# 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.CONSULTANTREPORT

March 2004 500-04-013Prepared 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

Dav i d H. Walker, P.E. Principal Investigator

Contract No. 500-98-039

Prepared For: 

## California Energy Commission

Chris Scruton, Project M a nager
Nancy Jenkins, Program Area Lead Buildings End-Use Energy Efficiency

Terry Surles, PIER Program M a nager

Robert L. Therkelsen, Executive Director



### DISCLAIMER

This report w a s prepared as the result of w o rk sponsored by the California Energy Commission. It does not necessarily represent the view s of the Energy Commission, its employ ees or the State of California. The Energy Commission, the State of California, its employ ees, contractors and subcontractors make no w a rrant,
ex press or implied, and assume no legal liability for the information in this report; nor does any party represent that the uses of this information w ill not infringe upon privately ow ned 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 Scru ton and Nancy Jenkins of the California Energy Commission for their help and guidance th roughout this project. We would also like to recognize the outstanding work and co operation 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.

iiiii

## Preface

The Public Interest Energy Research (PIER) Program supports public interest energy research and development that will help im prove 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 Ca lifornia 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 individu als, businesses, utilities, and public or private research institutions. PIER funding efforts are focused on th e following six RD&D program areas:

• Buildings End-Use Energy Efficiency

• Industrial/Agricultural/Wate r 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 Inve stigation of Secondary Loop Supermarket Refrigeration Systems. This project co ntributes 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.

ivv

## Table of Contents

Prefac e

............................................................................................................................... ............. iv

Executive Summa ry

........................................................................................................................ 1

Abstract

............................................................................................................................... ............ 5

1.0 Introductio n

......................................................................................................................... 7

1.1 Present Supermarket Re frigeration Sy stems

................................................................... 7 1.2 Low-charge Refrigeration Sy stem 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 L oop Refrigeration S y stem

................................................... 15 2.1.1 Description o f the Secondary L oop Refrigeration S y st em

.................................... 15 2.1.2 High-Efficiency Features for Secondar y Loop Refrigeration

............................... 17 2.1.3 Display Cases Designed for Use with Secondar y Flui d

........................................ 18 2.1.4 High-efficiency Refrigeration Com p ress ors

.......................................................... 19 2.1.5 Close-Coupling of Com p ressors and Evaporator

.................................................. 19 2.1.6 Multiple Parallel Pum p s

........................................................................................ 19 2.1.7 Evaporative Heat R e jectio n

................................................................................... 20 2.1.8 Low Viscosity Secondar y Fluid

............................................................................ 20 2.1.9 Refrigerant Subcoolin g fro m Warm Brine Defrost

............................................... 22 2.2 Analy s is of t h e Secondary Loop and M u ltiplex Refrigeration S y st em s

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

.............................................................................. 23 2.2.2 Modeling of Secondary L o op Refrigeration

.......................................................... 27 2.3 Field Testing of the Second ary Lo op Ref r igeration S y s t em

......................................... 30 2.3.1 Description o f the Secondary L oop Refrigeration Test Site

.................................. 30 2.3.2 Measurement Plan to Moni tor Secondar y Loop Refrigeration

.............................. 35 2.3.2.1 Refrigeration Energ y Cons um ption

................................................................... 35 2.3.2.2 Refrigeration Supplied

....................................................................................... 36 2.3.2.3 Refrigeration Sy stem Oper ating State Points

.................................................... 37 2.3.2.4 Brine Pum p Operation

....................................................................................... 37 2.3.2.5 Brine Defrost Operation

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

..................................................................................... 38 2.3.2.7 Am bient Conditions

........................................................................................... 38 2.3.2.8 Data Collection at the Secondar y Lo op Store

.................................................... 39 2.3.3 Field Testing of the Multi pl ex Refrigeration S y stem

............................................ 39 2.3.3.1 Description of the Multiple x Refrigeration Test Site

........................................ 39 2.3.4 Measurement Plan for the Multiplex Refrigeration System

.................................. 43 2.3.4.1 Refrigeration Energ y Cons um ption

................................................................... 43 2.3.4.2 Refrigeration Sy stem Oper ating State Points

.................................................... 43 2.3.4.3 Display Case Operation

..................................................................................... 44 2.3.4.4 Am bient Conditions

........................................................................................... 44 2.3.5 Data Collecti on at the Mul tiplex Refrigeration Test Store

.................................... 44

3.0 Project Outcomes

............................................................................................................... 45

vi3.1 Analy s is Results

............................................................................................................ 45 3.2 Field Test Re sults

.......................................................................................................... 46 3.2.1 Energy Com p arison

............................................................................................... 46 3.2.2 Power De ma nd

...................................................................................................... 50 3.2.3 Refrigeration Efficiency

........................................................................................ 52 3.2.4 Product Stora g e Tem p eratu r es

............................................................................... 57 3.2.5 Refrigerant L eakage Co m p arison

.......................................................................... 60

4.0 Conclusion & Reco mmen d ations

...................................................................................... 63

5.0 Referen ces

......................................................................................................................... 65



## List of Tables

Table 1. Com p arison of Operating Tem p eratur es for Direct-Expansion and Secondary -Fluid Refrigerated Display Cases

................................................................................................... 19 Table 2. Transport Properties of Propy lene Gly c ol-Water and Organic Salt-Water (Dy n alene) at Operating Tem p eratures of Secondary L oop Refrigeration

.................................................. 21 Table 3. Low Tem p erature Circuits for the Secondary Loop Refrig e ration Test Store

................ 31 Table 4. Medium Temperature Circuits for the Second ary Lo op Ref r igeration Sto r e (Loop B1) 32 Table 5. Medium Temperature Circuits for the Second ary Lo op Ref r igeration Sto r e

.................. 33 Table 6. Low Tem p erature Circuits at the Multiplex R e frigeration Store (Rack A)

..................... 40 Table 7. Medium Temperature Circuits at the Multipl e x Refrigerati on Store (Ra c k B)

.............. 40 Table 8. Medium Temperature Circuits at the Multipl e x Refrigerati on Store (Ra c k C)

.............. 41 Table 9. Medium Temperature Circuits at the Multipl e x Refrigerati on Store (Ra c k D)

.............. 41 Table 10. Secondar y Lo op Modeling Re sults

............................................................................... 45 Table 11. Multiplex Model i ng R esults – Air-Cooled Condensing

................................................ 45 Table 12. Multiplex Model i ng Resu lts – Evaporative Condensing

............................................... 46 Table 13. Multiplex Refrigeration Average Daily Ener gy Consum ption

...................................... 47 Table 14. Secondar y Lo op Refrigeration Average Daily Energ y Co n s u m ption

........................... 47 Table 15. Performance of Multiple Pum p A rrangem e nt for Secondary Fl uid Pum p ing

............... 50 Table 16. Peak Power Demand of the Multiplex St ore

................................................................. 51 Table 17. Peak Power Demand for the Se condar y Loo p Store

..................................................... 51 Table 18. Multiplex Refrigeration

................................................................................................. 53 Table 19. Secondar y Lo op Refrigeration

...................................................................................... 53 Table 20. Com p arison of Secondary Loop and Multiplex Refrigeration Using Norm alized EER Values, Multiplex Store Refrigeration Lo ads and Am bient Wet-bulb Tem p eratures

........... 54 Table 21. Com p arison of Secondary Loop and Multiplex Refrigeration Using Norm alized EER Values, Seco ndary Loop Store Refrigera tion Loads and Am bient W e t-bulb Tem p eratures . 5 5 Table 22. Com p arison of Display Case Perform ance for the Multiplex and Secondary Loop Refrigeration Sy stem s

........................................................................................................... 57 Table 23. Refrigerant Additions at the S econdary L o o p and Mult ipl e x Stores

............................. 60 Table 24. Estim ated Annual Refrigerant Leak Ra tes for the Secondary Loop and Multiplex Test Stores

............................................................................................................................... ...... 61

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## List of Figures

Figure 1. La yout of Refrigerated Display Cases and Walk-in Cool ers in a Ty pical Supermark e t

.. 7 Figure 2. The Multiplex Sy stem is the Most Com m on Ty pe of Refrigeration Used in Supermarkets

........................................................................................................................... 8 Figure 3. Fl o w Diagram fo r a Secondary Loop Refrig e ration S y stem

.......................................... 16 Figure 4. La yout of t h e Secondar y Lo op Refrigeration Test Store

............................................... 30 Figure 5. Piping Diagram for the Low Tem p erature Secondary Loop Refrigeration Sy stem (Patent by Hill PHOENIX, Inc.)

............................................................................................ 34 Figure 6. Piping for the Medium Tem p erature Secondary Loop Refrigeration Sy stem (Patent by Hill PHOENIX, Inc.)

............................................................................................................. 35 Figure 7. La yout of t h e Multiplex Refrigeration Test Store

.......................................................... 39 Figure 8. Pi pi ng Diagram for the Low T e m p erature Multiplex Refrigeration

.............................. 42 Figure 9. Pi pi ng Diagram for the Medium T e m p erature Multiplex Refrigeration

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

............................................................................................................................... ........ 48 Figure 11. Multiplex Refrigeration S y st em Energy C onsum ption

................................................ 49 Figure 12. Secondar y Loo p Refrigeration Energ y Cons um ption

.................................................. 49 Figure 13. Peak Power Demand for the Multiplex and Secondar y Loop St ores

........................... 51 Figure 14. Relation between Am bient Wet-Bulb Tem p erature and Low Pressure Refrigeration EER

............................................................................................................................... ........ 55 Figure 15. Relation between Am bient Wet-Bulb Tem p erature and Medium Tem p erature Refrigeration EER for the Multiplex S y stem

........................................................................ 56 Figure 16. Relation between Am bient Wet-Bulb Tem p erature and Medium Tem p erature Refrigeration EER for the S econdary L o o p S y stem

.............................................................. 56 Figure 17. Discharge Air and Product Tem p erature Profiles for the Single-Deck Meat Case - Multiplex Refrigeration

......................................................................................................... 58 Figure 18. Discharge Air and Product Tem p erature Profiles for the Single-Deck Meat Case - Secondary L o op Refrigeration

.............................................................................................. 58 Figure 19. Discharge Air and Product Tem p erature Profiles for the Multi-Deck Product Case – Multiplex Refrigeration

......................................................................................................... 59 Figure 20. Discharge Air and Product Tem p erature Profiles for the Multi-Deck Product Case – Secondary L o op Refrigeration

.............................................................................................. 59

viiiix

## Executive Summary

Background The largest single use of energy in a supermar ket is refrigeration, accounting for half of total energy use, on the order of 2-3 million kWh annually for a 35,000 ft 2 store. Refrigeration is typically provided by dire ct ex pansion 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 la rge. 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 refriger ant leakage. On average, a supermarket refrigeration system can be ex pected 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 req u iring signific antly 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
approx imately 300 – 500 lb. The anticipated refr igerant leak rate for the secondary loop refrigeration system is much less than that of the direct-ex p ansion systems, because all refrigerant is contained within the chiller syst em. 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, consider ing both energy and refrigerant use. This would make the environmentally benign second ary loop system economically attractive.
Project Approach
Southern California Edison and its subcontrac tor, Foster-Miller, Inc., obtained support from Safeway, Inc. to provide a superm arket 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 aske d 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 sa me time, a second supermarket operated by Safeway was identified that employed a state-of-the-art multiplex refrigeration system. The multiplex system was also instrument ed for performance measurement and both sites were monitored for approx imately 9 months.

1Project Objectives and Results The specific technical objectives of the proj ect were to build and test a secondary loop refrigeration system, which:

• Consumes approximately 14% less electricity th an 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 effici ent than air cooled condensing. This was appropriate because the secondary loop sy stem also used evaporative condensing. Modeling results predicted annual savings of 6,130 kWh, or just under 1% for this comparison, but the actual savi ngs achieved by the secondary loop refrigeration system were 37,266 kWh/yr, or 4.9%. This suggests th e benefit of the secondary loop system as tested over an air cooled multiplex sy stem 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 de sign. The annual savings achieved by each of these features compared to a more conven tional secondary loop refrigeration system

were:

• Multiple, parallel brine pumps – Estimate d 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 evapor ative 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 sy stem as tested had an initial refrigerant charge of 1,400 lb., which is considerably larg er than the stated goal. The reason for the larger refrigerant charge is that the seco ndary loop system also had heat reclaim capabilities for both hot water and space heat ing. The added charge was needed for piping and heat ex changers 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 approx imately 500 lbs. Further design

2enhancements such as including the heat re covery eq uipment 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 calc ulations 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/y r. 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 syst em will probably begin to leak more, which would increase the relative benefit of th e 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 produc t 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 l oop systems operated at similar rack SST values. The resulting average product te mperature 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 consumpt ion 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 q u antities of refrigerant, piping and fittings employed. Refrigerant leaks ar e difficult to find and repair, and after years of repairs and modifications the mu ltiplex 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 progra m objective of reducing the environmental costs of California’s electrical system, by de veloping an alternative refrigeration system which uses significantly less refrigerant than co nventional systems. It also contributed

3to the PIER program objective of improving energy value of California’s electricity by lowering electrical consumption of supermar ket secondary loop refrigeration systems.

4

## Abstract

Present multiplex supermarket refrigeratio n systems consume approx imately 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 refr igeration system employs fluid loops and a central chiller to provide refrigeration to th e 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 evapor ative condensing, respectively. Results of a field test comparison of the secondary l oop system, a multiplex refrigeration system with evaporative condensing developed ener gy savings of 4.9% for the secondary loop refrigeration system. Savings produced by the secondary loop refrigeration can be attributed to: minimum temper ature differences between disp lay case discharge air and saturated suction temperatures; the use of mu ltiple 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.

561.0 Introduction

1.1 Present Supermarket Refrigeration Sy stems One of the largest uses of energy in supermar kets is for refrigeration. Most of the product sold is perishable and must be kept refrigerated during storage and display. Typical energy consumption for supermarket refr igeration 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 ft 2 store). Compressors and condensers account for 30-35%. The remainder isconsumed 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 re turn between the machine room and the refrigerated fixtures.

Figure 1. Lay out of Refrigerated Display Cases and Walk-in Coolers in a Ty pical Supermarket

7Figure 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 suct ion and discharge manifolds, all installed on a skid. The skid contains all necessary pi ping, 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 co mpressors is piped to a remotely located condenser, which condenses the gas to a liq ui d. Liquid refrigerant returning from the condenser is piped back to the compressor rack, where a receiver, liq uid manifold and associated control valves are located, distri buting the liquid to the cases and coolers. Each case or cooler circuit is piped with liq uid 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 so lenoid valves used to control gas routing during defrost. The compressor rack is no rmally equipped with a number of pressure regulators used to control system head pressu re, 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 Sy stem is the Most Common Ty pe of Refrigeration Used in Supermarkets

Typically, three to five compressor racks are em ployed to provide all refrigeration in the supermarket. The display cases and cool ers are grouped, and attached to the compressor racks based on req u ired 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 requ irements. The low temperature racks will typically operate at a – 20°F SST. Refrigeration lo ads as low as – 30°F, or as high as – 10°F, will also be provided by the low temperatur e racks. In these situations, the suction

8manifold will be divided, and one or two compressors will provide the off-temperature refrigeration. The discharge of these “s atellite” compressors will be piped to the common discharge manifold with the other low temperature loads so that a common condenser and liq uid manifold can be used fo r all circuits on the rack. The remaining refrigeration circuits in the store are refe rred 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. Sate llite compressors are also used for medium temperature loads req u iring an SST sign ificantly higher or lower than 20°F. Multiplex systems commonly consist of three or four compressors that are sized such that operation of all compressors simultaneou sly can provide adequate capacity to meet the design refrigeration load. During off-des ign operation, the refrigeration load can be considerably less than the design value; at th e same time, the refrigeration capacity of the compressors can increase with a decrease in ambient temperature and the condenser operating pressure. In this situation, th e compressors can be selected so that the combined capacities of the compressors cl osely match the refrigeration load. The selection of, and the on-off cycling of the comp ressors is done based on suction pressure value measured at the compressor rack. The use of microprocessor-based controls allows more sophisticated control algorith ms 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 req u ire the least maintenance and have been shown to operate reliably in the non-operator envi ronment of supermarket refrigeration. Aircooled condensers employ finned coil constr uction 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 subs tantial difference in dry-bulb and wet-bulb temperatures exists. The evaporative condense r 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 tu be bundle where heat is removed and the refrigerant is condensed. The resulting co ndensing temperature can be close to the ambient wet-bulb. Evaporative condensers req u ire 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 issu es that prevent more use of evaporative condensers in supermarkets. Water treatmen t is needed because of the evaporation of the water, which concentrates dissolved mi nerals. 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 contro ller. The discharged water will carry away excess minerals and solids, which prevents prec ipitation. Biocides, such as chlorine, are added to the water by automatic feed ers to prevent biological growth.

9The use of evaporative condensers is strong ly recommended when energy-efficiency is an issue, because of the lower average condensi ng temperatures achieved with this type of heat rejection.

1.2 Low -charge Refrigeration Sy stem Options As a result of using the layout describe d above in supermarkets, the amount of refrigerant needed to charge a supermarket refr igeration 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 req u irin g significantly less refrigerant charge are now being considered. Advanced systems of this type include:

• Low-charge multiplex - Several refriger ation 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
approx imately 1/3.

• Distributed – Multiple scroll compressors are located in cabinets placed on or near the sales floor. Scroll compressors are empl oyed 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 th e use of a glycol loop that connects the cabinets to a fluid cooler, in orde r 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 capa city control, such as unloading, andwater-cooled condensers with a water loop for heat rejection. The advanced selfcontained refrigeration system would em ploy 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 case s 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 ra cks. 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 mec hanical hardware on or near the sales area, which is an arrangement not favored by supermarket operators. In contrast, the secondary loop refrigeratio n system employs central mechanical systems and reduces the total refrigerant charge to approx imately 300 to 500 lb. The anticipated refrigerant leak rate for the secondary loop re frigeration system is much less than that of the direct-expansion systems, because all re frigerant is contained within the chiller system.

10Another significant benefit derived from the op eration of secondary loop refrigeration is improved product temperature of stored pe rishables. Information provided in (2) suggests that with properly designed coils , display cases using secondary fluid for refrigeration produce a more uniform air temp erature throughout the case than seen in direct-expansion display cases. When warm br ine is employed for display case defrost, the time needed for defrost is significantly shorter, which prevents warming of the product. The resulting product temperatur es are lower and more consistent than the product temperature recorded in direct ex pansion cases operating at the same air temperature. Secondary loop refrigeration has been used extensively in Europe, particularly for medium temperature applications (3). Thes e present systems, in general, have used more energy than comparable multiplex refrig eration. The reasons for the increased energy use include:

• The use of display cases with evaporators designed for direct-ex p ansion 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 differenc e of the secondary fluid that must be provided to maintain the required refrig eration temperature. The lower saturated suction temperature reduces the efficien cy of the refrigeration compressors, increasing refrigeration energy use.

• Pumping energy associated with the circulat ion of the secondary fluid. Many of the secondary loop systems employ propylene gl ycol and water as the secondary fluid, because it is non-tox i c and acceptable for use near foods. At low temperature, the viscosity of the glycol-water mix t ure is much higher than that of water alone, greatly increasing the fluid friction and pumping power.

Presently, the installed cost of the secondar y loop refrigeration system may be higher than that of multiplex refrigeration. The reas on for this increased cost is the multiplex system is established and there are many q u alified contractors, which leads to cost competition. It is very likely that increasi ng the number of installations of secondary loop refrigeration could also cause the first cost of these systems to drop. Supermarket operators can just ify the higher first cost of the secondary loop system, if the benefits derived from its use provid e an acceptable payback and return on investment. The environmental and food storag e 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 redu ce it. Since the secondary loop system employs a chiller system similar to the mult iplex system, many of the energy-saving features now used with multiplex systems, such as floating head pressure, mechanical subcooling, and evaporative heat rejection c an be implemented to reduce energy use. Certain uniq ue characteristics of the second ary loop system can also be ex ploited for

11energy savings. These characteristics in clude: close-coupling between the chiller evaporator and the compressor suction which reduces pressure drop and refrigerant gas superheating; and liq uid refrigerant subcoolin g by brine warming for defrost. The use of a low viscosity fluid and the secondary fl uid along with better pump capacity control can greatly reduce the energy associated with brine pump. Finally, the use of display cases and fix t ures with heat transfer surfaces designed specifically for secondary loop systems can greatly reduce heat transfer penalties. Initial estimates for an advanced secondary l oop 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 sy stem now employed in many supermarkets.

1.3 Project Objectives The overall goal of this project was to de termine if a high-efficiency secondary loop refrigeration system could be economically attractive, due to reduced operating costs, than a state-of-the-art multiplex refrigeratio n system, which contains significantly more refrigerant. This project contributes to the PIER progra m objective of reducing the environmental costs of California’s electrical system, by de veloping an alternative refrigeration system, which uses significantly less refrigerant than co nventional refrigeration systems. It also contributes to the PIER program object ive of improving energy cost/value of California’s electricity by lowering electr ical 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 proj ect were to design and test a secondary loop refrigeration system, which:

• Consumes approx imately 14% less electric ity 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 re frigerant charge due to leakage.

The specific economic objective of the projec t 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 subc ontractor, 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 adv anced secondary loop system was prepared and several refrigeration manufacturers were asked to propose and bid on the

12construction. Hill PHOENIX, Inc. of Conyers, GA was selected as the system manufacturer and installing contractor base d on their response to the request for proposal. The advanced secondary loop refrigeration system was installed and instrumented for evaluation of its performance. Concurre ntly, a second supermarket operated by Safeway was identified in which a state-of-t he-art multiplex refrigeration system was installed. The multiplex system was inst rumented for performance measurement and both sites were monitored for approx imately 9 months. Originally, the test plan called for a one-year field test, but problems with th e 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) – d efined as:



### Required Power Supplied on i Refrigerat EER =

• Product temperature – Product sensors in similar cases were provided at each site so the average product temperatur e 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

13142.0 Project Approach

2.1 Design of the Secondary Loop Refrigeration Sy stem

2.1.1 Description of the Secondary Loop Refrigeration Sy stem Figure 3 shows the elements of the secondar y loop refrigeration system. The difference between secondary loop and direct ex pansion 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 th e 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 re frigeration 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 temp erature control and lower compressor energy use. The cooling of the secondary fluid ta kes place in a heat ex changer whereby the liquid refrigerant of the chiller system is ev aporated 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 reje cting heat to the lower ambient wet-bulb temperature. The lower condensing te mperature provides greater compressor refrigeration capacity and lower power dr aw, both of which result in reduced refrigeration energy consumption.

15Figure 3. Flow Diagram for a Secondary Loop Refrigeration Sy stem

The secondary loop refrigeration can be config ured 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 ex changers in the cases and coolers must be sized differently than those used for direct ex pansion 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 te mperature requirements can result in more energy efficient operation of the secondary loop system. Secondary fluid loop piping consists of wh ich 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 o ccurs 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

16branch piping is eq uipped 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 constr ucted of high-density polyethylene (4). Insulation is also required. For the me dium temperature piping, the recommended insulation is closed-cell foam (4). The reco mmended thickness of the insulation is 1⁄2 - 3⁄4 in. for piping underground or in air-cond itioned 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 ai r-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 refrigerat ing temperatures, and the fact the fluid is circulated continuously, the energy associ ated with pumping is substantial and therefore, is a major component of the over all energy consumption of the secondary loop system. The central chiller systems employ multiple co mpressors that are piped in parallel. The compressors are multiplexed and are on/off cycl ed 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 temperatur e refrigeration may be equipped with mechanical subcooling. A direct ex pansion he at ex changer is used to subcool the liq uid refrigerant used by the low temperature chiller system. Secondary loop refrigeration systems empl oy 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 comp ressors. The warm fluid is piped to the cases using an additional pipe. The fluid leav ing 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 substant ially (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 refr igeration 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 w i th Secondary Fluid A significant factor affecting the energy cons umption of the secondary loop system is the SST at which the refrigeration must operat e 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 ex pansion refrigeration system because of the added temperature differe nce associated with the secondary fluid. In previously installed secondary loop syste ms, the display cases were often the same as those used in direct ex pansion systems. The heat ex changer coils in these cases were designed specifically for refrigerant and ev aporation and did not provide sufficient heat transfer surface for secondary fluid use. Th e heat ex changers used req u ired a larger temperature difference between the secondar y fluid and the case air to achieve the desired display case air temperature. Prior to this project, Hill PHOENIX, Inc. (manu facturer and contractor for the field test secondary loop system) had initiated an ex tensive redesign of the display case heat

ex changer coils. As a result the coils are bette r suited for use with secondary fluid. The heat ex changer coils are larger than those used for direct refrigerant ex pansion, resulting in a lower temperature difference to achieve the desired air temperature. Table 1 shows a comparison of the req u ired operating conditions for secondary loop and direct ex pansion refrigeration for similar disp lay cases. For each display case type, the discharge air temperature of the case is lis ted. The expected saturated evaporator temperature and compressor rack SST needed to maintain this air temperature with direct ex pansion refrigeration is shown. The comparable secondary fluid supply temperature and SST of the chiller are also prov ided. 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 us e of the fluid loop. There are two reasons for this result: the first is that the heat ex changers of the secondary loop display cases have been sized to utilize a fluid at approxim ately the same temperature as the directex pansion display cases; and secondly, the co mpressors of the direct expansion system are located remotely from the case evapor ators 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 evaporat or temperature at the display case. The secondary loop refrigeration system has th e evaporator chiller and compressors close coupled (explained in detail below), which eliminates the pressure drop due to the piping runs necessary for the direct ex pansion system.

18Table 1. Comparison of Operating Temperatures for Direct-Expansion and Secondary

Fluid Refrigerated Display Cases

Display Case & Discharge Air T e mperature Direct-Expansion Refrigeration Secondary

- F l uid Refrigeration Display Case Discharge Air Te m p (°F) Evaporator Sat. Te m p (°F) Compressor Rack SST (°F) Secondary F l uid Supply Te m p (°F) Chiller SST (°F)

F r ozen F ood Reach-in - 1 2

- 1 9

- 2 5

- 2 0

- 2 5

Single-deck Meat 2 8 2 1 1 5 2 0 1 5

M u l t i

- d e c k D e l i 2 9 2 1 1 5 2 0 1 5 Multi-deck Produce 3 2 2 4 2 0 2 5 2 0

2.1.4 High-efficiency Refrigeration Compressors The chiller system may employ either reci procating or screw compressors. Screw compressors can provide the large refriger ation capacities needed to address the refrigeration load associated with each flui d loop. Screw compressors do not exhibit as high an energy efficiency value as reciprocat ing units, but are less susceptible to liq uid refrigerant damage and, in general, are cons idered 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 we re 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 seco ndary 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 ex pansio n systems incur substantial energy penalties because of the long pipe runs between th e 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 hi gher saturated suction temperature, which helps to maximize the EER.

2.1.6 Multiple Parallel Pumps The design secondary fluid flow rates repres ent the maximum brine flow rate required to refrigerate the display cases. When the re frigeration load is at less than design, a flow control valve in each case throttles the brin e flow to maintain discharge air temperature

19set point. Display cases in a defrost cycl e do not require chilled brine flow. For the above reasons, the total brine flow of each l oop is often considerably less than the design value. The total loop flow cannot be reduce d if a single pump is employed. Throttling of the pump to reduce flow reduces total pu mping head is required at the display cases which are in a cooling mode to insure prop er brine flow through the heat ex changer. 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, wh ile the full pressure head is maintained. Pump power is constant when a bypass is us ed to reduce system capacity when the system req u ires a constant head pressure. A variable-speed pump could be used, but th e matching of the head -flow characteristics of the pump to the head and flow values need ed 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 q u ite high and the energy savings often do not justify this added cost. The approach investigated is the use of mu ltiple, parallel pumps with each pump sized at approx imately 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 stop s 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 contin uous. A lowering of the pressure difference signifies more brine flow is required by th e system. One of the two remaining pumps is started and the second added pressure contin ues to drop to the start set point of the third pump.

2.1.7 Evaporative Heat Rejection Two types of condensers are available for su permarket 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 sign ificant, 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, resu lting in significant energy savings in the operation of the refrigeration compressors (5). Ambient airflow req u ired 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 req u ired to meet a

20refrigeration load, reducing the pumping powe r. Viscosity is a key parameter of fluid friction, and greatly influences pumping power. This is of particular importance for low temperature refrigeration applications, sinc e fluid viscosity increases as temperature decreases. Secondary loop systems often em ploy propylene glycol -water solutions as the brine, wherein the concentration of the glyc ol 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
req u ired operating temperatures and also hav e better transport properties. These fluids have been evaluated by others (6). The resu lts of these evaluations show the preferred secondary fluid is an organic-salt and wate r mix t ure, which has lower viscosity than propylene glycol-water and also has higher thermal capacitance. The organic-salt and water solution investigat ed in this project is marketed under the trade name of Dynalene. Table 2 gives a compar ison of the relevant transport properties of propylene glycol-water and Dynalene at fl uid temperatures of -20 and 20°F, which are standard values for low and medium temper ature 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 Dynal ene are most notable at -20°F, particularly in terms of viscosity.

Table 2. Transport Properties of Propy lene Gly c ol-Water and Organic Salt-Water (Dy n alene) at Operating Temperatures of Secondary Loop Refrigeration

Property Propy lene Gly c ol- W a ter Dy nalene Brine temperature = + 20°F Concentration (% by w e ight) 30 HC-10 Density - ρ (lb/ft 3 ) 6 4

. 9 7 5

. 0 Specific Heat - C (Btu/lb- °F ) 0.90 0.80 T hermal Capacitance ( ρ × C ) 5 8

. 4 6 0

. 0 Viscosity (centipoises) 9.91 2.84 Brine T e mperature = - 20 °F C o n c e n t r a t i o n 5 0 H C

- 3 0 Density - ρ (lb/ft 3 ) 6 6

. 5 8 1

. 2 Specific Heat - C (Btu/lb- °F ) 0

. 8 0

. 7 7 Thermal Capacitance ( ρ ×C) 5 3

. 2 6 2

. 5 Viscosity (centipoises) 156.1 8.78

The impact of the fluid transport properti es on the energy consumption of secondary loop refrigeration can be determined throug h evaluation of the relative pumping power (4,6).

21The relative pumping power is found by holding the refrigeration load, fluid temperature difference, pipe diameter and pi pe length constant. The relationship of heat transport, fluid friction and pump powe r are combined to give an expression for pumping power in terms of fluid properti es only, relative to a base fluid:

8

. 2

B A 2

. 0 8

. 1

B A

B A



### c c Power Pumping Power Pumping − − 

### ⎟ ⎟ ⎠ ⎞ ⎜ ⎜ ⎝ ⎛ ⎟ ⎟ ⎠ ⎞ ⎜ ⎜ ⎝ ⎛ ⎟ ⎟ ⎠ ⎞ ⎜ ⎜ ⎝ ⎛

=

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 temperat ures of -20 and 20°F, respectively.

These results indicate the use of lower vi scosity fluid in both the medium and low temperature fluid loops will reduce energy us e for brine pumping. For the advanced secondary loop refrigeration system, Dynal ene was specified for all fluid loops.

2.1.9 Refrigerant Subcooling from Warm Brine Defrost During operation, the cooling coils of the re frigerated 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 refrigeratio n loop and heated with refrigerant rejection heat to a temperature of approx imately 65 or 80°F for medium and low temperature refrigeration, respectively. The warm brin e 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 ex changer where heat from the brine melts the frost. The brine then flows to the return pipi ng 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 theref ore little excess heat is added to the total refrigeration load by the defrosting. The heating of the brine for defrost is do ne with two heat exchangers at the chiller system. The first ex changer uses heat fr om the liq uid refrigerant coming from the system receiver. The second heat ex changer 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 ho t liquid refrigerant results in lowering significantly the temperature of the refriger ant, producing a subcooled liquid. The use of this sub-cooled refrigerant liquid in th e chiller system increase s the chilling capacity of the system, which means fewer compressors need to operate to satisfy the refrigeration load. The reduction in comp ressor run time results in a reduction in compressor energy used.

22Field testing showed the refrigerant subcoo ling produced by the warm brine defrost averaged about 12°F, which corresponds to ca pacity increases of 6.9 and 6.4% for low and medium temperature refrigeration, respectively.

2.2 Analy s is of the Secondary Loop and Multiplex Refrigeration Sy stems Two refrigeration system models were formul ated for the analysis work done on this project. The first model was used to de termine the energy consumption of a direct

ex pansion 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 refrig eration system is calculated on the basis of ambient dry-bulb temperature bins, wher e each bin specifies: an ambient dry-bulb temperature value; the coincident value of th e 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 th e power needed by the refrigeration system and apply that power to the number of ho urs 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 evap orator temperature in the system, minimum condenser temperature, type of condense r (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 refrigerat ion load will vary with outside ambient drybulb temperature, decreasing as the ambi ent temperature decreases. The rate of decrease in load is greater for medium te mperature than low temperature refrigeration, and a minimum is reached, which is a resu lt of space heating the store. Store temperature will not fall below 68°F. The minimum store temperature is normally present at outside ambient temperatures le ss than 40°F. The max i mum values of store temperature and humidity are also constrained by the use of air conditioning, which tends to maintain the temperature at 75°F wi th 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. Ba sed 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 refriger ation, the value of the refrigeration load is set at design for ambient temperatures of 85°F or greater.

23At ambient temperatures of 40°F or less, the re frigeration 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:



## ( ) ⎟ ⎟ ⎠ ⎞ ⎜ ⎜ ⎝ ⎛

−
= 40)

- (85 ) T 85 ( min

- 1

- 1 factor Load amb

where

min = the minimum fraction of design load (0.66 for medium temperature and 0.8 for low temperature T am b = ambient dry-bulb temperature

The state points for the multiplex refrigeratio n 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 lowe r 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, increa sing with increased distance. In the analysis, these factors are taken into accoun t by values included with the system configuration description. The condenser temperature is also determin ed at this time. The most significant parameter in determining condensing temper ature 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 te mperature is reached. At that point, control of the condenser fans or a liq uid pressure regulato r maintains the condensing temperature at the minimum value. The model compar es the ambient temperature with the characteristic TD of the condenser type specified and calculates a condensing temperature. The calculated value is comp ared with the specifie d minimum condensing temperature. If the calculated temperat ure is less than the mi nimum, 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-cooledcondensers, 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 diffe rence 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

24is achieved at a TD value of approx imatel y 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 liq uid temperature is 10 degrees less than the condensing temperature. When mechanica l subcooling is employed in multiplex systems, the liq uid refrigerant temperature leaving the subcooler heat ex changer is typically 40°F. Some warming of the liquid o ccurs 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 norma lly 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 lo ad of the subcooled system, normally the low temperature refrigeration. For each temperature interval the low temperature refrigeration load is first determined, and th e 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 liq uid flow from 10° F less than the condensing temperature to 40°F. The subcooling load is added to the re frigeration 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 cons ists of refrigeration capacities (Btuh), refrigerant mass flow rate (lb/hr), and inpu t power (Watts) as functions of saturated suction temperature (SST) and saturated disc harge temperature (SDT). These data are obtained from the ARI equation:
Capacity, Mass Flow, or Power = C 0 + C 1 *SST + C 2 *SDT + C 3 *SST 2

+ C 4 *SST*SDT + C 5 *SDT 2 + C 6 *SST 3

+ C 7 *SDT*SST 2 + C 8 *SST*SDT 2

+ C 9 *SDT 3 where

C 0 ... C 9 = 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 ca pacity, mass flow rate, or compressor input
power. The compressor cooling capacity and mass fl ow 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 liq uid subcooling. Corrections are made to account for the values of superheat and refr igerant liquid temperature. The superheat correction takes into account the density and enthalpy change, while change in liq uid temperature affects the enthalpy of the re frigerant entering the evaporator. The correction applied to the compressor capacity or mass flow rate is:

25

### ) ( ) (h *
v v correction out r

rin rout in 

### h h h Capacity

− −
=

where
v = the specific volume of the refrigerant (ft 3 /lb) h in = the enthalpy of the refr igerant entering the evaporator h out = the enthalpy of the refrig erant 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 req u ired . Fractional values represent compressor on/off cycling. The analysis does not use sp ecific compressor models, but instead uses a single generic size for each type of compresso r. 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 temper ature 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 comp ressors operating and the number of hours associated with the ambient temperature. The fan power for remote condensers or fl uid 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 req u ire more fan power than condensers employed with medium temperature refrig eration. Fan req u irements are less for evaporative heat rejection than is needed fo r dry rejection, because less air flow is required. The power value listed in the ta ble for the evaporative units includes the sump pump used to spray water over the heat ex changer coil. The condenser fans operate continuously as long as the resulting condensing temperature is greater than the specified mini mum value. Fan cycling is employed with both the condensers and fluid coolers to re gulate 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 need ed 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 ambi ent dry-bulb temperature and the condenser TD are greater than the minimum condensing te mperature. 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:



### ⎟ ⎟ ⎠ ⎞ ⎜ ⎜ ⎝ ⎛

= 30)

- TD

- (T ) T

- TD

- (T * 0.75

- 1 Factor Fan

con amb con

where

26T con = the minimum condensi ng temperature (°F) TD = the temperature difference (°F) T am b = 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 installe d fan power, the fan factor, and the number of hours in the bin. The energy for mechanical subcooling of the lo w temperature refrigeration is addressed by the medium temperature system with the highest SST. As mentioned previously, the subcooling load is calculated for each temper ature bin and is added to the refrigeration load of the appropriate medium temperatur e 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 specifie d. Once the bin loop is completed, the model then obtains the nex t system description and the bin loop is repeated for this nex t 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 se condary loop refrigeration system is the operation of the secondary loop. The loop co nsists 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 th e 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 va ried 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 temperat ure from the display case heat ex changer. Since all loads on the loop behave in this fa shion, the estimated total fluid flow can be found from:

brine brine ref brine 

### T C Q M ∆ = •

where
Q ref = the refrigeration load addressed by the brine loop (Btuh)


### (lb/hr) brine of rate flow mass the = •
brine 

### M

27

### cases at the seen brine in the change re temperatu the = ∆ brine



### T


### brine the of heat specific the = brine 

### C

The secondary fluid loop will experience so me heat gain, while flowing between the cases and the central chiller. The most signific ant 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 desi gn 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 c hanges, 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 calc ulated as the ideal power for the maximum fluid flow and head divided by a pump effici ency 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 Power Pump

where Pump Power = the power input to the secondary fluid pump

Some line heat gain is also ex pected and was set at 0.25°F for the supply and return lines of the loop. Once the total temperature rise of the second ary fluid loop is determined, the load on the chiller evaporator can be found by:

brine brine brine evap 

### T C M Q ∆ = •

where



### fluid in the seen gain re temperatu total the = ∆ brine



### T

Mechanical subcooling is used in the second ary 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 subc ooling. The refrigeration load associated with the subcooling must be added to the tota l 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 ra te 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 th en determined. The evaporator temperature of the chiller heat exchanger is set at 7°F belo w the outlet temperature of the fluid loop.

28The outlet temperature of the refrigerant is set at 8°F higher than the evaporator temperature. The temperature rise is du e 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 superh eated 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 ex changer. The SDT of the compressor system is determined from the condensing temperat ure, 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 ea ch heat rejection type is the same as described previously.

The central chiller uses mult iple parallel compressors to address the chiller heat

ex changer load. The types of compressors now employed in presently installed systems are either screw or reciprocating. Scroll co mpressors can be analyzed if desired. The procedure for determining the number of co mpressors operating is the same as that used for the multiplex and distributed syste ms. 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 numb er of hours in the temperature bin being examined. The energy consumption of the secondary flui d 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 th e pump is constant. Variation in flow is achieved by flow bypassing around the pu mp. The head addressed by the pump was estimated based upon the temperature of th e 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 ex amined here consists of an organic salt-water mix t ure (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 rang es 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 pu mps 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 sa me 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 fl uid 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 usin g the procedures described previously.

292.3 Field Testing of the Secondary Loop Refrigeration Sy stem

2.3.1 Description of the Secondary Loop Refrigeration Test Site The design and analysis work of the projec t were used to specify a secondary loop refrigeration system for an operating superm arket. 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 supermarke t is located in Thousand Oaks, CA and has a total building and sales areas of 48,351 and 41,912 ft 2 , 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. Lay out of the Secondary Loop Refrigeration Test Store

30Table 3. Low Temperature Circuits for the Secondary Loop Refrigeration Test Store (Loop A)

C i r c u i t F i x t u r e D e s c r i p t i o n Refrig Load (Btuh) Disch. Air Te m p (°F) Brine F l ow (gpm)
A-1 Reach-in F r ozen Pasta 3,000 -5 0.9 A-2 W i de Island F r ozen M e a t 1 9

, 8 0 0

- 1 2 1 8

. 2 A-3 W i de Island F r ozen M e a t 1 7

, 4 0 0

- 1 2 1 6

. 2 A-4 F r eezer W a lk-in Box 31,300 -12 24 A-5 Reach-in F r ozen F ood 24,000 -5 6.9 A-6 Reach-in F r ozen F ood 21,000 -5 6 A-7 Reach-in F r ozen F ood 24,000 -5 6.9 A-8 Reach-in F r ozen F ood 21,000 -5 6 A

- 9 R e a c h

- i n Ice C r e a m 2 2

, 4 0 0

- 1 2 1 2

. 6 A

- 1 0 R e a c h

- i n Ice C r e a m 1 9

, 2 0 0

- 1 2 1 0

. 8 A

- 1 1 R e a c h

- i n Ice C r e a m 2 5

, 6 0 0

- 1 2 1 4

. 4 A

- 1 2 R e a c h

- i n Ice C r e a m 2 2

, 4 0 0

- 1 2 1 2

. 6 A

- 1 3 B a k e r y F r e e z e r 9

, 4 0 0

- 5 4 A

- 1 4 D e l i F r e e z e r 9

, 4 0 0

- 5 4 Pump Heat 11,400 T o tal Refrigeration Load 281,300 144

Brine Supply T e mperature = -20°F Chiller SST = -25°F

31Table 4. Medium Temperature Circuits for the Secondary Loop Refrigeration Store (Loop B1)

C i r c u i t F i x t u r e D e s c r i p t i o n Refrig Load (Btuh) Disch. Air Te m p (°F) Brine F l ow (gpm)
B1-1 Deli W a lk-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 W a lk-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 F oods Case 7,200 28 10.3 B1-16 Service Deli Cases 6,400 31 8 B 1

- 1 7 S a n d w ich C a s e s 1 4

, 6 0 0 2 8 1 3

. 5 Pump Heat 16,000 T o tal Refrigeration Load 337,800 159

Brine Supply T e mperature = 20°F Chiller SST = 15°F

32Table 5. Medium Temperature Circuits for the Secondary Loop Refrigeration Store (Loop B2)

C i r c u i t F i x t u r e D e s c r i p t i o n Refrig Load (Btuh) Disch. Air Te m p (°F) Brine F l ow (gpm)
B2-1 F l ow er W a lk-in Cooler 8,600 38 4 B 2

- 2 M e l o n T a b l e 6

, 4 0 0 3 5 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 W a lk-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 W a lk-in Cooler w i th 14 Doors 46,200 36 27 B2-13 Bakery W a lk-in Cooler 7,300 36 4.5 B 2

- 1 4 M u l t i

- d e c k B e v e r a g e 5 6

, 6 0 0 3 6 2 0

. 1 B 2

- 1 5 M u l t i

- d e c k B e v e r a g e 4 4

, 7 0 0 3 6 1 5

. 9 B 2

- 1 6 F l oral C a s e 7

, 8 0 0 3 6 4 B 2

- 1 7 P a s t a C a s e 1 2

, 0 0 0 3 1 4

. 3 Pump Heat 21,800 T o tal Refrigeration Load 446,300 196

Low T e mp Subcooling 49,800

Brine Supply T e mperature = 25°F Chiller SST = 20°F

Figure 5 and Figure 6 show diagrams descri bing the operation of the secondary loop refrigeration system for low and medium temp erature, 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 te mperatures 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 thro ugh 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 heat ing, or for defrost brine heating. The discharge refrigerant flow is piped to the ev aporative condenser. From the condenser, the liquid refrigerant flow from the two disc harge circuits is combined at a single receiver tank. The liquid refrigerant flow from the receiver is divided into two streams, which pass through heat ex changers that subc ool the refrigerant by warming brine that is used for display case defrost. After thes e heat ex changers, the liq uid refrigerant flow is combined at the liquid manifold wher e 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

33system is further sub-cooled in a mec hanical sub-cooling heat ex changer. 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 Sy stem (Patent by Hill PHOENIX, Inc.)

34Figure 6. Piping for the Medium Temperat ure Secondary Loop Refrigeration Sy stem (Patent by Hill PHOENIX, Inc.)

In each fluid loop, the secondary fluid is circulated between the chiller evaporator and the refrigerated fixtures by three paralle l 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 rest arted when the pressure falls.

2.3.2 Measurement Plan to Monitor Secondary Loop Refrigeration A measurement plan was developed to dete rmine 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 concer n to the project, power measurement of the refrigeration system was the primary measur ement monitored. Power measurements of refrigeration compressors, brine pumps, and condenser fans and pumps were noted at regular intervals and used to determine ener gy consumption of the refrigeration system.

35Power measurements were recorded at 10-se cond intervals which were later combined into 15-minute averages. The average values were stored at the data acquisition for remote collection. Watt transducers we re 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 a ddressed 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 take n at regular intervals. The refrigeration supplied by each brine loop was determined from:



## ( ) bout bin brine 

### T T Q

− = c

m where

T b 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 refr igeration to the chiller evaporator is found from: 

# ( ) liq out 

### h h m

− = chiller 

### Q
where h is the enthalpy of the refrigerant at th e 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 suctio n and discharge temperatures at which the compressor is operating. Corrections ar e needed to account for the refrigerant temperature at the suction of the compressor and for the refrigerant liq uid. 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/o ff cycles of each compressor determined the run fraction. The results from the power and refrigeratio n measurements were then combined to determine the energy efficiency ratio (EER) of the refrigeration system. The EER is defined as:

36

### hr) (Btu/Watt quired Power nSupplied frigeratio EER



### − = Re Re

For the secondary loop refrigeration, the po wer required included both the compressor and brine pumping power. Two EER values were calculated, one for low temperature refrigeration and a second for medium te mperature 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 fo r mechanical subcooling was subtracted from the medium temperature refrigeration load.

2.3.2.3 Refrigeration Sy stem Operating State Points The significant operating state points that we re 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 liqui d temperature was measured at several locations in the system to characteri ze 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 defr ost heat ex changer. Additional liq uid temperature measurements were needed at the outlet of the mechanical subcooling heat exchanger for the low temperature chiller.

• Refrigerant discharge temperature – The te mperature was measured at the inlet of the brine defrost heat ex changer, and at the inlets of the heat reclaim water heater (an outlet temperature was also req u ired) and space heater coil.

• Condenser Pan Water temperature – the pr imary 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 monitore d to evaluate the pump control strategy. Along with pump power, digital on/off signal s were included to monitor the operation of the brine pumps.

372.3.2.5 Brine Defrost Operation The warm brine defrost system specified in this project used a combination of refrigerant liq uid subcooling and discharge he at to raise the brine temperature. The subcooling of the refrigerant increases the re frigeration 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 ex changer.

• Brine temperature in and out of defrost/subcooling heat ex changer.

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 eq uippe d 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 – t hree display cases

• Multi-deck produce cases – t wo display cases

Similar display cases at the multiplex store were also eq uipped with product simulators to allow comparison of product temperature and temperature variation for the two sites.

Other case measurements that were addresse d 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 te mperature and humidity were monitored. The inside reading consisted of multiple re adings of the dry-bulb temperature and a single reading of the inside relative humi dity. The outside ambient instrumentation consisted of single readings of the dry-bu lb temperature and the relative humidity.

Problems were incurred with the outside humi dity reading shortly after the start of testing. An alternate approach for outside ambient data was employed after the sensor failed. Southern California Edison maintain s weather stations throughout the service territory. One station is located in Moorpa rk, CA, which is in close proximity to the secondary loop store. Good agreement wa s 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 us ed for the entire test period to represent the outside ambient conditions at the store. Si milarly, weather station readings were also

38used to characterize the ambient for the mult iplex 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 secondar y loop store were wired to the energymanagement 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 Sy stem

2.3.3.1 Description of the Multiplex Refrigeration Test Site A second Vons supermarket, located in Vale ncia, CA, was selected and instrumented for performance comparison with the secondar y loop refrigeration system. The store selected employed a direct expansion, mu ltiplexed refrigeration system to provide refrigeration to all display cases and walk-in co olers. The total building area of the store is 56,526 ft 2 . Figure 7 shows the layout of the sales area of the store, which measured 41,912 ft 2 . Table 6, Table 7, Table 8, and Table 9 de scribe the refrigerated fixtures at the multiplex test store.

Figure 7. Lay out of the Multiplex Refrigeration Test Store

39Table 6. Low Temperature Circuits at the Multiplex Refrigeration Store (Rack A)

C i r c u i t F i x t u r e D e s c r i p t i o n Refrig Load (Btuh) Evap Te m p (°F) Disch Air Te m p (°F)
A1 Reach-in Ice Cream/F rozen F ood Case 28,050 -19 -12

A2 Reach-in Ice Cream/F rozen F ood Case 28,050 -19 -12 A3 Reach-in Ice Cream/F rozen F ood Case 29,700 -19 -12 A4 Reach-in Ice Cream/F rozen F ood Case 28,050 -19 -12 A5 Reach-in Ice Cream/F rozen F ood Case 26,400 -19 -12 A6 Reach-in Ice Cream/F rozen F ood Case 24,750 -19 -12 A7 Reach-in Ice Cream/F rozen F ood Case 28,050 -19 -12 A8 Reach-in Ice Cream/F rozen F ood Case 33,000 -19 -12 A9 W i de Island F r ozen F i sh/Meat Case 17,050 -20 -12 A10 W i de Island F r ozen F i sh/Meat Case 13,950 -20 -12 A 1 1 D e l i F r e e z e r 8,200

- 1 5

- 5 A 1 2 B a k e r y F r e e z e r 10,000

- 1 5

- 5 A 1 3 G r o c e r y F r e e z e r 35,000

- 2 0

- 1 0 A14 Reach-in Ice Cream/F rozen F ood Case 4,950

- 1 9

- 1 2
T o tal Refrigeration Load (Btuh) 315,200 Low T e mp Rack A SST = -25°F

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

C i r c u i t F i x t u r e D e s c r i p t i o n Refrig Load (Btuh) Evap Te m p (°F) Disch Air Te m p (°F) B2 Deli Walk-in Cooler 5

, 8 0 0 2 0 3 6 B 4 S a n d w i c h C a s e 5

, 6 0 0 1 8 2 8 B1 Deli Island Case 1 4

, 4 0 0 1 5 2 5 B3 Cheese Island Case 1 4

, 4 0 0 1 5 2 5 B5 Service Deli Cases 7

, 0 4 0 1 5 2 8 B6 Multi-deck Deli Cases 2 2

, 8 0 0 2 4 3 2 B7 Bakery Island Cases 9

, 6 0 0 1 5 2 5 B 8 R e t a r d e r Box 8

, 0 0 0 2 0 3 6 B9 Self-Service Bakery Case 2 6

, 8 0 0 2 0 3 4 B10 Multi-deck Dairy Case 1 7

, 1 0 0 2 4 3 2 B18 Produce Island Case 1 4

, 4 0 0 1 5 2 5 B11 Multi-deck Meat Cases 4 5

, 7 2 0 2 1 2 9 B12 Single-deck Meat Cases 1 2

, 9 6 0 2 1 2 8 B13 Self-Service Fish Case 5

, 4 4 8 1 9 2 8 B14 Service Fish Cases 2

, 4 0 0 1 5 B15 W i de Island Meat Cases 2 2

, 6 8 0 1 8 2 6 B16 Multi-deck Lunch Meat Cases 3 9

, 9 0 0 2 4 3 2 B17 Multi-deck Lunch Meat Cases 1 1

, 4 0 0 2 4 3 2 T o tal Refrigeration Load (Btuh) 286,448 Medium T e mp Rack B SST = 13°F

40Table 8. Medium Temperature Circuits at the Multiplex Refrigeration Store (Rack C)

C i r c u i t F i x t u r e D e s c r i p t i o n Refrig Load (Btuh) Evap Te m p (°F) Disch Air Te m p (°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 F l oral W a lk-in Cooler 8,100 20 38 C11 F l oral Display Case 12,200 20 38 C7 Produce W a lk-in Cooler 39,120 20 36 C8 Seafood W a lk-in Cooler 9,700 20 36 C9 Dairy W a lk-in Cooler w i th 16 Doors 50,000 20 32 C10 Meat W a lk-in Cooler 31,400 20 32 T o tal Refrigeration Load (Btuh) 499,830

Medium T e mp Rack C SST = 20°F

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

C i r c u i t F i x t u r e D e s c r i p t i o n Refrig Load (Btuh) Evap Te m p (°F) Disch Air Te m p (°F) D1 Produce Prep Area 24,000 3 7 5 0 D2 Meat Prep Area 95,000 3 7 5 0 T o tal Refrigeration Load (Btuh) 119,000

Subcooling Sy stem A 49,950 Subcooling Sy stem B 44,708 Subcooling Sy stem C 81,634
T o tal Subcooling Load (Btuh) 176,292

Medium T e mp 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 refr igeration, respectively. The refrigeration was divided over three compressor racks, co nsisting of one low temperature and two medium temperature racks. One of the medi um 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 cons idered unusual in that normally only the low temperature rack is mechanically subcoo led. An evaporative condenser provided heat rejection. The refrigeration system was eq uipped to provide heat reclaim for both hot water and space heating.

41Figure 8. Piping Diagram for the Low Temperature Multiplex Refrigeration

Figure 9. Piping Diagram for the Medium Temperature Multiplex Refrigeration

422.3.4 Measurement Plan for the Multiplex Refrigeration Sy stem A measurement plan was also developed for th e multiplex system that included all key parameters to allow comparison of performance with that of the secondary loop refrigeration. The elements needed fo r 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 in tervals. The measurements were used to determine energy consumption and for th e multiplex refrigeration system EER calculation. The total power and energy of the store was also monitored. Refrigeration Supplied The refrigeration load delivered by ea ch refrigeration rack was found from:



# ( ) liq out 

### h h m



### − = •
rack 

### Q
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 estimate d through the published capacity data for the compressors and digital on/off readings that described compressor operation as

ex plained above for the secondary loop compressors. Several of the medium temperature compressors employed cylinder un loading for capacity control. Operation of the unloaders was monitored and counted by digital on/off readings.

2.3.4.2 Refrigeration Sy stem Operating State Points The significant operating state points t hat 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 ex changers to characterize the subcooling provided.

• Refrigerant discharge temperature – Measured at the inlets of the heat reclaim water heater (an outlet temperature was al so req u ired) and space heater coils.

• Condenser pan water temperature – the primary control point for the condenser fans.

• On/off digital signals were used to track even ts such as the initiation of defrost or heat reclaim.

2.3.4.3 Display Case Operation The operation of the display cases and walk-i n boxes were monitored during testing. Each display case is equipped with a dischar ge air temperature reading used to control operation of the refrigeration system. Several display cases were eq uipped with prod uct simulators, similar to those used with the secondary loop system, so an evaluation of product temperature maintenance could be made. The display cases eq uipped with product simulators were

• Single-deck meat – 2 display cases

• Multi-deck produce – 2 display cases

The display cases selected for product simulato rs 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 addresse d 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 monito red with several dry-bulb temperature measurements and one relative humidity read ing. 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 collect ed routinely by Southern California Edison and provided to the project on a monthly basis.

4445 3.0 Project Outcomes

3.1 Analysis Results The analytical models were used to pred ict 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 estimate s for the secondary loop refrigeration system. The modeled secondary loop sy stem included evaporative condensing, multiple parallel secondary fluid pumping, and warm brine defrost subcooling to help decrease energy consumption. Mechani cal 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 condensi ng. 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 791,491 187,982 979,47346 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 syst em 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 secondar y 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 refrigeratio n 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 Case s 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

O

c t

' 0

2 N o

v D e

c
J

a

n

' 0

3 F e b M a

r A p

r M a

y

J

u

n

eAverage DailyEnergy (kWh/day)



### Sec o ndary Loop Ref r iger at ion Sec o ndary Loop St or e

Mult iplex Ref r iger at ion

Mult iplex St ore

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

Figure 11 and Figure 12 show the breakdown s 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 th e multiplex system. Both display case and condenser energy was higher at the secondary lo op site. One reason for the higher case energy use was the display case lights were turn ed off at the multiplex site during late evening hours; this practice was not follo wed 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 on ly briefly during the test period, meaning minimum fan operation was req u ired to ma intain the condenser minimum set point temperature.

48

### 0 50 0 1,00 0 1,50 0 2,00 0 2,50 0

O

c t

' 0

2 N o

v D e

c

J

a

n

' 0

3 F e b M a

r A p

r M a

y J u

n

eAverageDaily Energy(kWh/day)



### C o m p ressors Co nd ens er Ca s e s

Figure 11. Multiplex Refrigeration Sy stem Energy Consumption



### 0 500 1, 000 1, 500 2, 000 2, 500

O

c t

' 0

2 N o

v D e

c

J

a

n

' 0

3 F e b M a

r A p

r M a

y J u

n

eAverage Daily Energy(kWh/day)



### Com p ress ors Pu m p s Condenser Cases

Figure 12. Secondary Loop Refrigeration Energy Consumption

4950 One of the major energy differences between the multiplex and secondary loop refrigeration systems was the operation of th e 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 arrangem ent. 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 fl ow 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 consumptio n 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 measur ed for the secondary loop store during January because of data acquisition pr oblems. Figure 13 shows a graphical representation of these power values. The peak power demand of the multiplex store was higher than the secondary loop siteTable 16. Peak Pow e r Demand of the Multiplex Store

Date T i me Store Refrig % of Store O c t ' 0 2 9

- O c t 1 6 : 5 5 3 7 0

. 3 1 3 5

. 2 4 6

. 2 N o v 9

- N o v 1 5 : 3 9 3 2 9

. 3 1 1 8

. 6 4 7

. 3 D e c 1 6

- D e c 1 5 : 4 7 3 4 8

. 6 1 0 2

. 3 4 1

. 5 Jan ' 03 16-Jan 8: 3 1 3 3 5

. 2 1 0 2

. 2 4 2

. 9 F e b 11-F e b 1 1 : 1 3 3 5 1

. 9 1 1 7

. 7 4 3

. 4 M a r 1 4

- M a r 1 4 : 2 9 3 2 4

. 3 1 2 6

. 2 5 0

. 2 A p r 4

- A p r 9 : 5 0 3 2 2

. 7 1 0 0

. 1 4 1

. 5 May 28-May 1 4 : 3 7 3 3 2

. 7 1 3 3

. 5 5 0

. 9 J u n 1 6

- J u n 1 4 : 2 2 3 4 1

. 9 1 2 9

. 3 4 8

. 5

Table 17. Peak Pow e r Demand for the Secondary Loop Store

Date T i me Store Refrig % of Store Oct ' 02 6-Oct 16:30 312 111.2 51.5 N o v 9

- N o v 1 7 : 1 5 3 1 2 1 2 4

. 8 5 6

. 7 D e c 4

- D e c 1 7 : 1 5 3 0 4 1 2 4

. 8 5 6

. 6 Jan ' 03 NR F e b 13-F e b 9 : 4 5 2 9 6 1 2 0 5 7

. 6 M a r 3 1

- M a r 1 2 : 0 0 3 2 0 1 3 0

. 4 5 6

. 5 A p r 1 6

- A p r 1 3 : 3 0 3 2 0 1 4 6

. 4 6 2 May 19-May 1 2 : 0 0 3 2 0 1 3 6

. 8 5 7

. 3 J u n 1 6

- J u n 1 6 : 0 0 3 2 0 1 5 0

. 4 6 0

. 5

0 50 10 0 15 0 20 0 25 0 30 0 35 0 40 0

O

c t

' 0 2 N o v D e

c
J

a

n

'0

3 F e b M a

r A p r M a y J u nPowerDemand - 15min (kW)

M u lt iplex - Store

M u lt iplex - T o t a l Ref r igerat ion Se con dar y Loo p - St or e Se con dar y Loo p - T o t a l Ref r igerat ion

Figure 13. Peak Pow e r Demand for the Multiplex and Secondary Loop Stores

513.2.3 Refrigeration Efficiency Table 18 and Table 19 show the average refr igeration loads, power, and EER for the multiplex and secondary loop systems, respecti vely. The refrigeration load at each site has been divided into two segments wh ich are the low and medium temperature refrigeration loads. The low temperature re frigeration load is addressed by the low temperature system at each site. The remain der of the refrigeration load was combined as the medium temperature load. The compresso r power was also divided on this basis. The only ex ception is the mechanical subcoo ling 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 th e medium temperature refrigeration. Mechanical subcooling is not normally used with medium temperature refrigeration and would not be representative of a baseline mult iplex 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 additi onal 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 liq uid temperatur es, 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 action NoSubRunFr ref 

### Q

where
Q ref 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

.

5253 Table 18. Multiplex Refrigeration

Low Temperature Medium Tem perature 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 refr igeration show the secondary loop system operating at an approximately 6.6% higher efficiency than the multiplex system. For medium temperature refrigeration, the mult iplex 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 effi ciency data can be achieved by considering the ambient wet-bulb temperature variation, si nce both refrigeration systems reject heat to an evaporative heat exchang er. Figure 14 shows the plot for the EER versus the ambient wet-bulb temperature for low temperature refrigeration. Figure 15 and Figure54 16 show this same relation for medium temp erature refrigeration for the multiplex and secondary loop systems, respectively. A lin ear 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 sa me 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 Bt u/hr for low and medium temperature refrigeration, respectively, and the aver age ambient wet-bulb temperature at the multiplex site was 49.2°F. For the secondar y loop store the refrigeration loads were 344,800 and 617,000 Btu/hr for the low and medi um temperature, respectively, and the average ambient wet-bulb temp erature was 50.3°F. Using these values, and the energy consumption of the secondary loop and mult iplex 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, resp ectively. 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 Temperatur e Load – 530,100 Btu/hr Average Ambient Wet-Bulb Temperature – 49.2°FTable 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 T e mperature Refrigeration Secondary Loop 8.26 1,001.2 M u l t i p l e x 7

. 5 8 1, 0 6 6

. 8 6 5

. 6 ( 6

. 1 % ) Medium T e mperature Refrigeration Secondary Loop 12.41 1,192.9 M u l t i p l e x 1 2

. 5 4 1,181. 1

- 1 1

. 8 (

- 1

. 0 % ) Multiplex – No Subcooling 1 1

. 9 4 1

, 2 4 0

. 1 4 7

. 2 ( 3

. 8 % )

T o tal Refrigeration Secondary Loop 2,194.1 M u l t i p l e x 2

, 2 4 7

. 9 5 3

. 8 ( 2

. 4 % ) Multiplex – No Subcooling 2

, 3 0 6

. 9 1 1 2

. 8 ( 4

. 9 % )

Low T e mperature Load – 340,800 Btu/hr Medium T e mperatur e Load – 617,000 Btu/hr Average Ambient W e t-Bulb T e mperature – 50.3°F



### y = - 0 .0066x + 8.0887 R 2 

### = 0. 0506 y

= - 0 .0261x + 9.5784 R 2 

### = 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
A m bient Wet- Bulb T e m p erat ure (F)System EER (Btu/W-hr)

S e c o nd ar y Lo o p
M u l t ip le x S e condary L

Mult iplex

Figure 14. Relation betw een Ambient Wet-Bulb Temperature and Low Pressure Refrigeration EER

55

### y = -0.0597x + 15. 538 R 2 

### = 0.2474

y

= -0.0599x + 14. 954 R 2 

### = 0.2458 8.00 9.00 10.00 11.00 12.00 13.00 14.00 15.00 16.00 17.00

30 40 50 60 70
A m bient Wet-bulb

T e mperatur e (F)SystemEER (Btu/W-hr)



### M u lt ip le x

M u lt ip le x - N o Subcooling

Figure 15. Relation betw een Ambient Wet-Bul b Temperature and Medium Temperature Refrigeration EER for the Multiplex Sy stem



### y

= - 0 .0248x + 13.661 R 2 

### = 0.0629

8.00 9.00 10.00 11.00 12.00 13.00 14.00

30 40 50 60 70
A m bi e n t W e t-bu l b

Temp erat ure ( F )SystemEER(Btu/W-hr)



### Sec ondar y Loop

Figure 16. Relation betw een Ambient Wet-Bul b Temperature and Medium Temperature Refrigeration EER for the Secondary Loop Sy stem

5657 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 th e temperature of a thermal mass to be monitored. Table 22 shows the average temperature measur ements for the instrumented cases for a 24-hour period. The uniformity of the pr oduct temperature is characterized by the standard deviation ( σ ) of the product temperature. By definition, 95% of the product temperature values will fall within +/- 2 σ . Temperature plots for the meat cases are shown in Figure 17 and Figure 18, for th e 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 th e 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 di scharge 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 temperat ure includes values during defrost. The uniformity of the product temperature was better for the secondary loop produce case by a small amount.

## Mu lt ip le x R e f r ig er at ion - Sing le

- d ec k Mea t Ca se

0 5 10 15 20 25 30 35 40 45024791214171922


### Hour of DayTemperature(F)


### Ca se Dishcar g e Air Product

Figure 17. Discharge Air and Product Temperat ure Profiles for the Single-Deck Meat Case - Multiplex Refrigeration



## Secondary Loop - Single -deck Meat Display Case (B1-6a)

0 10 20 30 40 50 60024791214171922


### Hour of DayTemperature(F)



### Case Dischar ge Air Pr o duct

Figure 18. Discharge Air and Product Temperat ure Profiles for the Single-Deck Meat Case - Secondary Loop Refrigeration

58

## Multiplex Refri g er ati o n - M u l t i-deck Produce Cases

0 10 20 30 40 50 6002469111316182023 

### Hour of DayTemperature(F)



### Disc harge Air Pr oduct

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 50024791214171922


### Hour of DayTemperature(F)



### Case Dischar ge Air Pr o duct

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

593.2.5 Refrigerant Leakage Comparison The refrigerant charge for the secondary l oop 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 sign ificant amount of energy in a supermarket. The refrigerant charge for the multiplex syst em 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 insta llation 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 rec harge req u ired by the secondary loop system since its start on December 18, 2001. The records for the refrigerant charge additi on for the multiplex store are very limited, because of a change in site maintenance cont ractor 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 ve ry limited and may not be representative of the true refrigerant leakage of this multip lex 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

D a t e R e f r i g e r a n t P o u n d s D a t e R e f r i g e r a n t P o u n d s

9 / 1 / 2 0 0 2 5 0 7 1 0 0 2 / 2 1 / 2 0 0 3 4 0 4 A 8 9 / 4 / 2 0 0 2 5 0 7 1 0 0 3 / 1 4 / 2 0 0 3 R 2 2 6 0 4 / 1 8 / 2 0 0 3 5 0 7 2 7 5 4 / 2 5 / 2 0 0 3 R 4 0 4 A 4

T o t a l 4 7 5 7 2

The annual refrigerant leak rate from each site can be determined using the methods provided by the USEPA (9). The percent of to tal refrigerant charge that is lost through leakage is found from:



### 100 additions t refrigeran last since days days 365 charge t refrigeran total added t refrigeran lb Rate Leak Annual % × × =

6061 Table 24 shows the results for the two stores. With the exception of the addition of field test instrumentation, the only refrigerant c harge 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 ra tes 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 mu ltiplex refrigeration system. In terms of absolute refrigerant loss, the annual leak ra te 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 3706263 4.0 Conclusion & Recommendations
The specific technical objectives of the projec t were to design and test a secondary loop refrigeration system, which:

• Consumes approximately 14% less electricity th an 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 multiple x with evaporative condensing, suggesting that the largest portion of the savings c an 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 re frigeration 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 de sign. The annual savings achieved by each of these features compared to a more conven tional secondary loop refrigeration system

were:

• Multiple parallel brine pumps – Estimate d 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 us e of a minimum delta T between the display case discharge air and refrigeration satura ted 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 sy stem as tested had an initial refrigerant charge of 1,400 lb., which is considerably larg er than the stated goal. The reason for the larger refrigerant charge is the secondary lo op 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 th e 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 comp lete 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 si ngle-deck meat cases showed the multiplex and secondary loop systems maintained the pr oduct 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 ra ck SST in order to achieve this condition (2.3°F for multiplex vs. 14.1°F for secondary l oop). For the multi-deck produce cases, the multiplex and secondary loop systems operated at similar rack SST values. The resulting average product temperature was a pproximately 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 progra m 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 supermar ket secondary loop refrigeration systems.5.0 References

1. Sand, James R, Steven K. Fischer, Van D. Bax t er , 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 Pr otection Technologies, Baltimore, MD, November 12-13, 1997.

4. Walker, D.H., Development and Demo nstration of an Advanced Supermarket Refrigeration/HVAC System, Foster-Miller, Inc., Subcontract Nu mber 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 Technolo gies, 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., “Dev elopment of an Evaporative Condenser for Northern Climates,” Report prepared for Niagara Blower Co., Buffalo, NY, April, 1997.

9. Compliance Guidance For Industrial Proc ess Refrigeration Leak Repair Regulations Under Section 608 of The Clean Air Act , The Chemical Manufactur ers Association and The Environmental Protection Agency, October 1995.

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