycol-Water and Organic Salt-Water 

(Dynalene) at Operating Temperatures of Secondary Loop Refrigeration 

Property Propylene Glycol- Water Dynalene 

Brine temperature = + 20°F

Concentration (% by weight) 30 HC-10 

Density - ρ (lb/ft3) 64.9 75.0

Specific Heat - C (Btu/lb- °F) 0.90 0.80 

Thermal Capacitance (ρ×C) 58.4 60.0

Viscosity (centipoises) 9.91 2.84 

Brine Temperature = - 20 °F

Concentration 50 HC-30

Density - ρ (lb/ft3) 66.5 81.2

Specific Heat - C (Btu/lb- °F) 0.8 0.77

Thermal Capacitance (ρ×C) 53.2 62.5

Viscosity (centipoises) 156.1 8.78 

The impact of the fluid transport properties on the energy consumption of secondary 

loop refrigeration can be determined through evaluation of the relative pumping power 

(4,6). 

21 
The relative pumping power is found by holding the refrigeration load, fluid 

temperature difference, pipe diameter and pipe length constant. The relationship of 

heat transport, fluid friction and pump power are combined to give an expression for 

pumping power in terms of fluid properties only, relative to a base fluid: 

2.8

B

A

1.8 0.2

B

A

B

A

c

c

Pumping Power

Pumping Power − −

⎟

⎟

⎠

⎞

⎜

⎜

⎝

⎛

⎟

⎟

⎠

⎞

⎜

⎜

⎝

⎛

⎟

⎟

⎠

⎞

⎜

⎜

⎝

⎛ =

B

A

μ

μ

ρ

ρ

Using propylene glycol-water as the base fluid (B), the relative pumping power of 

Dynalene are 0.44 and 0.84 for fluid temperatures of -20 and 20°F, respectively. 

These results indicate the use of lower viscosity fluid in both the medium and low 

temperature fluid loops will reduce energy use for brine pumping. For the advanced 

secondary loop refrigeration system, Dynalene was specified for all fluid loops. 

2.1.9 Refrigerant Subcooling from Warm Brine Defrost 

During operation, the cooling coils of the refrigerated display cases and walk-in coolers 

develop frost build-up from condensation and subsequent freezing of water vapor from 

air passing through the coils. The frost must be removed periodically. Otherwise the 

coils will clog, preventing airflow and reducing the cooling capability of the fixture. The 

secondary loop refrigeration system employs a flow of warmed brine for this defrosting 

process. Brine is taken from the refrigeration loop and heated with refrigerant rejection 

heat to a temperature of approximately 65 or 80°F for medium and low temperature 

refrigeration, respectively. The warm brine flows to the display cases by a separate 

piping loop used exclusively for defrosting. When a display case is scheduled to defrost, 

a valve in the warm brine supply line opens, providing warm brine to the display case. 

The brine passes through the case heat exchanger where heat from the brine melts the 

frost. The brine then flows to the return piping of the loop where it is pumped back to 

the chiller system. It is important to note that the melting of the frost chills the brine 

close to the normal return temperature therefore little excess heat is added to the total 

refrigeration load by the defrosting. 

The heating of the brine for defrost is done with two heat exchangers at the chiller 

system. The first exchanger uses heat from the liquid refrigerant coming from the 

system receiver. The second heat exchanger uses the heat contained in the compressor 

discharge gas to raise the temperature of the brine to the desired temperature. 

The initial heating of the brine by the hot liquid refrigerant results in lowering 

significantly the temperature of the refrigerant, producing a subcooled liquid. The use 

of this sub-cooled refrigerant liquid in the chiller system increases the chilling capacity 

of the system, which means fewer compressors need to operate to satisfy the 

refrigeration load. The reduction in compressor run time results in a reduction in 

compressor energy used. 

22 
Field testing showed the refrigerant subcooling produced by the warm brine defrost 

averaged about 12°F, which corresponds to capacity increases of 6.9 and 6.4% for low 

and medium temperature refrigeration, respectively. 

2.2 Analysis of the Secondary Loop and Multiplex Refrigeration Systems 

Two refrigeration system models were formulated for the analysis work done on this 

project. The first model was used to determine the energy consumption of a direct 

expansion multiplex refrigeration. The second model for the secondary loop system 

used a modified version of the multiplex system to describe the compressors and 

condensers. Models for the brine loops and pumping were then added to the 

compressor and condenser models to form the full secondary loop refrigeration system. 

2.2.1 Multiplex Refrigeration Model 

The annual performance of the multiplex refrigeration system is calculated on the basis 

of ambient dry-bulb temperature bins, where each bin specifies: an ambient dry-bulb 

temperature value; the coincident value of the wet-bulb temperature; and the number of 

hours at which the ambient temperature occurs during the year. The general procedure 

is, for each temperature bin, to calculate the power needed by the refrigeration system 

and apply that power to the number of hours at each ambient temperature. The 

procedure is repeated for all ambient temperatures present at the site. 

A refrigeration configuration must also be specified, where each refrigeration system 

employed in the supermarket is described in detail. System information identifies the 

design refrigeration load, display case evaporator temperature in the system, minimum 

condenser temperature, type of condenser (air-cooled or evaporative), and the 

refrigerant employed. Other information specified includes saturated temperature 

change between the display cases and the suction of the compressors, which is 

representative of the suction pressure drop, and the temperature rise in the return gas. 

The first step in the analysis is to determine the refrigeration load on the system. Past 

experience (7) has shown that the refrigeration load will vary with outside ambient drybulb temperature, decreasing as the ambient temperature decreases. The rate of 

decrease in load is greater for medium temperature than low temperature refrigeration, 

and a minimum is reached, which is a result of space heating the store. Store 

temperature will not fall below 68°F. The minimum store temperature is normally 

present at outside ambient temperatures less than 40°F. The maximum values of store 

temperature and humidity are also constrained by the use of air conditioning, which 

tends to maintain the temperature at 75°F with a corresponding relative humidity of 

55%. Maintaining the indoor conditions at these levels can be expected at outdoor 

ambient temperatures of greater than 85°F. Based on these constraints, a load factor can 

be calculated which is applied to the design load to determine the refrigeration load for 

each temperature bin. The refrigeration load factor is determined from the following 

relationships: 

For both low and medium temperature refrigeration, the value of the refrigeration load 

is set at design for ambient temperatures of 85°F or greater. 

23 
At ambient temperatures of 40°F or less, the refrigeration load is at its smallest fraction 

of design value, which is 66% for medium temperature and 80% for low temperature. 

For ambient temperatures between 40 and 85°F, the load factor is found from: 

( ) ⎟

⎟

⎠

⎞

⎜

⎜

⎝

⎛ − = (85 - 40)

(85 T ) Load factor 1- 1- min amb

where 

min = the minimum fraction of design load (0.66 for medium temperature and 

0.8 for low temperature 

Tamb = ambient dry-bulb temperature 

 

The state points for the multiplex refrigeration system are then determined. The system 

configuration specifies the desired evaporator temperature for the display cases. In 

operation, pressure drop will occur between the case evaporators and the compressor 

suction. This drop is reflected in a lower saturated temperature value at the compressor 

suction. Heat gain to the return gas will also take place, which affects the mass flow rate 

of refrigerant seen by the compressor. Both of these factors tend to decrease the capacity 

of the compressors and increase the run time need to meet the refrigeration load. The 

amount of pressure drop and superheating is a strong function of the distance between 

the display cases and the compressors, increasing with increased distance. In the 

analysis, these factors are taken into account by values included with the system 

configuration description. 

The condenser temperature is also determined at this time. The most significant 

parameter in determining condensing temperature is the outdoor ambient temperature, 

since heat is rejected to ambient conditions. The operation of the condenser can be 

characterized by the temperature difference (TD) between the condensing refrigerant 

and the ambient. The condensing temperature is allowed to vary with the ambient until 

a certain minimum condensing temperature is reached. At that point, control of the 

condenser fans or a liquid pressure regulator maintains the condensing temperature at 

the minimum value. The model compares the ambient temperature with the 

characteristic TD of the condenser type specified and calculates a condensing 

temperature. The calculated value is compared with the specified minimum condensing 

temperature. If the calculated temperature is less than the minimum, the minimum 

value is used to set the state point of the refrigeration system. 

The condenser TD is determined by the type of condenser modeled. For air-cooled 

condensers, the TD refers to the difference between the condensing and ambient drybulb temperatures. The standard values of TD for air-cooled condensers in supermarket 

refrigeration are 10 and 15°F for low and medium temperature, respectively. The TD of 

an evaporative condenser refers to the difference between condensing and ambient wetbulb temperatures. Evaporative condensers are often sized to produce a condensing 

temperature of 100°F at the design wet-bulb, however, analysis and field measurements 

by Foster-Miller (8) showed the lowest combined compressor and condenser fan power 

24 
is achieved at a TD value of approximately 12°F. This lower value was used for all 

relevant analysis. 

The refrigerant liquid temperature is determined as part of the state points of operation. 

For non-subcooled systems, the liquid temperature is 10 degrees less than the 

condensing temperature. When mechanical subcooling is employed in multiplex

systems, the liquid refrigerant temperature leaving the subcooler heat exchanger is 

typically 40°F. Some warming of the liquid occurs as the liquid is piped to the display 

cases .The temperature of the liquid entering the cases is set at 42°F. The refrigeration 

needed for mechanical subcooling is normally provided by the highest temperature 

refrigeration system in operation, typically at a SST of 25 - 35°F. The size of the 

mechanical subcooling load varies with the load of the subcooled system, normally the 

low temperature refrigeration. For each temperature interval the low temperature 

refrigeration load is first determined, and the liquid refrigerant flow required for this 

load is then determined. The subcooling load is the amount of cooling needed to lower 

the temperature of the liquid flow from 10°F less than the condensing temperature to 

40°F. The subcooling load is added to the refrigeration load of the medium temperature 

system and is used to determine the energy consumption for that system. 

Once the state points are determined, the capacity and power of the compressors is 

found. These calculations are made using the compressor performance data supplied by 

the manufacturers. Performance data consists of refrigeration capacities (Btuh), 

refrigerant mass flow rate (lb/hr), and input power (Watts) as functions of saturated 

suction temperature (SST) and saturated discharge temperature (SDT). These data are 

obtained from the ARI equation: 

Capacity, Mass Flow, or Power = C0 + C1*SST + C2*SDT + C3*SST2

+ C4*SST*SDT + C5*SDT2 + C6*SST3

+ C7*SDT*SST2 + C8*SST*SDT2

+ C9*SDT3

where 

C0 ... C9 = Performance equation coefficients 

SST = Saturated suction temperature (°F) 

SDT = Saturated discharge temperature (°F) 

Compressor manufacturers provide three sets of coefficients for each compressor, where 

each set is to determine either cooling capacity, mass flow rate, or compressor input 

power. 

 The compressor cooling capacity and mass flow rate given by the above equations are 

determined at particular rating conditions. One such condition commonly seen is a 

return gas temperature of 65°F and 0°F of liquid subcooling. Corrections are made to 

account for the values of superheat and refrigerant liquid temperature. The superheat 

correction takes into account the density and enthalpy change, while change in liquid 

temperature affects the enthalpy of the refrigerant entering the evaporator. The 

correction applied to the compressor capacity or mass flow rate is: 

25 
( )

(h ) *

v

v correction r out

rout rin

in

h h

h Capacity −

− =

where 

v = the specific volume of the refrigerant (ft3

/lb) 

hin = the enthalpy of the refrigerant entering the evaporator 

hout = the enthalpy of the refrigerant leaving the evaporator 

the subscript, r, designates that the property is at the rating conditions 

The capacity value and refrigeration load are then used to find the number of 

compressors operating by taking the ratio of refrigeration load to capacity. Typically, 

three or four compressors are needed to meet the load at design conditions. At other 

conditions less than this number is required. Fractional values represent compressor 

on/off cycling. The analysis does not use specific compressor models, but instead uses a 

single generic size for each type of compressor. The generic size is based upon the most 

commonly compressor, which is a 7.5 HP unit for the reciprocating compressor and a 6 

HP unit for a scroll compressor. 

Compressor energy consumption for the temperature bin is found by first determining 

the power needed by the compressor at the state point and load conditions. Compressor 

power is multiplied by the number of compressors operating and the number of hours 

associated with the ambient temperature. 

The fan power for remote condensers or fluid coolers is based upon the type of 

condenser or cooler being utilized. Air-cooled condensers for low temperature 

refrigeration are sized for a smaller TD and require more fan power than condensers 

employed with medium temperature refrigeration. Fan requirements are less for 

evaporative heat rejection than is needed for dry rejection, because less air flow is 

required. The power value listed in the table for the evaporative units includes the 

sump pump used to spray water over the heat exchanger coil. 

The condenser fans operate continuously as long as the resulting condensing 

temperature is greater than the specified minimum value. Fan cycling is employed with 

both the condensers and fluid coolers to regulate the condensing temperature when full 

operation of the fans reduces condensing temperature below the minimum value. Fan 

energy is estimated by multiplying the installed fan power by a fan factor that 

represents the amount of fan operation needed to maintain the condensing temperature 

at the minimum. For air-cooled condensers and dry heat rejection, the analysis sets the 

fan factor at 1.0 when the sum of the ambient dry-bulb temperature and the condenser 

TD are greater than the minimum condensing temperature. The fan factor is set at 0.25 

when the ambient dry-bulb temperature is less than 30°F. For ambient temperatures 

greater than 30°F where continuous fan operation is not needed, the fan factor is 

calculated from: 

⎟

⎟

⎠

⎞

⎜

⎜

⎝

⎛ = (T - TD - 30)

(T - TD - T ) Fan Factor 1- 0.75*

con

con amb

where 

26 
Tcon = the minimum condensing temperature (°F) 

TD = the temperature difference (°F) 

Tamb = the ambient dry-bulb temperature (°F) 

For evaporative rejection, the above relation is also employed, but the ambient wet-bulb 

is used instead of the dry-bulb temperature. The fan energy for the temperature bin is 

determined from the product of the installed fan power, the fan factor, and the number 

of hours in the bin. 

The energy for mechanical subcooling of the low temperature refrigeration is addressed 

by the medium temperature system with the highest SST. As mentioned previously, the 

subcooling load is calculated for each temperature bin and is added to the refrigeration 

load of the appropriate medium temperature system. The compressor run time for the 

medium temperature system is calculated on the basis of this combined load. 

The bin calculation is repeated until energy values are set for each temperature bin for 

the refrigeration system configuration specified. Once the bin loop is completed, the 

model then obtains the next system description and the bin loop is repeated for this next

system. The procedure continues until all systems are analyzed. 

2.2.2 Modeling of Secondary Loop Refrigeration 

The major difference in the analysis of the secondary loop refrigeration system is the 

operation of the secondary loop. The loop consists of secondary fluid that is pumped 

between a central chiller and the display cases. Two to four fluid loops are employed in 

a supermarket, depending upon the composition of the refrigeration load. The analysis 

is, therefore, conducted separately for each of these loops at each ambient temperature. 

Energy results are combined at the completion of the analysis to determine total energy 

consumption. 

The system configuration specifies the design refrigeration load for each secondary loop. 

The analysis first considers the variation on the refrigeration load seen at each ambient 

temperature, using the method described previously. The model assumes that the 

refrigeration load at the display cases is handled by a constant temperature change of 

the fluid, while the flow through the cases is varied as the refrigeration load varies. This 

flow arrangement is an attempt to simulate the operation of a temperature control valve 

that maintains constant fluid outlet temperature from the display case heat exchanger. 

Since all loads on the loop behave in this fashion, the estimated total fluid flow can be 

found from: 

brine brine

ref brine C T

Q M

∆ = •

where 

Qref = the refrigeration load addressed by the brine loop (Btuh) 

= the massflow rate of brine (lb/hr) •M brine 

27 
∆ = the temperature change in the brine seen at the cases Tbrine

= the specific heat of the brine Cbrine

The secondary fluid loop will experience some heat gain, while flowing between the 

cases and the central chiller. The most significant of these gains is the addition of energy 

due to operation of the secondary fluid loop pump. The pump power is based on the 

maximum fluid flow needed to meet the design refrigeration load, which is found with 

the above relation. The required pump head is set at 75 ft of water column (WC) for low 

temperature refrigeration loop and 50 ft (WC) for the medium temperature loop. While 

the flow to the cases varies as the load changes, the total fluid flow through the pump 

remains constant, since a by-pass is used to regulate the operation of the pump in the 

loop. The power input to the pump is calculated as the ideal power for the maximum 

fluid flow and head divided by a pump efficiency of 55%. The power input to the pump 

is converted into heat in the fluid, which must be removed by the chiller system. The 

rise in temperature is calculated from: 

brine brine

pump

C

T • ∆ =

M

Pump Power

where 

Pump Power = the power input to the secondary fluid pump 

Some line heat gain is also expected and was set at 0.25°F for the supply and return lines 

of the loop. 

Once the total temperature rise of the secondary fluid loop is determined, the load on 

the chiller evaporator can be found by: 

brine brine brine Qevap = M C ∆T •

where 

∆Tbrine = the total temperature gain seen in the fluid 

Mechanical subcooling is used in the secondary loop system for the low temperature 

chiller system in the same fashion as is seen in multiplex systems. A portion of the 

medium temperature system provides the subcooling. The refrigeration load associated 

with the subcooling must be added to the total load of the medium temperature system. 

The subcooling load is calculated based upon the refrigeration load of the low 

temperature system, which sets the flow rate of refrigerant needed. The subcooling 

reduces the temperature of the refrigerant from the temperature leaving the condenser 

to 40°F. The sub-cooled liquid temperature is factored into the available capacity of the 

low temperature system in meeting its refrigeration load. 

The state points of the chiller system are then determined. The evaporator temperature 

of the chiller heat exchanger is set at 7°F below the outlet temperature of the fluid loop. 

28 
The outlet temperature of the refrigerant is set at 8°F higher than the evaporator 

temperature. The temperature rise is due primarily to the control action of the 

thermostatic expansion valve of the chiller heat exchanger, which regulates the outlet 

temperature of the refrigerant at a superheated condition. Pressure drop of the 

refrigerant vapor to the suction of the compressors is negligible due to the close 

coupling of the compressors and the heat exchanger. The SDT of the compressor system 

is determined from the condensing temperature, which is calculated depending upon 

the type of heat rejection system analyzed. The secondary loop refrigeration system can 

be modeled with air-cooled, water-cooled, or evaporative condensing. The method of 

determining the condensing temperature for each heat rejection type is the same as 

described previously. 

The central chiller uses multiple parallel compressors to address the chiller heat 

exchanger load. The types of compressors now employed in presently installed systems 

are either screw or reciprocating. Scroll compressors can be analyzed if desired. The 

procedure for determining the number of compressors operating is the same as that 

used for the multiplex and distributed systems. Manufacturer’s performance data are 

used to determine compressor cooling capacity and power. The number of compressors 

operating is found from the ratio of the chiller cooling load to the total available 

compressor capacity. Compressor energy is the product of the compressor power, 

number of compressor operating, and the number of hours in the temperature bin being 

examined. 

The energy consumption of the secondary fluid pump is determined using the method 

outlined previously. While the fluid flow to the display cases varies with changing 

refrigeration load, the total fluid flow across the pump is constant. Variation in flow is 

achieved by flow bypassing around the pump. The head addressed by the pump was 

estimated based upon the temperature of the secondary fluid loop, the secondary fluid 

employed, the length and diameter of pipe between the cases and chiller, and the 

pressure drop at the display cases and through the chiller heat exchanger. The 

secondary fluid examined here consists of an organic salt-water mixture (Dynalene) for 

the low and medium temperature loops. Loop piping diameter was sized to maintain a 

velocity of 4to 6 ft/sec, while the typical length of piping was estimated at 250 ft. Value 

of the pressure drop for the display cases ranges from 5 to 7 ft. (WC), while the pressure 

drop of the chiller heat exchanger was set at 20 ft (WC). The head requirements 

calculated for the secondary fluid loop pumps were 50 and 75 ft. (WC) for the medium 

and low temperature systems, respectively. The pump and motor efficiencies were 

taken at 55 and 85%, respectively. 

During part-load operation, the flow rate of the secondary loops is less than design but 

the total head remains approximately the same due to the pressure drops through the 

operating display case coils. Pump system control was modeled as continuous. The 

model consisted of calculating the required fluid flow to meet the refrigeration load and 

then using that flow to calculate the pump power. The energy for that temperature bin 

was the product of the calculate pump power and the number of hours associated with 

that temperature bin. 

The energy requirement for heat rejection is dependent upon the type of heat rejection 

employed. The energy is calculated using the procedures described previously. 

29 
2.3 Field Testing of the Secondary Loop Refrigeration System 

2.3.1 Description of the Secondary Loop Refrigeration Test Site 

The design and analysis work of the project were used to specify a secondary loop 

refrigeration system for an operating supermarket. A supermarket was chosen for the 

testing of the secondary loop system and is a store operated by Vons Supermarkets (a 

division of Safeway, Inc.). The supermarket is located in Thousand Oaks, CA and has a 

total building and sales areas of 48,351 and 41,912 ft2, respectively. The layout of the 

store is shown in Figure 4 and the circuits that are refrigerated by the secondary loop 

system are listed in, Table 3, Table 4, and Table 5. 

Figure 4. Layout of the Secondary Loop Refrigeration Test Store 

30 
Table 3. Low Temperature Circuits for the Secondary Loop Refrigeration Test Store 

(Loop A) 

Circuit Fixture Description

Refrig Load 

 (Btuh) 

Disch. Air 

Temp 

(°F) 

Brine Flow 

(gpm) 

A-1 Reach-in Frozen Pasta 3,000 -5 0.9 

A-2 Wide Island Frozen Meat 19,800 -12 18.2

A-3 Wide Island Frozen Meat 17,400 -12 16.2

A-4 Freezer Walk-in Box 31,300 -12 24 

A-5 Reach-in Frozen Food 24,000 -5 6.9 

A-6 Reach-in Frozen Food 21,000 -5 6 

A-7 Reach-in Frozen Food 24,000 -5 6.9 

A-8 Reach-in Frozen Food 21,000 -5 6 

A-9 Reach-in Ice Cream 22,400 -12 12.6

A-10 Reach-in Ice Cream 19,200 -12 10.8

A-11 Reach-in Ice Cream 25,600 -12 14.4

A-12 Reach-in Ice Cream 22,400 -12 12.6

A-13 Bakery Freezer 9,400 -5 4

A-14 Deli Freezer 9,400 -5 4

 Pump Heat 11,400 

 Total Refrigeration Load 281,300 144 

Brine Supply Temperature = -20°F Chiller SST = -25°F 

31 
Table 4. Medium Temperature Circuits for the Secondary Loop Refrigeration Store (Loop 

B1) 

Circuit Fixture Description

Refrig Load 

 (Btuh) 

Disch. Air 

Temp 

(°F) 

Brine Flow 

(gpm) 

B1-1 Deli Walk-in Cooler 8,600 30 2.5 

B1-2 Multi-deck Lunch Meat Cases 50,200 30 15.7 

B1-3 Meat Holding Box 6,300 30 2 

B1-4 Service Seafood Case 5,000 26 2.4 

B1-5 Single-deck Meat Cases 3,300 26 1.6 

B1-6 Single-deck Meat Cases 11,600 26 5.6 

B1-7 Multi-deck Meat Cases 37,800 28 19.6 

B1-8 Meat Walk-in Cooler 28,000 30 16 

B1-9 Island Cake Display Case 5,000 28 6.9 

B1-10 Self-service Bakery Case 5,000 28 5 

B1-11 Multi-deck Deli Case 56,500 30 17.6 

B1-12 Multi-deck Deli Case 44,000 30 13.7 

B1-13 Multi-deck Deli Case 25,100 30 7.8 

B1-14 Island Cheese Case 7,200 28 10.3 

B1-15 Island Prepared Foods Case 7,200 28 10.3 

B1-16 Service Deli Cases 6,400 31 8 

B1-17 Sandwich Cases 14,600 28 13.5

 Pump Heat 16,000 

 Total Refrigeration Load 337,800 159 

Brine Supply Temperature = 20°F Chiller SST = 15°F 

32 
Table 5. Medium Temperature Circuits for the Secondary Loop Refrigeration Store 

(Loop B2) 

Circuit Fixture Description

Refrig Load 

 (Btuh) 

Disch. Air 

Temp 

(°F) 

Brine Flow 

(gpm) 

B2-1 Flower Walk-in Cooler 8,600 38 4 

B2-2 Melon Table 6,400 35 3.8

B2-3 Multi-deck Produce Cases 37,600 36 23.8 

B2-4 Multi-deck Produce Cases 34,200 36 21.6 

B2-5 Multi-deck Produce Cases 9,000 36 3.2 

B2-6 Multi-deck Produce Cases 9,000 36 3.2 

B2-7 Multi-deck Produce Cases 36,000 36 12.7 

B2-8 Multi-deck Produce Cases 27,400 36 17.3 

B2-9 Produce Walk-in Cooler 23,900 38 14 

B2-10 Meat Prep Room 45,800 50 12 

B2-11 Multi-deck Egg Case 12,000 35 4.2 

B2-12 Dairy Walk-in Cooler with 14 Doors 46,200 36 27 

B2-13 Bakery Walk-in Cooler 7,300 36 4.5 

B2-14 Multi-deck Beverage 56,600 36 20.1

B2-15 Multi-deck Beverage 44,700 36 15.9

B2-16 Floral Case 7,800 36 4

B2-17 Pasta Case 12,000 31 4.3

 Pump Heat 21,800 

 Total Refrigeration Load 446,300 196 

 

 Low Temp Subcooling 49,800 

Brine Supply Temperature = 25°F Chiller SST = 20°F 

Figure 5 and Figure 6 show diagrams describing the operation of the secondary loop 

refrigeration system for low and medium temperature, respectively. The refrigeration 

load is divided into 3 secondary fluid circuits operating at -20, 20, and 25°F, respectively. 

There are three compressor suction groups to provide refrigeration to the secondary 

fluid loops operating at saturated suction temperatures of 25, 15, and 20°F. Multiple 

parallel compressors are employed in each of the suction groups so that compressors can 

be cycled on and off to provide capacity control by maintaining a constant suction 

pressure. The compressed gas is routed through two discharge manifolds; one manifold 

transports the gas from the -20°F compressors while the seconds carries the combined 

flow from the 15 and 20°F compressors. The compressor discharge gas may be used for 

heat reclaim for either water or space heating, or for defrost brine heating. The 

discharge refrigerant flow is piped to the evaporative condenser. From the condenser, 

the liquid refrigerant flow from the two discharge circuits is combined at a single 

receiver tank. The liquid refrigerant flow from the receiver is divided into two streams, 

which pass through heat exchangers that subcool the refrigerant by warming brine that 

is used for display case defrost. After these heat exchangers, the liquid refrigerant flow 

is combined at the liquid manifold where it is distributed to the three chiller 

evaporators. The only exception is the liquid refrigerant flow used at the low 

temperature (-25°F) chiller evaporator. The refrigerant liquid for the low temperature 

33 
system is further sub-cooled in a mechanical sub-cooling heat exchanger. The 

refrigeration for the mechanical subcooling is provided by the 20°F suction group 

compressors. 

Figure 5. Piping Diagram for the Low Temperature Secondary Loop Refrigeration System 

(Patent by Hill PHOENIX, Inc.)

34 
Figure 6. Piping for the Medium Temperature Secondary Loop Refrigeration System 

(Patent by Hill PHOENIX, Inc.) 

In each fluid loop, the secondary fluid is circulated between the chiller evaporator and 

the refrigerated fixtures by three parallel centrifugal pumps. The number of pumps 

operating is determined by the pump discharge pressure. Pumps are cycled off on a rise 

in discharge pressure and restarted when the pressure falls. 

2.3.2 Measurement Plan to Monitor Secondary Loop Refrigeration 

A measurement plan was developed to determine the necessary instrumentation to 

evaluate the performance of the secondary loop refrigeration. The key elements 

included in the evaluation were the following: 

• Refrigeration system energy consumption 

• Refrigeration supplied 

• Operating state points 

• Display case operating temperatures 

• Ambient temperature and humidity 

Details of these measurements are as follows: 

2.3.2.1 Refrigeration Energy Consumption 

Since energy efficiency is of primary concern to the project, power measurement of the 

refrigeration system was the primary measurement monitored. Power measurements of 

refrigeration compressors, brine pumps, and condenser fans and pumps were noted at 

regular intervals and used to determine energy consumption of the refrigeration system. 

35 
Power measurements were recorded at 10-second intervals which were later combined 

into 15-minute averages. The average values were stored at the data acquisition for 

remote collection. Watt transducers were employed for these measurements. 

Total store power and electric energy were also monitored and recorded. 

2.3.2.2 Refrigeration Supplied 

Differences in the refrigeration loads addressed at the two sites (Secondary and 

Multiplex), must be taken into account when comparing energy use of the two systems. 

Refrigeration supplied by the secondary loops and the chiller systems were calculated 

based upon several instrument readings taken at regular intervals. The refrigeration 

supplied by each brine loop was determined from: 

( ) Qbrine = Tbin − Tbout mc

where 

Tb is the temperature of the brine in supply and return lines at the chiller 

evaporator 

m is the mass flow rate of brine in the loop 

c is the specific heat of the brine 

The refrigeration load delivered by the refrigeration to the chiller evaporator is found 

from: 

( ) out liq = m h − h Qchiller

where 

h is the enthalpy of the refrigerant at the outlet (out) and inlet (liq) of the chiller 

m is the mass flow rate of refrigerant entering the chiller. 

A third method was available to estimate the refrigeration provided by using the 

published capacity data for the compressors. The rated capacity of each compressor is 

given as an equation of the saturated suction and discharge temperatures at which the 

compressor is operating. Corrections are needed to account for the refrigerant 

temperature at the suction of the compressor and for the refrigerant liquid. The rated 

capacity was calculated with measured data for these quantities. The refrigeration load 

was found by multiplying the capacity by the operating run fraction for each 

compressor. Monitoring and recording on/off cycles of each compressor determined 

the run fraction. 

The results from the power and refrigeration measurements were then combined to 

determine the energy efficiency ratio (EER) of the refrigeration system. The EER is 

defined as: 

36 
(Btu/Watt hr) Power quired

frigerationSupplied EER = − Re

Re

For the secondary loop refrigeration, the power required included both the compressor 

and brine pumping power. Two EER values were calculated, one for low temperature 

refrigeration and a second for medium temperature refrigeration. The medium 

temperature EER combined the refrigeration and power of the 15 and 20°F fluid loops. 

Refrigeration provided by the 20°F loop for mechanical subcooling was subtracted from 

the medium temperature refrigeration load. 

2.3.2.3 Refrigeration System Operating State Points 

The significant operating state points that were monitored for the secondary loop system 

are the following: 

• Saturated Suction Temperature (SST) – This is the saturation temperature 

corresponding to the refrigerant pressure measured at the compressor suction. 

• Saturated Discharge Temperature (SDT)– The saturation temperature of the 

refrigerant based upon the pressure measured at the compressor discharge. 

• Return Gas Temperature – The temperature of the refrigerant gas measured at the 

suction of the compressor. 

• Refrigerant Liquid temperature – The liquid temperature was measured at several 

locations in the system to characterize the subcooling provided. The liquid 

temperature was measured at the outlets of the receiver and ambient subcooling coil, 

and at the inlet and outlet of the defrost heat exchanger. Additional liquid 

temperature measurements were needed at the outlet of the mechanical subcooling 

heat exchanger for the low temperature chiller. 

• Refrigerant discharge temperature – The temperature was measured at the inlet of 

the brine defrost heat exchanger, and at the inlets of the heat reclaim water heater 

(an outlet temperature was also required) and space heater coil. 

• Condenser Pan Water temperature – the primary control point for the condenser 

fans. 

• On/off digital signals were included to track events such as the initiation of defrost 

or heat reclaim. 

2.3.2.4 Brine Pump Operation 

The operation of the brine pumps was monitored to evaluate the pump control strategy. 

Along with pump power, digital on/off signals were included to monitor the operation 

of the brine pumps. 

37 
2.3.2.5 Brine Defrost Operation 

The warm brine defrost system specified in this project used a combination of 

refrigerant liquid subcooling and discharge heat to raise the brine temperature. The 

subcooling of the refrigerant increases the refrigeration capacity of the compressors and 

is an energy-saving feature for this system. Operation of the hot brine defrost was fully 

characterized as part of the site monitoring. The following measurements were included 

for each secondary loop: 

• Brine flow rate for defrost – a flow meter was installed in each brine defrost supply 

line. 

• Brine temperature in and out of defrost heat exchanger. 

• Brine temperature in and out of defrost/subcooling heat exchanger. 

2.3.2.6 Display Case Operation 

The operation of the display cases and walk-in boxes was monitored during testing. 

“Degree Master” modules were employed to control operation of the display cases. 

These modules regularly measured discharge and return air temperatures, along with 

brine inlet and outlet temperatures. 

Two display case circuits were also equipped with product simulators in each display 

case that allow measurement and tracking of product temperature. The 2 circuits 

selected consisted of: 

• Single-deck meat cases –three display cases 

• Multi-deck produce cases –two display cases 

Similar display cases at the multiplex store were also equipped with product simulators 

to allow comparison of product temperature and temperature variation for the two sites. 

Other case measurements that were addressed included the energy use for case lights, 

fans, and anti-sweat heaters (frozen food door cases). Power monitoring was included 

for this purpose. 

2.3.2.7 Ambient Conditions 

Both inside and outside ambient dry-bulb temperature and humidity were monitored. 

The inside reading consisted of multiple readings of the dry-bulb temperature and a 

single reading of the inside relative humidity. The outside ambient instrumentation 

consisted of single readings of the dry-bulb temperature and the relative humidity. 

Problems were incurred with the outside humidity reading shortly after the start of 

testing. An alternate approach for outside ambient data was employed after the sensor 

failed. Southern California Edison maintains weather stations throughout the service 

territory. One station is located in Moorpark, CA, which is in close proximity to the 

secondary loop store. Good agreement was found between dry-bulb temperature and 

relative humidity readings from the weather station and those taken at the store during 

a site visit. Weather station readings were used for the entire test period to represent the 

outside ambient conditions at the store. Similarly, weather station readings were also 

38 
used to characterize the ambient for the multiplex test store. Both the store and the 

weather station are located in Valencia, CA. 

2.3.2.8 Data Collection at the Secondary Loop Store 

All instrument readings at the secondary loop store were wired to the energy 

management system (EMS) at the site. The instruments were read by the EMS at twominute intervals and these readings were stored for later retrieval. Data were obtained 

from the EMS by a modem/phone line connection. The site was called daily and the 

data were downloaded. 

2.3.3 Field Testing of the Multiplex Refrigeration System 

2.3.3.1 Description of the Multiplex Refrigeration Test Site 

A second Vons supermarket, located in Valencia, CA, was selected and instrumented for 

performance comparison with the secondary loop refrigeration system. The store 

selected employed a direct expansion, multiplexed refrigeration system to provide 

refrigeration to all display cases and walk-in coolers. The total building area of the store 

is 56,526 ft2

. Figure 7 shows the layout of the sales area of the store, which measured 

41,912 ft2

. Table 6, Table 7, Table 8, and Table 9 describe the refrigerated fixtures at the 

multiplex test store. 

Figure 7. Layout of the Multiplex Refrigeration Test Store 

39 
Table 6. Low Temperature Circuits at the Multiplex Refrigeration Store (Rack A) 

Circuit Fixture Description

Refrig Load 

(Btuh) 

Evap 

Temp

(°F) 

Disch Air 

Temp

(°F) 

A1 Reach-in Ice Cream/Frozen Food Case 28,050 -19 -12 

A2 Reach-in Ice Cream/Frozen Food Case 28,050 -19 -12 

A3 Reach-in Ice Cream/Frozen Food Case 29,700 -19 -12 

A4 Reach-in Ice Cream/Frozen Food Case 28,050 -19 -12 

A5 Reach-in Ice Cream/Frozen Food Case 26,400 -19 -12 

A6 Reach-in Ice Cream/Frozen Food Case 24,750 -19 -12 

A7 Reach-in Ice Cream/Frozen Food Case 28,050 -19 -12 

A8 Reach-in Ice Cream/Frozen Food Case 33,000 -19 -12 

A9 Wide Island Frozen Fish/Meat Case 17,050 -20 -12 

A10 Wide Island Frozen Fish/Meat Case 13,950 -20 -12 

A11 Deli Freezer 8,200 -15 -5

A12 Bakery Freezer 10,000 -15 -5

A13 Grocery Freezer 35,000 -20 -10

A14 Reach-in Ice Cream/Frozen Food Case 4,950 -19 -12

 Total Refrigeration Load (Btuh) 315,200 

Low Temp Rack A SST = -25°F

Table 7. Medium Temperature Circuits at the Multiplex Refrigeration Store (Rack B) 

Circuit Fixture Description

Refrig Load 

(Btuh) 

Evap 

Temp

(°F) 

Disch Air 

Temp 

(°F) 

B2 Deli Walk-in Cooler 5,800 20 36

B4 Sandwich Case 5,600 18 28

B1 Deli Island Case 14,400 15 25

B3 Cheese Island Case 14,400 15 25

B5 Service Deli Cases 7,040 15 28

B6 Multi-deck Deli Cases 22,800 24 32

B7 Bakery Island Cases 9,600 15 25

B8 Retarder Box 8,000 20 36

B9 Self-Service Bakery Case 26,800 20 34

B10 Multi-deck Dairy Case 17,100 24 32

B18 Produce Island Case 14,400 15 25

B11 Multi-deck Meat Cases 45,720 21 29

B12 Single-deck Meat Cases 12,960 21 28

B13 Self-Service Fish Case 5,448 19 28

B14 Service Fish Cases 2,400 15

B15 Wide Island Meat Cases 22,680 18 26

B16 Multi-deck Lunch Meat Cases 39,900 24 32

B17 Multi-deck Lunch Meat Cases 11,400 24 32

 Total Refrigeration Load (Btuh) 286,448 

Medium Temp Rack B SST = 13°F

40 
Table 8. Medium Temperature Circuits at the Multiplex Refrigeration Store (Rack C) 

Circuit Fixture Description

Refrig Load 

(Btuh) 

Evap 

Temp 

(°F) 

Disch Air 

Temp 

(°F) 

C1 Multi-deck Dairy/Beverage Cases 74,100 24 32 

C2 Multi-deck Beverage Cases 108,300 24 32 

C3 Multi-deck Produce Cases 77,600 21 31 

C4 Multi-deck Produce Cases 42,750 24 32 

C5 Single-deck Produce Cases 46,560 21 31 

C6 Floral Walk-in Cooler 8,100 20 38 

C11 Floral Display Case 12,200 20 38 

C7 Produce Walk-in Cooler 39,120 20 36 

C8 Seafood Walk-in Cooler 9,700 20 36 

C9 Dairy Walk-in Cooler with 16 Doors 50,000 20 32 

C10 Meat Walk-in Cooler 31,400 20 32 

 Total Refrigeration Load (Btuh) 499,830 

Medium Temp Rack C SST = 20°F

Table 9. Medium Temperature Circuits at the Multiplex Refrigeration Store (Rack D) 

Circuit Fixture Description

Refrig Load 

(Btuh) 

Evap 

Temp 

(°F) 

Disch Air 

Temp 

(°F) 

D1 Produce Prep Area 24,000 37 50

D2 Meat Prep Area 95,000 37 50

 Total Refrigeration Load (Btuh) 119,000 

 

 Subcooling System A 49,950 

 Subcooling System B 44,708 

 Subcooling System C 81,634 

 Total Subcooling Load (Btuh) 176,292 

Medium Temp Rack D SST = 35°F

Figure 8 and Figure 9 show diagrams of the multiplex refrigeration system at the test 

store, for the low and medium temperature refrigeration, respectively. The refrigeration 

was divided over three compressor racks, consisting of one low temperature and two 

medium temperature racks. One of the medium temperature racks (labeled as C/D in 

the diagram) has two suction pressure groups. The highest temperature suction group, 

D, was also used for mechanical subcooling of the remaining low and medium 

temperature racks. The arrangement is considered unusual in that normally only the 

low temperature rack is mechanically subcooled. An evaporative condenser provided 

heat rejection. The refrigeration system was equipped to provide heat reclaim for both 

hot water and space heating. 

41 
Figure 8. Piping Diagram for the Low Temperature Multiplex Refrigeration 

Figure 9. Piping Diagram for the Medium Temperature Multiplex Refrigeration 

42 
2.3.4 Measurement Plan for the Multiplex Refrigeration System 

A measurement plan was also developed for the multiplex system that included all key 

parameters to allow comparison of performance with that of the secondary loop 

refrigeration. The elements needed for this comparison include the following: 

• Refrigeration system energy consumption 

• Refrigeration supplied 

• Operating state points 

• Display case operating temperatures 

• Ambient temperature and humidity 

2.3.4.1 Refrigeration Energy Consumption 

Power measurements of refrigeration compressors, and condenser fans and pumps were 

recorded from watt transducers at regular intervals. The measurements were used to 

determine energy consumption and for the multiplex refrigeration system EER 

calculation. 

The total power and energy of the store was also monitored. 

Refrigeration Supplied 

The refrigeration load delivered by each refrigeration rack was found from: 

( ) = m hout − hliq

•

Qrack

where 

h is the enthalpy of the refrigerant at the suction of the compressor rack (out) and 

at the liquid manifold (liq). 

m is the mass flow rate of refrigerant circulated by the compressors. 

The refrigeration provided was also estimated through the published capacity data for 

the compressors and digital on/off readings that described compressor operation as 

explained above for the secondary loop compressors. Several of the medium 

temperature compressors employed cylinder unloading for capacity control. Operation 

of the unloaders was monitored and counted by digital on/off readings. 

2.3.4.2 Refrigeration System Operating State Points 

The significant operating state points that were monitored for each multiplex 

compressor rack and for each compressor suction group are the following: 

• Saturated Suction Temperature (SST) – the saturation temperature corresponding to 

the refrigerant pressure measured at the compressor suction. 

43 
• Saturated Discharge Temperature (SDT) – The saturation temperature of the 

refrigerant based upon the pressure measured at the compressor discharge. 

• Return gas temperature – The temperature of the refrigerant gas measured at the 

suction of the compressor. 

• Refrigerant liquid temperature – The liquid temperature was measured at the outlet 

of the receiver and before and after each mechanical subcooling heat exchangers to 

characterize the subcooling provided. 

• Refrigerant discharge temperature – Measured at the inlets of the heat reclaim water 

heater (an outlet temperature was also required) and space heater coils. 

• Condenser pan water temperature – the primary control point for the condenser 

fans. 

• On/off digital signals were used to track events such as the initiation of defrost or 

heat reclaim. 

2.3.4.3 Display Case Operation 

The operation of the display cases and walk-in boxes were monitored during testing. 

Each display case is equipped with a discharge air temperature reading used to control 

operation of the refrigeration system. 

Several display cases were equipped with product simulators, similar to those used with 

the secondary loop system, so an evaluation of product temperature maintenance could 

be made. The display cases equipped with product simulators were 

• Single-deck meat – 2 display cases 

• Multi-deck produce – 2 display cases 

The display cases selected for product simulators were similar in type and operated at 

the same discharge air temperature as those cases instrumented at the secondary loop 

store. 

Other case measurements that were addressed included power readings for case lights, 

fans, and anti-sweat heaters (frozen food case doors). 

2.3.4.4 Ambient Conditions 

Inside ambient conditions were monitored with several dry-bulb temperature 

measurements and one relative humidity reading. The outside ambient conditions were 

determined through dry-bulb temperature and relative humidity readings taken at a 

Southern California Edison weather station also located in Valencia, CA. 

2.3.5 Data Collection at the Multiplex Refrigeration Test Store 

All instrumentation used for the evaluation of the multiplex refrigeration system was 

wired to the store EMS, where readings were taken at two-minute intervals and stored. 

The EMS was polled on a daily basis through a modem and phone line to gather the 

data. The weather station data were collected routinely by Southern California Edison 

and provided to the project on a monthly basis.

44 
45 

3.0 Project Outcomes 

3.1 Analysis Results 

The analytical models were used to predict the annual energy consumptions of the 

secondary loop and multiplex refrigeration systems. The design refrigeration loads for 

the secondary loop store were used to estimate the energy consumptions of both 

systems. Weather data for Moorpark, CA for 2001 were used to generate the ambient 

temperature bins for the analysis. 

Table 10 provides the annual energy estimates for the secondary loop refrigeration 

system. The modeled secondary loop system included evaporative condensing, 

multiple parallel secondary fluid pumping, and warm brine defrost subcooling to help 

decrease energy consumption. Mechanical subcooling was modeled for the low 

temperature refrigeration. Table 11 has the energy estimates for the multiplex system 

with air-cooled condensing, while Table 12 shows the energy consumption for the 

multiplex system with evaporative condensing. The multiplex refrigeration system 

employed mechanical subcooling for the low temperature refrigeration only. 

Table 10. Secondary Loop Modeling Results 

System Brine 

Loop 

Temp 

Design Load Annual Energy (kWh) 

 °F Btuh Compressors Pumping Condenser Total 

Low Temp -20 281,300 263,951 19,270 31,863 315,084 

Med Temp 1 20 337,800 169,977 22,406 39,938 232,321 

Med Temp 2 25 446,300 214,914 25,598 45,628 286,140 

Subcooling 49,750 

Total 648,842 67,328 117,429 833,599 

Table 11. Multiplex Modeling Results – Air-Cooled Condensing 

System SST Design Load Annual Energy (kWh) 

 °F Btuh Compressors Condenser Total 

Low Temp -23 281,300 304,235 70,418 374,653 

Med Temp 1 13 337,800 216,580 25,658 242,238 

Med Temp 2 19 446,300 270,676 91,906 362,582 

Subcool 49,750 

Total 791,491 187,982 979,473 
46 

Table 12. Multiplex Modeling Results – Evaporative Condensing 

System SST Design Load Annual Energy (kWh) 

 °F Btuh Compressors Condenser Total 

Low Temp -23 281,300 290,823 29,965 320,788 

Med Temp 1 13 337,800 186,366 21,836 208,202 

Med Temp 2 19 446,300 231,637 79,102 310,739 

Subcool 49,750 

Total 708,826 130,903 839,729 

The results of the analysis show that the secondary loop refrigeration system had lower 

energy consumption than either of the multiplex system configurations. Annual energy 

savings achieved by the secondary loop system versus the multiplex system with aircooled condensing were 145,874 kWh, which were savings of 14.9%. Versus multiplex 

with evaporative condensing, the annual savings were 6,130 kWh, or 0.3% 

3.2 Field Test Results 

3.2.1 Energy Comparison 

Summaries of the electric energy data collected at the two sites are given in Table 13 and 

Table 14 for the multiplex and secondary loop stores, respectively. Values are listed for 

each month of testing and are expressed as average daily energy consumptions. A 

breakdown is provided for the refrigeration system in terms of the rack (compressors 

only for the multiplex, and compressors and pumps for the secondary loop), the 

condenser, and the display cases. The display case energy includes the energy 

consumptions for fans, lights, and anti-sweat heaters. No values for store or display 

case energy were recorded for the secondary loop store during January, because of 

problems incurred with the data acquisition. Figure 10 shows the average daily energy 

values for the store and for total refrigeration for both the multiplex and secondary loop 

sites. The secondary loop refrigeration system represents a larger percentage of the total 

store energy, because the secondary loop site is smaller than the multiplex site and has 

lower total energy use. 
47 

Table 13. Multiplex Refrigeration Average Daily Energy Consumption 

 Average Daily Energy (kWh/day) 

Month Compressors Condenser Cases Refrigeration Store % of Store 

Oct '02 1,994 189 726 2,909 6,530 44.5 

Nov 1,871 157 708 2,736 6,609 41.4 

Dec 1,920 132 961 3,013 7,021 42.9 

Jan '03 2,039 157 799 2,995 7,066 42.4 

Feb 1,989 126 764 2,879 6,851 42.0 

Mar 2,030 144 732 2,906 6,396 45.4 

Apr 2,072 131 739 2,942 6,471 45.5 

May 2,016 302 760 3,078 6,446 47.8 

June 2,183 297 781 3,261 6,822 47.8 

 

Overall 2,013 182 774 2,969 6,690 44.4 

Table 14. Secondary Loop Refrigeration Average Daily Energy Consumption 

 Average Daily Energy (kWh/day) 

Month Compressors Pumps Condenser Cases Refrigeration Store % of Store 

Oct '02 2,069 191 299 1,140 3,699 6,265 59.0 

Nov 1,952 187 253 1,103 3,495 6,134 57.0 

Dec 1,901 188 199 1,113 3,401 6,034 56.4 

Jan '03 1,937 201 249 NR 2,387 NR - 

Feb 1,853 213 208 1,116 3,390 5,968 56.8 

Mar 2,035 204 243 1,083 3,565 6,131 58.1 

Apr 2,038 195 243 1,127 3,603 6,147 58.6 

May 2,151 195 309 1,061 3,716 6,281 59.2 

June 2,303 203 316 988 3,810 6,369 59.8 

 

Overall 2,027 197 258 1,091 3,452 6,166 58.0 
0

1,000

2,000

3,000

4,000

5,000

6,000

7,000

8,000

Oct '02

Nov

Dec

Jan '03

Feb

Mar

Apr

May

June

Average Daily Energy (kW

h/day)

Secondary Loop

Refrigeration

Secondary Loop Store

Multiplex Refrigeration

Multiplex Store

Figure 10. Store and Total Refrigeration Energy use for the Multiplex and Secondary Loop 

Test Sites 

Figure 11 and Figure 12 show the breakdowns of energy use for refrigeration for the 

multiplex and secondary loop systems. Compressor energy was the largest component 

at each site followed by display case energy. The compressor energy at the secondary 

loop site was approximately the same as the multiplex system. Both display case and 

condenser energy was higher at the secondary loop site. One reason for the higher case 

energy use was the display case lights were turned off at the multiplex site during late 

evening hours; this practice was not followed at the secondary loop store. The 

condenser energy use at the multiplex store was lower than at the secondary loop store. 

The multiplex condenser was sized to handle heat rejection for both the refrigeration 

and store air conditioning, which meant the condenser was over-sized for refrigeration 

operation alone. The air conditioning ran only briefly during the test period, meaning 

minimum fan operation was required to maintain the condenser minimum set point 

temperature. 

48 
0

500

1,000

1,500

2,000

2,500

Oct '02

Nov

Dec

Jan '03

Feb

Mar

Apr

May

June

Avera

ge Daily Energ

y (

k

W

h/day)

Compressors

Condenser

Cases

Figure 11. Multiplex Refrigeration System Energy Consumption 

0

500

1,000

1,500

2,000

2,500

Oct '02

Nov

Dec

Jan '03

FebMar

Apr

May

June

A

verage D

aily E

nerg

y (

k

W

h/day)

Compressors

Pumps

Condenser

Cases

Figure 12. Secondary Loop Refrigeration Energy Consumption 

49 
50 

One of the major energy differences between the multiplex and secondary loop 

refrigeration systems was the operation of the secondary loop pumps. To minimize the 

pump energy, multiple parallel pumps were installed on each fluid loop; these pumps 

were multiplexed based upon the pressure difference across the pumps. Table 15 shows 

the overall results achieved by this arrangement. The pump run fractions are shown for 

each of the fluid loops, along with the average daily pump energy consumed. 

Maximum daily pump energy was then calculated, which represents the energy 

consumed if all pumps ran continuously. This maximum consumption would occur if 

the fluid loops were operated at a constant flow rate at all times. The energy savings are 

the difference between the maximum and average consumptions. The use of the 

multiple pump arrangement resulted in an energy savings of 58%. The savings of 273.2 

kWh/day are 7.9% of the energy consumption of the secondary loop refrigeration 

system. 

Table 15. Performance of Multiple Pump Arrangement for Secondary Fluid Pumping 

Pump Run Fraction Pump Energy (kWh/day) 

Fluid 

Loop 

Pump 1 Pump 2 Pump 3 Avg. Pump 

Energy 

Max. Pump 

Energy 

Savings 

Low Temp 0.99 0.42 0.00 83.3 177.2 93.9 

Med. 

Temp 1 

0.99 0.33 0.05 60.0 131.4 71.4 

Med. 

Temp 2 

0.99 0.01 0.00 54.0 161.9 107.9 

Overall 197.3 470.5 273.2 

3.2.2 Power Demand 

Table 16 and Table 17 give the peak power demand (15 min.) recorded for each test 

month for the multiplex and secondary loop stores, respectively. The day and time of 

the peak are also noted in the tables. The power for the total refrigeration that occurred 

at the same time is also shown, along with the contribution of the refrigeration to the 

peak demand. No store demand was measured for the secondary loop store during 

January because of data acquisition problems. Figure 13 shows a graphical 

representation of these power values. The peak power demand of the multiplex store 

was higher than the secondary loop site 
Table 16. Peak Power Demand of the Multiplex Store 

 Date Time Store Refrig 

% of 

Store 

Oct '02 9-Oct 16:55 370.3 135.2 46.2

Nov 9-Nov 15:39 329.3 118.6 47.3

Dec 16-Dec 15:47 348.6 102.3 41.5

Jan '03 16-Jan 8:31 335.2 102.2 42.9

Feb 11-Feb 11:13 351.9 117.7 43.4

Mar 14-Mar 14:29 324.3 126.2 50.2

Apr 4-Apr 9:50 322.7 100.1 41.5

May 28-May 14:37 332.7 133.5 50.9

Jun 16-Jun 14:22 341.9 129.3 48.5

Table 17. Peak Power Demand for the Secondary Loop Store 

 Date Time Store Refrig 

% of 

Store 

Oct '02 6-Oct 16:30 312 111.2 51.5 

Nov 9-Nov 17:15 312 124.8 56.7

Dec 4-Dec 17:15 304 124.8 56.6

Jan '03 NR 

Feb 13-Feb 9:45 296 120 57.6

Mar 31-Mar 12:00 320 130.4 56.5

Apr 16-Apr 13:30 320 146.4 62

May 19-May 12:00 320 136.8 57.3

Jun 16-Jun 16:00 320 150.4 60.5

0

50

100

150

200

250

300

350

400

Oct '02

Nov

Dec

Jan '03

Feb

Mar

Apr

May

Jun

Power De

m

and - 1

5 min (

kW)

Multiplex - Store

Multiplex - Total

Refrigeration

Secondary Loop -

Store

Secondary Loop - Total

Refrigeration

Figure 13. Peak Power Demand for the Multiplex and Secondary Loop Stores 

51 
3.2.3 Refrigeration Efficiency

Table 18 and Table 19 show the average refrigeration loads, power, and EER for the 

multiplex and secondary loop systems, respectively. The refrigeration load at each site 

has been divided into two segments which are the low and medium temperature 

refrigeration loads. The low temperature refrigeration load is addressed by the low 

temperature system at each site. The remainder of the refrigeration load was combined 

as the medium temperature load. The compressor power was also divided on this basis. 

The only exception is the mechanical subcooling load, which is subtracted from the 

medium temperature refrigeration. 

For the multiplex refrigeration, the power and EER were also determined if no 

mechanical subcooling were applied to the medium temperature refrigeration. 

Mechanical subcooling is not normally used with medium temperature refrigeration and 

would not be representative of a baseline multiplex system found in most supermarkets. 

An estimate of the power level expected if the medium temperature racks were not 

subcooled is also given to provide an additional comparison with the secondary loop 

system. The estimate performance for non-subcooled compressors was made by first 

determining the total refrigeration capacity of the medium temperature compressors 

using the measured saturated suction and discharge temperatures along with the 

measured return gas and liquid temperatures, and the performance curves for the 

compressors. The total run fraction for the compressors was estimated from this 

capacity and the refrigeration load measured at the site: 

∑

= Cap

NoSubRunFraction Qref

where 

Qref is the measured refrigeration load 

Σ Cap is the total non-subcooled capacity of the compressors 

The estimated power of the compressors is then found by multiplying the combined 

power for the compressors with the NoSubRunFraction. 

52 
53 

Table 18. Multiplex Refrigeration 

 Low Temperature Medium Temperature Med Temp No Subcooling 

 

Load 

(kBtuh) 

Rack 

Power 

(kW) 

EER 

(Btu/W-hr) 

Load 

(kBtuh) 

Rack 

Power 

(kW) 

EER 

(Btu/W-hr) 

Rack 

Power 

(kW) 

EER 

(Btu/W-hr) 

Oct '02 323.2 42.2 7.65 554.6 44.0 12.60 46.2 12.00 

Nov 316.7 40.9 7.73 508.3 39.5 12.87 41.4 12.28 

Dec 344.6 43.4 7.94 479.1 36.7 13.06 38.3 12.51 

Jan '03 333.2 43.1 7.73 540.3 41.9 12.90 44.0 12.28 

Feb 321.9 41.6 7.74 526.1 41.0 12.82 43.2 12.17 

Mar 328.3 42.6 7.71 533.9 42.4 12.58 44.6 11.96 

Apr 336.5 43.0 7.81 536.9 43.3 12.39 45.5 11.79 

May 336.3 43.2 7.80 514.4 44.3 11.61 46.8 11.00 

Jun 337.5 43.4 7.78 582.3 47.3 12.30 49.3 11.81 

 

Overall 331.5 42.7 7.76 530.1 42.2 12.60 44.3 12.00 

Table 19. Secondary Loop Refrigeration 

Low Temperature Medium Temperature 

Load 

(kBtuh) 

Rack 

Power 

(kW) 

EER 

(Btu/W-hr) 

Load 

(kBtuh) 

Rack 

Power 

(kW) 

EER 

(Btu/W-hr) 

Oct '02 350.9 42.9 8.18 587.6 47.2 12.47 

Nov 348.8 41.9 8.32 560.6 45.3 12.39 

Dec 350.2 41.7 8.40 546.0 45.5 12.00 

Jan '03 349.1 41.7 8.37 585.8 47.1 12.43 

Feb 340.0 40.6 8.37 568.7 45.4 12.51 

Mar 346.6 41.7 8.31 648.9 51.6 12.57 

Apr 338.2 41.1 8.23 653.8 51.9 12.62 

May 338.5 41.5 8.16 699.8 56.2 12.51 

Jun 343.3 42.7 8.04 751.8 61.8 12.18 

Overall 344.8 41.7 8.27 617.0 49.8 12.42 

The EER values for the low temperature refrigeration show the secondary loop system 

operating at an approximately 6.6% higher efficiency than the multiplex system. For 

medium temperature refrigeration, the multiplex system EER was higher than that of 

the secondary loop system by 1.4%; but, when the multiplex system does not use 

mechanical subcooling for medium temperature, the EER of the secondary loop system 

is higher by 5%. 

Further normalization of the refrigeration efficiency data can be achieved by considering 

the ambient wet-bulb temperature variation, since both refrigeration systems reject heat 

to an evaporative heat exchanger. Figure 14 shows the plot for the EER versus the 

ambient wet-bulb temperature for low temperature refrigeration. Figure 15 and Figure 
54 

16 show this same relation for medium temperature refrigeration for the multiplex and 

secondary loop systems, respectively. A linear regression was performed for each plot. 

The normalized EER data can be used to estimate the energy consumption for each 

refrigeration system when addressing the same refrigeration loads and when operating 

at the same ambient wet-bulb temperature. For the multiplex store, the overall average 

refrigeration loads were 331,500 and 530,100 Btu/hr for low and medium temperature 

refrigeration, respectively, and the average ambient wet-bulb temperature at the 

multiplex site was 49.2°F. For the secondary loop store the refrigeration loads were 

344,800 and 617,000 Btu/hr for the low and medium temperature, respectively, and the 

average ambient wet-bulb temperature was 50.3°F. Using these values, and the energy 

consumption of the secondary loop and multiplex refrigeration was estimated for each 

of the test stores. The results of this comparison are given in Table 20 and Table 21 for 

the multiplex and secondary loop stores, respectively. The secondary loop refrigeration 

used less total energy than the multiplex with either subcooled or no subcooled medium 

temperature refrigeration for both site descriptions. 

Table 20. Comparison of Secondary Loop and Multiplex Refrigeration Using Normalized 

EER Values, Multiplex Store Refrigeration Loads and Ambient Wet-bulb Temperatures 

EER (Btu/W-hr) Energy (kWh/day) Savings (kWh/day) 

Low Temperature Refrigeration 

Secondary Loop 8.29 959.7 

Multiplex 7.76 1,025.2 65.5 (6.4%) 

Medium Temperature Refrigeration 

Secondary Loop 12.44 1,022.7 

Multiplex 12.60 1,009.7 -13.0 (-1.3%) 

Multiplex – No 

Subcooling 

12.01 1,059.3 36.6 (3.5%) 

Total Refrigeration 

Secondary Loop 1982.4 

Multiplex 2034.9 52.5 (2.6%) 

Multiplex – No 

Subcooling 

2084.4 102.1 (4.9%) 

Low Temperature Load – 331,500 Btu/hr 

Medium Temperature Load – 530,100 Btu/hr 

Average Ambient Wet-Bulb Temperature – 49.2°F 
Table 21. Comparison of Secondary Loop and Multiplex Refrigeration Using Normalized 

EER Values, Secondary Loop Store Refrigeration Loads and Ambient Wet-bulb 

Temperatures 

EER (Btu/W-hr) Energy (kWh/day) Savings (kWh/day) 

Low Temperature Refrigeration 

Secondary Loop 8.26 1,001.2 

Multiplex 7.58 1,066.8 65.6 (6.1%)

Medium Temperature Refrigeration 

Secondary Loop 12.41 1,192.9 

Multiplex 12.54 1,181.1 -11.8 (-1.0%)

Multiplex – No 

Subcooling 

11.94 1,240.1 47.2 (3.8%)

Total Refrigeration 

Secondary Loop 2,194.1 

Multiplex 2,247.9 53.8 (2.4%)

Multiplex – No 

Subcooling 

2,306.9 112.8 (4.9%)

Low Temperature Load – 340,800 Btu/hr 

Medium Temperature Load – 617,000 Btu/hr 

Average Ambient Wet-Bulb Temperature – 50.3°F

y = -0.0066x + 8.0887

R2

 = 0.0506

y = -0.0261x + 9.5784

R2

 = 0.5015

5.00

5.50

6.00

6.50

7.00

7.50

8.00

8.50

9.00

30 40 50 60 70

Ambient Wet-Bulb Temperature (F)

System EER (Btu/W-hr)

Secondary Loop

Multiplex

Secondary

L

Multiplex

Figure 14. Relation between Ambient Wet-Bulb Temperature and Low Pressure 

Refrigeration EER 

55 
y = -0.0597x + 15.538

R2

 = 0.2474

y = -0.0599x + 14.954

R2 8.00 = 0.2458

9.00

10.00

11.00

12.00

13.00

14.00

15.00

16.00

17.00

30 40 50 60 70

Ambient Wet-bulb 

Temperature (F)

System EER (Bt

u/W-hr)

Multiplex

Multiplex - No

Subcooling

Figure 15. Relation between Ambient Wet-Bulb Temperature and Medium Temperature 

Refrigeration EER for the Multiplex System 

y = -0.0248x + 13.661

R2

 = 0.0629

8.00

9.00

10.00

11.00

12.00

13.00

14.00

30 40 50 60 70

Ambient Wet-bulb

Temperature (F)

Syste

m E

E

R (Btu/W-hr)

Secondary Loop

Figure 16. Relation between Ambient Wet-Bulb Temperature and Medium Temperature 

Refrigeration EER for the Secondary Loop System 

56 
57 

3.2.4 Product Storage Temperatures 

The ability of the two refrigeration systems to maintain product storage temperatures 

was evaluated. Two display case types were considered, which were: 

• Single-deck meat 

• Multi-deck produce 

Cases of these types were available at the two test stores. These cases were equipped 

with product simulators that allowed the temperature of a thermal mass to be 

monitored. 

Table 22 shows the average temperature measurements for the instrumented cases for a 

24-hour period. The uniformity of the product temperature is characterized by the 

standard deviation (σ) of the product temperature. By definition, 95% of the product 

temperature values will fall within +/- 2σ. Temperature plots for the meat cases are 

shown in Figure 17 and Figure 18, for the multiplex and secondary loop stores, 

respectively. Similar plots for the produce cases are shown in Figure 19 and Figure 20. 

Table 22. Comparison of Display Case Performance for the Multiplex and Secondary Loop 

Refrigeration Systems 

Compressor Rack 

SST 

Discharge Air 

Temp 

Product Temp Standard Deviation 

in Product Temp 

Single-deck Meat Display Case 

Multiplex 2.3 26.5 30.7 +/- 0.95 

Secondary Loop 14.1 29.4 33.8 +/- 1.81 

Multi-deck Produce Case 

Multiplex 21.0 36.0 37.7 +/- 1.35 

Secondary Loop 20.9 37.4 37.1 +/- 0.88 

For the single-deck meat cases, the product temperatures achieved by both systems are 

acceptable for storage. Both cases also operate at a similar temperature difference 

between the product and the discharge air – 4.2°F for multiplex and 4.3°F for the 

secondary loop. The multiplex system is operated at a significantly lower SST than the 

secondary loop system. It should also be noted from Figure 16 and Figure 17 the peak 

discharge air temperature of the secondary loop case is higher than the multiplex case. 

The maximum discharge air temperature for the secondary loop case was 56.4°F, while 

the maximum for the multiplex case was 42.4°F. The higher air temperature for the 

secondary loop case suggests that the defrost duration used was too long, causing the air 

and the product to warm up unnecessarily. With a proper defrost time setting, a more 

uniform product temperature can be expected. 

The temperature results for the multi-deck produce cases were similar for both the 

multiplex and secondary loop systems. The SST of both racks was essentially the same. 

The temperature difference between the discharge air and the product was small for 

both systems. For the secondary loop case, the average discharge air temperature was 

slightly higher than the product temperature, which may be explained by an error in the 

temperature measurements and the air temperature includes values during defrost. The 

uniformity of the product temperature was better for the secondary loop produce case 

by a small amount. 
Multiplex Refrigeration - Single-deck 

Meat Case

0

5

10

15

20

25

30

35

40

45

02479

1214171922

Hour of Day

Temperature (F)

Case Dishcarge

Air

Product

Figure 17. Discharge Air and Product Temperature Profiles for the Single-Deck Meat Case - 

Multiplex Refrigeration 

Secondary Loop - Single -deck Meat 

Display Case (B1-6a)

0

10

20

30

40

50

60

02479

1214171922

Hour of Day

Temperature (

F)

Case Discharge Air

Product

Figure 18. Discharge Air and Product Temperature Profiles for the Single-Deck Meat Case - 

Secondary Loop Refrigeration 

58 
Multiplex Refrigeration - Multi-deck 

Produce Cases

0

10

20

30

40

50

60

02469

111316182023

Hour of Day

Temperature (F)

Discharge Air

Product 

Figure 19. Discharge Air and Product Temperature Profiles for the Multi-Deck Product 

Case – Multiplex Refrigeration 

Secondary Loop - Multi-deck Produce 

Case (B2-7b)

0

10

20

30

40

50

02479

1214171922

Hour of Day

Temperature (

F)

Case Discharge Air

Product

Figure 20. Discharge Air and Product Temperature Profiles for the Multi-Deck Product 

Case – Secondary Loop Refrigeration 

59 
3.2.5 Refrigerant Leakage Comparison 

The refrigerant charge for the secondary loop refrigeration system was 1400 lb. This 

value is higher than the project goal of 500 lb or less. The added charge is primarily due 

to the of the additional refrigerant necessary to provide heat reclaim for both hot water 

and space heat. Heat reclaim is an effective use of the discharge heat generated by the 

refrigeration system which can replace a significant amount of energy in a supermarket. 

The refrigerant charge for the multiplex system was not measured, but was estimated at 

3,000 lb, or roughly 1,000 lb. per compressor rack. 

Table 23 provides a record of the refrigerant additions at the two test stores. The 

information was obtained from Safeway, Inc for both stores. The refrigerant additions 

for the secondary loop store include 2, 100-lb additions that occurred in September 2002. 

These charges were required because of installation of flow meters in the refrigeration 

used for the field test work and should not be considered normal refrigerant loss by the 

system. The remaining 275 lb addition was due to a fitting break at the compressor rack. 

This one loss is believed to be the only recharge required by the secondary loop system 

since its start on December 18, 2001. 

The records for the refrigerant charge addition for the multiplex store are very limited, 

because of a change in site maintenance contractor as of January 2003. Refrigerant 

records from the previous contractor were not available from Safeway, Inc. at the time of 

this report. The available information is very limited and may not be representative of 

the true refrigerant leakage of this multiplex refrigeration system. The total charge 

addition for the first four months of 2003 listed for the multiplex site is 72 lb. 

Table 23. Refrigerant Additions at the Secondary Loop and Multiplex Stores 

Vons 1610 -Secondary Loop Store Vons 2030 - Multiplex Store 

Date Refrigerant Pounds Date Refrigerant Pounds

9/1/2002 507 100 2/21/2003 404A 8

9/4/2002 507 100 3/14/2003 R22 60

4/18/2003 507 275 4/25/2003 R404A 4

Total 475 72

The annual refrigerant leak rate from each site can be determined using the methods 

provided by the USEPA (9). The percent of total refrigerant charge that is lost through 

leakage is found from: 

100

dayssince last refrigerant additions

365 days

total refrigerant charge

lb refrigerant added % Annual Leak Rate = × × 

60 
61 

Table 24 shows the results for the two stores. With the exception of the addition of field 

test instrumentation, the only refrigerant charge addition credited to secondary loop 

store since the opening was a single recharge of 275 lb. The number of days between the 

store opening and the refrigerant addition was approximately 486 days. The limited 

records for the multiplex store suggest that the system was fully charged on February 

21. Between that date and April 25, 64 lb of refrigerant was needed to maintain the 

system charge. The calculated annual % leak rates for the 2 stores were very close at 14.8 

and 12.4 % for the secondary loop and multiplex stores, respectively, because of the 

large system charge associated with the multiplex refrigeration system. In terms of 

absolute refrigerant loss, the annual leak rate for the secondary loop store was less than 

that of the multiplex store by 44.3%. 

Table 24. Estimated Annual Refrigerant Leak Rates for the Secondary Loop and Multiplex 

Test Stores 

Store Refrigerant 

Added (lb) 

Days Since 

Last Addition 

Total System 

Charge (lb) 

% Annual Leak 

Rate 

Annual Leak 

Rate (lb/yr) 

Secondary 

Loop 

275 486 1,400 14.8 206 

Multiplex 64 63 3,000 12.4 370 
62 
63 

4.0 Conclusion & Recommendations 

The specific technical objectives of the project were to design and test a secondary loop 

refrigeration system, which: 

• Consumes approximately 14% less electricity than a state-of-the-art multiplex refrigeration 

system (baseline system) installed in a comparable store. 

The modeling results for the secondary loop refrigeration system configured for this 

project show annual energy savings of 145,894 kWh, or 14.9%, can be obtained when 

compared to the performance of a multiplex refrigeration system employing air-cooled 

condensing. This particular multiplex system configuration is the most common 

configuration installed in supermarkets and may be considered the baseline system. 

Modeling results also show that the savings were reduced to 6,130 kWh/yr when the 

secondary loop was compared to multiplex with evaporative condensing, suggesting 

that the largest portion of the savings can be credited to the use of evaporative 

condensing. 

The field test results suggest the analytical results are in the correct direction. The 

savings achieved by the secondary loop refrigeration system versus the multiplex 

system with no medium temperature subcooling were 37,266 kWh/yr, or 4.9%. 

The savings achieved by the secondary loop refrigeration system may be attributed to 

energy-saving features incorporated in its design. The annual savings achieved by each 

of these features compared to a more conventional secondary loop refrigeration system 

were: 

• Multiple parallel brine pumps – Estimated annual savings versus single large 

pumps are 99,718 kWh. 

• Subcooling from warm brine defrost – Estimated annual savings versus no 

subcooling are 49,570 kWh. 

Additional savings were achieved by the use of a minimum delta T between the display 

case discharge air and refrigeration saturated suction temperatures, the use of a lowviscosity secondary fluid, and evaporative condensing. . 

• Has a refrigerant charge that is ten times less (less than 500 lbs.) than the baseline system. 

The advanced secondary loop refrigeration system as tested had an initial refrigerant 

charge of 1,400 lb., which is considerably larger than the stated goal. The reason for the 

larger refrigerant charge is the secondary loop system also had heat reclaim capabilities 

for both hot water and space heating. The added charge was needed for the additional 

piping and heat exchangers associated with heat reclaim. Heat reclaim is of great value 

to the operation of the supermarket since the heat reclaim can displace all energy use 

associated with hot water heating and space heating for the store. 
64 

• Loses annually no more than 15% of its refrigerant charge due to leakage. 

Service records for the two stores indicated the refrigerant leak rate for the secondary 

loop store was on the order of 14.8%/yr, or 206 lb/yr. The multiplex system was 

estimated to have a leak rate of 12.4%/yr, or 370 lb/yr. The refrigerant data for the 

multiplex store were very limited because the maintenance contractor for the store had 

changed at the beginning of 2003 and a complete year of service records was not 

available. It is likely that a full year of data would show a significantly higher loss of 

refrigerant for the multiplex refrigeration system. 

The ability of the two refrigeration systems to maintain product storage temperature 

was also assessed. The comparison for single-deck meat cases showed the multiplex 

and secondary loop systems maintained the product at acceptable temperature levels. 

The case associated with the multiplex system had a lower and more uniform product 

temperature than for the case operating in the secondary loop system. The multiplex 

display case had to operate a much lower rack SST in order to achieve this condition 

(2.3°F for multiplex vs. 14.1°F for secondary loop). For the multi-deck produce cases, the 

multiplex and secondary loop systems operated at similar rack SST values. The 

resulting average product temperature was approximately the same for both systems, 

but the product temperature of the multi-deck case in the secondary loop system was 

more uniform. 

Benefits to California 

This project contributed to the PIER program objective of reducing the environmental 

costs of California’s electrical system, by developing an alternative refrigeration system 

which uses significantly less refrigerant than conventional systems. It also contributed 

to the PIER program objective of improving energy value of California’s electricity by 

lowering electrical consumption of supermarket secondary loop refrigeration systems. 
5.0 References 

1. Sand, James R, Steven K. Fischer, Van D. Baxter, Energy and Global Warming 

Impacts of HFC Refrigerants and Emerging Technologies, Oak Ridge National Laboratory, 

sponsored by Alternative Fluorocarbons Environmental Acceptability Study (AFEAS), 

U.S. Department of Energy, 1997. 

2. Terrell, Wilson, Y. Mao, and P. Hrnjak, Tests of Supermarket Display Cases when 

Operating with Secondary Refrigerants, International Conference on Ozone Protection 

Technologies, Baltimore, MD, November 12-13, 1997. 

3. Cooper, P.J., Experience with Secondary Loop Refrigeration Systems in European 

Supermarkets, International Conference on Ozone Protection Technologies, Baltimore, 

MD, November 12-13, 1997. 

4. Walker, D.H., Development and Demonstration of an Advanced Supermarket 

Refrigeration/HVAC System, Foster-Miller, Inc., Subcontract Number 62X-SX363C Oak 

Ridge National Laboratory, Oak Ridge, TN, September 2001. 

5. Walker, D.H., and G.I. Deming, Supermarket Refrigeration Modeling and Field 

Demonstration, Foster-Miller, Inc., EPRI Report No. CU-6268, Electric Power Research 

Institute, Palo Alto, CA., March, 1989.

6. Kazachki, G., E. Bayoglu, and C. Gage, “Comparative Evaluation of HeatTransfer Fluids for Secondary Loop Systems,” presentation given at the International 

Conference on Ozone–Protection Technologies, Baltimore, MD, November 13, 1997. 

7. Walker, D.H., Foster-Miller, Inc., Field Testing of High-Efficiency Supermarket 

Refrigeration, EPRI Report No. TR-100351, Electric Power Research Institute, Palo Alto, 

CA., December, 1992. 

8. Walker, D.H., Foster-Miller, Inc., “Development of an Evaporative Condenser for 

Northern Climates,” Report prepared for Niagara Blower Co., Buffalo, NY, April, 1997. 

9. Compliance Guidance For Industrial Process Refrigeration Leak Repair Regulations 

Under Section 608 of The Clean Air Act, The Chemical Manufacturers Association and The 

Environmental Protection Agency, October 1995. 

65