IEA Annex 26: Advanced 

Supermarket Refrigeration/Heat 

Recovery Systems 

Final Report Volume 1 – Executive Summary 

Canada 

Denmark 

Sweden 

United Kingdom 

United States (Operating Agent)

Compiled by 

Van D. Baxter 

Oak Ridge National Laboratory 

Oak Ridge, TN, USA 

April, 2003
ii

List of Contributors 

This volume of the IEA Annex 26 final report contains a detailed summary of the Annex 

activities including principal conclusions from each participant. The following persons provided 

the information for this report. 

Canada - Daniel Giguère 

Natural Resources Canada 

CANMET Energy Technology Centre –Varennes (CETC) 

1615 Lionel-Boulet Blvd., Box 4800 

Varennes, Québec J3X 1S6 

Canada 

1-450-652-5015 phone 

1-450-652-5177 fax 

daniel.giguere@nrcan.gc.ca

Vasile Minea 

Hydro Québec, Institut de recherché

Laboratoire des technologies de l'énergie (LTE) 

600, avenue de la Montagne 

Shawinigan, Québec G9N 7N5 

Canada 

1-819-539-1400, ext 1507 phone 

1-819-539-1409 fax 

minea.vasile@lte.ireq.ca

Denmark - Hans Jørgen Høgaard Knudsen 

 Department of Mechanical Engineering 

 Technical University of Denmark 

 Nils Koppels Allé 

 DTU – Building 402 

 DK-2800 Kgs. Lyngby, Denmark 

 45-45-254121 phone 

 45-45-935215 fax 

 hk@mek.dtu.dk

Sweden - Per Lundqvist & Jaime Arias 

The Royal Institute of Technology 

Department of Energy Technology 

Div. of Applied Thermodynamics 

and Refrigeration 

S-100-44 Stockholm, Sweden 

46-8-790-7452 phone (Lundqvist) 

46-8-790-8642 phone (Arias) 

46-8-20-3007 fax 
iii

perlundq@egi.kth.se

jaime@thermo.kth.se

UK - Alison Crompton 

 Faber Maunsell 

Marlborough House 

Upper Marlborough Road 

St Albans, Hertfordshire AL1 3UT 

United Kingdom 

44-20-8784-5586 phone 

44-20-8784-5562 fax 

alison.crompton@fabermaunsell.com

John Palmer 

Faber Maunsell 

Beaufort House 

94/96 Newhall Street 

Birmingham B3 1PB 

United Kingdom 

44-121-262-6750 phone 

44-121-262-1999 fax 

john.palmer@fabermaunsell.com

US - Van D. Baxter 

Steven K. Fischer 

Oak Ridge National Laboratory 

Box 2008, Bldg 3147, MS-6070 

Oak Ridge, TN 37831-6070 

USA 

1-865-574-2104 phone (Baxter) 

1-865-574-2017 phone (Fischer) 

1-865-574-9329 or 9338 fax 

vdb@ornl.gov

fis@ornl.gov

David H. Walker 

Foster-Miller, Inc. 

350 Second Avenue 

Waltham, MA 02451 

1-781-684-4237 phone 

1-781-890-3489 fax 

dwalker@foster-miller.com
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Table of Contents 

List of Contributors ........................................................................................................................ ii 

Table of Contents........................................................................................................................... iv 

List of Figures................................................................................................................................ vi 

List of Tables ................................................................................................................................. vi 

Introduction to IEA Annex 26 ......................................................................................................1 

Background ....................................................................................................................................2 

General summary of system technology approaches to reduce supermarket TEWI..............5 

Secondary loop type systems...................................................................................................6 

Estimates of cost premium ..................................................................................................9 

Estimates of TEWI reduction ............................................................................................10 

Distributed compressor refrigeration system......................................................................10 

 Estimates of cost premium ................................................................................................12 

Estimates of TEWI reduction ............................................................................................12 

Low charge multiplex systems..............................................................................................12 

 Estimates of cost premium ................................................................................................14 

Estimates of TEWI reduction ............................................................................................14 

Advanced self-contained systems..........................................................................................................14 

 Estimates of cost premium ................................................................................................15 

Estimates of TEWI reduction ............................................................................................15 

HVAC integration approaches .............................................................................................15 

Country report summaries..........................................................................................................18 

Canada - CANMET Energy Technology Centre –Varennes (CETC–Varennes) of 

Natural Resources Canada..............................................................................................18 

Canada - Hydro-Quebec Research Institut, Laboratoire des technologies de l’énergie

(LTE) .................................................................................................................................20 

Total Heat Recovery (”System1-phase1”)........................................................................21 

Total Heat Recovery (”System2–phase1”)........................................................................21 

High Speed Defrost............................................................................................................23 

Liquid Pumping and Injection ...........................................................................................23 

Geothermal Cooling...........................................................................................................24 

Further Developments .......................................................................................................24 

Denmark .................................................................................................................................25 

Sweden ....................................................................................................................................27 

United Kingdom.....................................................................................................................33 

Secondary Systems.............................................................................................................33 

Defrost................................................................................................................................34
v

CHP/CCHP........................................................................................................................35 

Case and Store Environment Interaction...........................................................................36 

UK Overall Conclusions....................................................................................................37 

United States...........................................................................................................................37 

System analysis summary...................................................................................................38 

Payback estimates..............................................................................................................40 

Field test summary.............................................................................................................41 

Supermarket energy use screening level model ........................................................................42 

Overall Summary.........................................................................................................................42 

Canada ....................................................................................................................................43 

Denmark .................................................................................................................................43 

Sweden ....................................................................................................................................44 

United Kingdom.....................................................................................................................44 

United States...........................................................................................................................45 

Summary comments..............................................................................................................45 

References.....................................................................................................................................46 

Acknowledgements ......................................................................................................................47 

Appendix A - Case Study Summaries........................................................................................48 

Canada - LTE.........................................................................................................................48 

Denmark .................................................................................................................................62 

United States...........................................................................................................................66 
vi

List of Figures 

Fig. 1. Layout of a typical modern supermarket.........................................................................3 

Fig. 2. Total equivalent warming impact (TEWI) for low temperature supermarket 

refrigeration in North America. .............................................................................................4 

Fig. 3. Multiplex refrigeration system..........................................................................................5 

Fig. 4. Elements of the secondary loop refrigeration system......................................................7 

Fig. 5. Propane/carbon dioxide cascade refrigeration system....................................................8 

Fig. 6. Distributed compressor refrigeration system schematic. .............................................11 

Fig. 7. Low-charge multiplex system compressor rack concept 1. ..........................................13 

Fig. 8. Low-charge multiplex system compressor rack concept 2. ..........................................13 

Fig. 9. Roof top heat pump unit. .................................................................................................17 

Fig. 10. Plate heat exchangers (HX) for heat pump evaporators and refrigeration 

desuperheaters and subcooler...............................................................................................17 

Fig. 11. Results from CyberMart................................................................................................29 

Fig. 12. Indoor temperature, outdoor temperature and indoor relative humidity in 

Hedemora and Sala................................................................................................................30 

Fig. 13. Temperatures, relative humidity and compressor power in Sala..............................31 

Fig. 14. Temperatures in a deep freeze cabinet and indoor and outdoor temperatures 

during four days in July in Hedemora.................................................................................32 

Fig. 15. Refrigeration energy usage in The United Kingdom by end use. ..............................33 

Fig. 16. Example of freezer case type tested by UK team in environmental chamber to 

determine impact of store temperature, relative humidity and ventilation mode 

(conventional or displacement) on performance.................................................................36 

Fig. A1. Comparison of energy use of Fakta Beder store (with cascade refrigeration system) 

to that of eight other Fakta stores (with conventional R404A DX refrigeration systems); 

July 2000 through February 2002. .......................................................................................65 

List of Tables 

Table 1. Predicted Energy Consumption for Low-Charge Refrigeration Systems................39 

Table 2. Total Equivalent Warming Impact (TEWI) for Supermarket Refrigeration..........40 
1

Annex 26 Final Report 

Advanced Supermarket Refrigeration/Heat Recovery Systems 

Executive Summary 

Introduction to IEA Annex 26 

With increased concern about the impact of refrigerant leakage on global 

warming, a number of new supermarket refrigeration system configurations requiring 

significantly less refrigerant charge are being considered. In order to help promote the 

development of advanced systems and expand the knowledge base for energy-efficient 

supermarket technology, the International Energy Agency (IEA) established IEA Annex 

26 (Advanced Supermarket Refrigeration/Heat Recovery Systems) under the IEA 

Implementing Agreement on Heat Pumping Technologies. Annex 26 focuses on 

demonstrating and documenting the energy saving and environmental benefits of 

advanced systems design for food refrigeration and space heating and cooling for 

supermarkets. Advanced in this context means systems that use less energy, require less 

refrigerant and produce lower refrigerant emissions. Stated another way, the goal is to 

identify supermarket refrigeration and HVAC technology options that reduce the total 

equivalent warming impact (TEWI) of supermarkets by reducing both system energy use 

(increasing efficiency) and reducing total refrigerant charge. 

The Annex has five participating countries: Canada, Denmark, Sweden, the United 

Kingdom, and the United States. The working program of the Annex has involved 

analytical and experimental investigation of several candidate system design approaches 

to determine their potential to reduce refrigerant usage and energy consumption. 

Advanced refrigeration system types investigated include the following: 

distributed compressor systems – small parallel compressor racks are located in close 

proximity to the food display cases they serve thus significantly shortening the 

connecting refrigerant line lengths; 

secondary loop systems – one or more central chillers are used to refrigerate a 

secondary coolant (e.g. brine, ice slurry, or CO2) that is pumped to the food display 

cases on the sales floor; 

self-contained display cases – each food display case has its own refrigeration unit; 

low-charge direct expansion – similar to conventional multiplex refrigeration systems 

but with improved controls to limit charge. 

Means to integrate store HVAC systems for space heating/cooling with the 

refrigeration system have been investigated as well. One approach is to use heat pumps to 

recover refrigeration waste heat and raise it to a sufficient level to provide for store 

heating needs. Another involves use of combined heating and power (CHP) or combined 

cooling, heating, and power (CCHP) systems to integrate the refrigeration, HVAC, and 

power services in stores. Other methods including direct recovery of refrigeration reject 

heat for space and water heating have also been examined. 
2

A workshop on advanced supermarket refrigeration was held October 2-4, 2000 in 

Stockholm. Reports were made by Annex participants and other invited experts on the 

status of refrigeration research and development activities in the US and Europe. The 

workshop proceedings (in cd-rom format) are available from the IEA Heat Pump Centre 

(Lundqvist, ed, 2000). 

This report provides an account of the results of the work programs of the Annex 

participants. This volume gives an executive summary of the results. Full details are 

given in the individual country reports in volume 2. 

Background 

Supermarkets are one of the most energy-intensive types of commercial buildings. 

Significant energy is used to maintain chilled and frozen food in both product display 

cases and storage refrigerators. The refrigeration systems also produce a large amount of 

rejected heat that can be recovered and used by heat pumps or other equipment to provide 

space and water heating for store requirements. There is a wide range in size of 

supermarkets among the Annex 26 countries. In Europe stores range in size from about 

500 m2

 to 3000 m2 or somewhat larger. Stores are typically larger in Canada and the US 

ranging from a minimum of about 1000 m2 to 10000 m2. Plant capacities range from 30-

60 kW for small markets to over 400 kW for the largest stores. Similarly, annual energy 

use ranges from about 100,000 kWh/y for the smaller stores to 1.5 million kWh/y or 

more for the largest. 

Refrigeration is the largest component of supermarket energy use, accounting for 

half or more of the store total. Perishable products must be kept refrigerated during 

display and for storage. Compressors and condensers account for 60-70% of 

refrigeration energy consumption. The remainder is consumed by the display and storage 

cooler fans, display case lighting, evaporator defrosting, and for anti-sweat heaters used 

to prevent condensate from forming on doors and outside surfaces of display cases. 

Figure 1 shows a representative layout for a supermarket showing refrigerated 

display cases and storage areas located generally around the store perimeter. The most 

commonly used refrigeration system for supermarkets today is the multiplex direct 

expansion (DX) system. All display cases and cold store rooms use direct expansion airrefrigerant coils that are connected to the system compressors in a remote machine room 

located in the back or on the roof of the store. This requires thousands of meters of pipe 

with case connections that have historically been designed for ease and rapidity of service 

rather than low leakage. This practice is changing for new supermarkets with more 

emphasis on reducing leakage. Heat rejection is usually done with air-cooled condensers 

because these are the least cost to install and maintain. Evaporative condensers can be 

used as well and will reduce condensing temperature and system energy consumption. 

However, they carry the burden of increased maintenance effort and cost. In either case, 

system controls are usually set to allow the condensing temperature to float with the 

outdoor dry bulb (or wet bulb) temperature, usually to a minimum level of around 21°C 

(about the lowest condensing temperature for reciprocating compressors which are the 

most common type used in supermarkets). 
3

The amount of refrigerant needed to charge multiplex DX systems is very large - 

typically 4-5 kg/kW of refrigeration capacity. The large amount of piping and pipe joints 

required can also result in large refrigerant losses – historically 30% or more of the total 

charge annually. New systems can achieve annual refrigerant leakage rates of around 

15% or somewhat lower (Sand, et al, 1997). The large refrigerant charge and high loss 

rate for multiplex direct expansion refrigeration systems results in high values of TEWI 

(total equivalent warming impact) with direct refrigerant loss impact accounting for about 

half of the total as shown in Figure 2. 

Figure 1 – Layout of a typical modern supermarket 

Figure 3 shows the major elements of a multiplex refrigeration system. Multiple 

compressors operating at the same saturated suction temperature are mounted on a skid, 

or rack, and are piped with common suction and discharge refrigeration lines. Using 

multiple compressors in parallel provides a means of capacity control, since the 

compressors can be selected and cycled as needed to meet the refrigeration load. 
4

Figure 2 – Total equivalent warming impact (TEWI) for low temperature 

supermarket refrigeration in North America 
5

Figure 3 – Multiplex refrigeration system 

General summary of system technology approaches to reduce supermarket TEWI 

Part of the UK’s contribution to the Annex was a detailed review of existing 

technologies and new technologies for supermarkets (see UK country report, Volume 2). 

The bottom line from this state of the art review of advanced supermarket refrigeration is 

that there are a number of areas in which significant energy savings may be realized and 

environmental impact reduced. From the overall design of the energy supply, from CHP 

and CCHP, through to the interaction of the store environment and the cabinets, savings 

can be made. These factors offer a wide range of opportunities to bring about reductions 

in the energy use and TEWI of supermarket refrigeration systems. 

The advanced supermarket refrigeration systems described in this section are 

designed primarily to reduce the amount of refrigerant needed for operation and thereby 

the amount of annual refrigerant loss. This is generally accomplished by use of 

secondary refrigerants in a secondary loop type system, reduction of refrigerant 

connecting line length between compressors and display cases, improved refrigerant 

management and flow controls, or a combination of these approaches. Advantages and 

disadvantages of these systems with respect to energy efficiency and possible cost 

implications are briefly discussed in this section also. These issues are dealt with in more 
6

comprehensive fashion in the individual country reports. Four generic system types have 

been investigated as part of the Annex 26 work effort. They are secondary loop, 

distributed, low-charge multiplex, and advanced self-contained systems. 

Secondary loop type systems

Secondary loop refrigeration systems can take many forms, but they generally 

employ one or more chillers to refrigerate a secondary fluid that is then pumped to the 

display cases and storage rooms. Figure 4 shows the elements of one secondary loop 

approach. In this case the chillers, similar in configuration to multiplex compressor 

racks, provide chilled secondary refrigerant for the cases. The discharge of the 

compressors is to a common manifold and the discharge gas is piped to a remote 

condenser, normally located on the roof above the machine room. The use of evaporative 

condensing can produce the lowest average condensing temperature with lower fan 

energy than seen with air-cooled condensers, thus minimizing overall energy 

consumption especially in warmer climate areas. Primary refrigerant charge requirement 

can be reduced to about 10-15% of that needed for conventional direct expansion system. 

A variation on the system shown in Figure 4 is to use a secondary fluid on the heat 

rejection (condenser) side of the chiller as well as on the refrigeration (evaporator) side. 

This approach can reduce primary refrigerant requirements to less than 5% of direct 

expansion but would introduce an additional secondary heat exchanger thermodynamic 

energy penalty to the system and increase energy usage. 

Secondary loop systems have features that tend to improve the efficiency of the 

primary system. These include close coupling of the compressors to the chiller 

evaporators, and the ability to subcool the primary refrigerant with the secondary fluid 

(brine) and use the warmed brine to defrost the case heat exchangers. 
7

Figure 4 - Elements of the secondary loop refrigeration system 

For large stores secondary loop systems generally use at least two separate 

secondary fluid loops and chillers – one each for the low temperature (frozen food) cases 

and storage rooms and medium temperature (chilled food) cases and rooms. Using only 

two fluid loops for refrigeration, however means that all display cases and storage coolers 

must operate with these two temperatures. The use of multiple secondary fluid loops 

with temperatures more closely matching the case air temperature requirements can 

improve energy efficiency because it raises the effective average evaporator temperature 

of the system. The US country report includes an analysis of such a system using four 

different fluid loops. It must be stated however, that each loop will require a separate 

chiller and controls and that will increase the system cost. 

 Another approach to secondary systems is the cascade-type system. Figure 5 

schematically illustrates one example of a cascade system. The high temperature loop 

(using propane as the refrigerant in this illustration) removes heat from the low 

temperature loop (using CO2) condenser and also chills a brine (propylene glycol), which 

is circulated to the chilled food cases. The CO2 refrigerates the frozen food cases through 

direct expansion evaporators. Other combinations of refrigerants are possible for this 

system concept. 
8

Figure 5 - Propane/carbon dioxide cascade refrigeration system (from Danish 

country report) 

The following is a listing of several candidate fluids for use in secondary loop 

systems that have been investigated in prior studies (Melinder, 2000; Kazachki, et al, 

1997; Hrnjak, 1996; Hrnjak, 2000; Horton, 2002; Sherwood, 1999). Under Annex 26 

some detailed property measurements and comparisons were conducted by the UK team 

(see country report in Volume 2). 

Ethylene glycol/water 

Propylene glycol/water 

Potassium Formate/water 

 Pekasol 50 

 Freezium 

 Hycool 

Trichloroethylene 

Inhibited alkali ethanate solution 

 Tyfoxit 

Hydrofluoroether 

 HFE-L-13938 

 HFE-7100 

Cyclohexene 

 D-Limonene 

Polydimethilsiloxane (Silicon Oil) 

 Syltherm 

 Dowtherm 

Synthetic Isoparaffinic Petroleum Hydrocarbons 

 Therminol 
9

No single one of these fluids can be considered ideal for all secondary refrigerant 

applications. Those most widely used in supermarket systems appear to be propylene 

glycol/water solutions for medium temperature loops and potassium formate/water 

solutions for low temperature loops. However, some analytical studies indicate thermal 

and pressure droop advantages for HFE-7100 over potassium formate solutions for low 

temperature applications -- below about -20 °C (Sherwood, 1999; Horton, 2002). Liquid 

CO2 is also seeing increasing use as a secondary fluid for both temperature levels 

particularly in Europe. Means must be considered to deal with internal tube pressures in 

CO2 systems in the event of power outages or other failures. 

Estimates of cost premium. A number of studies have been conducted to estimate 

cost premiums with the added cost estimates ranging from none to about 40%. A study 

by Sherwood (1999) compared installed costs for a secondary loop and a multiplex direct 

expansion system for a low-temperature system typical of North American supermarkets 

– about 80 kW total cooling capacity. The secondary system used NH3 in the primary 

loop and HFE-7100 as the brine while the DX used R-404A. Installation cost (material 

and labor) for the DX system was the actual figure taken from the refrigeration schedule 

for the store. Costs for the secondary loop system were estimated based on interviews 

with supermarket industry professionals. This study considered the total cost of the 

system (including cases, piping, refrigerant, brine, and labor in addition to the compressor 

rack or primary chiller) with the exception of the condenser subsystem. It was 

determined that installed costs for both approaches were equal – i.e., 0% cost premium 

for the secondary approach. 

 Vana (2000) noted in discussions at the Annex 26 workshop that installation cost 

premiums for secondary loop approaches (using R404A or R507 as primary refrigerant 

and propylene glycol or potassium formate brines for secondary loops) were about 15% 

for typical US markets. He also noted that maintenance costs for the secondary system 

should be less. Horton (2002) conducted an operating (energy and maintenance) cost 

comparison between secondary loop (NH3 primary/HFE-7100 secondary) and multiplex 

DX (R22 medium temperature/R404A low temperature) systems and concluded that 

maintenance costs were 25% lower and operating costs 15% lower for the secondary 

approach. All of the operating savings were attributed to the high efficiency of the NH3

primary system – high enough to overcome the secondary pumping power and heat 

transfer penalties in this analysis. 

 A cost and payback analysis is included in the US country report for secondary 

loop (R507 primary/ propylene glycol and potassium formate secondary brines) and 

multiplex DX (R22/R404A) systems serving a supermarket with a 330 kW refrigeration 

load. The DX system was assumed to have standard air-cooled condensers while the 

secondary system was assumed to have evaporative condensers (to minimize energy use). 

The secondary system had an estimated installation cost premium of $147,000 US that 

was based on interviews with engineering departments of two supermarket companies. 

Payback in this case ranged from 8 to 17 years depending on climate conditions. More 

details are given in the US report (in Volume 2). 

 The Danish country report (Volume 2) compares installation costs and operating 

efficiencies for a cascade system based on that shown in Figure 5 and R404A DX 

systems. A test system installed in a small store (30 kW load) was estimated to cost 
10

about 20% more than a traditional DX system and to have about the same energy 

efficiency. With more experience for installers the premium is estimated to drop to under 

15%. For larger systems (90 kW) the premium is estimated to be 10% or less. 

Estimates of TEWI reduction. TEWI reduction estimates from the investigations 

carried out under Annex 26 range from <10% to about 60% compared to the multiplex 

DX/air-cooled condensing baseline. The wide range in TEWI estimates is attributable to 

assumptions made about the overall system design (compressor type, heat rejection 

approach, selection of secondary fluid, etc.) and operation. Secondary loop systems must 

be designed and operated carefully with an eye to minimizing the inherent secondary heat 

exchange and pumping power penalties in order to achieve the greatest TEWI benefit. 

Distributed compressor refrigeration system

Figure 6 shows a diagram of the distributed compressor refrigeration system. The 

big difference between this system and conventional multiplex systems is that several 

small compressor racks are located in cabinets that are distributed throughout the store 

and close-coupled to the display case lineups or storage rooms they serve. With this 

approach the long lengths of piping needed to connect the cases with large remote 

compressor racks in a machine room are eliminated. The cabinets may be placed either at 

the end of a case lineup on the sales floor or behind the cases around the perimeter of the 

store. 

 With this arrangement, the saturated suction temperature (SST) employed for 

each compressor cabinet can closely match the evaporator temperature of the display 

cases and walk-in coolers to which it is connected. This is not always possible with 

conventional multiplex systems, since a single rack will often provide refrigeration to 

display cases with three or four different evaporator temperatures. Each compressor rack 

of a multiplex system must operate at a SST value that will satisfy the temperature 

requirements of all display cases connected to it. The better temperature matching seen 

with distributed refrigeration can benefit the energy consumption of the overall system. 

 The refrigerant charge requirement for the distributed system is much less than for 

multiplex systems. The reduction in charge is due to the shortening of the suction and 

liquid lines to the display cases. If a secondary fluid loop is used for heat rejection (as 

shown in Figure 6) then the refrigerant heat rejection piping to a remote condenser and 

it’s associated charge is eliminated as well. With a secondary loop for heat rejection, the 

refrigerant charge required for a distributed system would be about 30-35% of that 

required for multiplex systems. If separate rooftop condensers are used for each cabinet, 

the total charge requirement will be about 50-60% that of multiplex systems. A 

secondary heat rejection loop will result in higher condensing temperatures and energy 

consumption than if direct roof top condensers are used. 
11

Figure 6 - Distributed compressor refrigeration system schematic 

Each compressor cabinet in a distributed system is similar to a multiplex rack 

only smaller. All necessary electrical and piping connections are provided within the 

cabinet, such that the only field connections are the refrigerant liquid and suction lines, 

fluid inlet and outlet piping for heat rejection, and electric service wiring. Multiple 

compressors of several sizes are employed, which are piped in parallel so that multiplex 

operation can be used to match the refrigeration capacity to the load. Usually, three to 

five compressors are installed in each cabinet. The cabinets are equipped with discharge 

and suction manifolds for parallel piping of the compressors. The suction manifold can 

be divided so that multiple suction temperatures can be provided from a single cabinet. 

Current distributed refrigeration systems generally employ scroll compressors, 

because of the very low noise and vibration levels encountered with this type of 

compressor. These characteristics are necessary if the compressor cabinets are located in 
12

or near the sales area. Current refrigeration scroll compressors are typically less efficient 

at evaporating and condensing temperatures of interest for supermarket refrigeration than 

the reciprocating types used in most conventional systems. However, they do have 

several features that can offset this efficiency disadvantage in system applications and 

potentially yield system efficiency improvements. For instance, scroll compressors have 

no valves, and, therefore, can be operated at significantly lower condensing temperatures 

than can reciprocating compressors. The lowest condensing temperature possible is at a 

suction-to-discharge pressure ratio of two (2), which means for supermarket systems that 

the lowest condensing temperature possible is on the order of 10 – 15 °C. 

The close coupling of the display cases to the distributed refrigeration cabinets 

has other ramifications to energy consumption. The shorter suction lines mean that the 

pressure drop between the case evaporator and the compressor suction manifold should 

be less than that seen with conventional multiplex systems, which means that the 

compressors’ SST will be closer to the display case evaporator temperature. Typically, 

the SST of conventional multiplex racks will be 1-2.5 °C less than the case evaporator 

temperature. SST values for the distributed system will be about 0.5-1 °C less than the 

case evaporator temperature. The shorter suction lines also mean that less heat gain to 

the return gas should be experienced. The cooler return gas has a higher density and 

results in higher compressor mass flow rates, which means that less compressor on time 

is needed to satisfy the refrigeration load. The return gas temperature rise expected for 

distributed refrigeration systems is normally on the order of 3 to 9 °C, depending on the 

distance between the compressor cabinet and the display cases and the evaporator 

temperature of the display cases. A greater return gas temperature rise is seen in low 

temperature systems than is seen in medium temperature. In comparison, the return gas 

temperature rise typically seen with multiplex systems falls between 20 and 35 °C, due to 

the longer length of suction lines employed. The liquid temperature can also be 

adversely affected by line length if subcooling is employed. Heat gain to a subcooled 

liquid line will result in a rise in liquid refrigerant temperature before reaching the display 

cases. 

Estimates of cost premium. The US country report includes cost and payback 

estimates for this system. Installed cost for a distributed (using a secondary fluid loop 

and evaporative cooling tower for condenser heat rejection) was estimated at $60,000 

(about 15%) over a baseline multiplex DX system (R22/R404A with air-cooled 

condensers) for a supermarket with a 330 kW refrigeration load. Payback ranged fro 

about 3 to 7 years depending on location. 

Estimates of TEWI reduction. Potential TEWI reductions from this work range 

from a high of about 60% (analytical estimate for well designed system) to 40% (from a 

field test). As in the secondary loop case, careful attention to design and operation is 

essential to achieving maximum benefits. 

Low charge multiplex systems.

Several approaches have been taken by refrigeration system manufacturers to 

reduce the refrigerant charge required by traditional multiplex system compressor racks. 

These approaches reduce the overall system refrigerant charge but retain the long 

connecting lines between compressors and display cases. Figures 7 and 8 illustrate two 

such low charge multiplex rack options. 
13

Figure 7 – Low-charge multiplex system compressor rack concept 1

Figure 8 – Low-charge multiplex system compressor rack concept 2 
14

 The first system (Figure 7) limits refrigerant charge by minimizing the refrigerant 

inventory in the receiver during normal operation (the receiver’s primary function in this 

design is to provide refrigerant storage during system servicing). Total charge required 

for this system is expected to be about 2/3 that of a conventional multiplex system. This 

charge control approach also can yield some energy savings because it can allow the 

system compressors to operate at lower condensing temperatures than would normally be 

the case. The control valve and ambient temperature sensor are intended to maintain a 

constant temperature difference between the condenser outlet liquid and the ambient. 

Liquid refrigerant flow through the valve is vaporized by heat exchange with the 

discharge manifold. Any refrigerant not vaporized is stored in the receiver. Saturated 

discharge temperatures can be as low as about 5 °C for low temperature racks and about 

15 °C for medium temperature racks. Normally reciprocating compressors are limited to 

minimum condensing temperatures of about 21 °C and scroll compressors to a minimum 

of about 15 °C. 

 The second charge control approach (Figure 8) tries to reduce charge to the 

minimum needed for correct operation of the system evaporators. A mixture of vapor 

and liquid refrigerant is sent to each display case lineup and storage room. Operation of 

the evaporators is done using a combination of relay and balancing valves to control 

refrigerant flow and evaporation. Refrigerant mixture flow to each case lineup is 

controlled at the compressor rack. Total refrigerant charge required by this approach is 

expected to be as little as 30% of that needed by conventional multiplex systems and the 

energy efficiency is claimed to equal that of conventional systems. Heat reclaim can be 

done with a heat exchanger and fluid loop as shown in Figure 8. 

Estimates of cost premium. System economic analysis in the US country report 

(Volume 2) assumed the installed cost for this system type to be the same as for baseline 

multiplex DX systems (if both use the same type condenser). 

Estimates of TEWI reduction. TEWI could be reduced by from 20% to over 40% 

by these systems (compared to the baseline) if the charge reduction and energy saving 

targets can be achieved in real systems. 

Advanced self-contained systems.

A self-contained refrigeration system consists of display cases or storage coolers 

each having their own condensing units. Self-contained systems are presently used in 

supermarkets for a limited number of cases, where the cases are in a location inaccessible 

to the central refrigeration piping. An example is a refrigerated beverage case placed at 

the cash registers for spot sales. Self-contained units are also employed as add-on cases 

or for temporary display of special sales items. Present self-contained display cases use 

small reciprocating compressors and air-cooled condensers. Heat is rejected directly into 

the sales area. Only a limited number of self-contained units of this type can be 

employed before noise and heat rejection levels interfere with store operation. Problems 

of this type caused store designers in the past to go to the remote machine room approach 

now used in most supermarkets. 

The self-contained system approach could be attractive for reduction of 

refrigerant charge. It is estimated that the total charge for a supermarket could be 
15

reduced to about 5-10% that of conventional multiplex systems if self-contained systems 

were used for all refrigeration needs. 

An advanced self-contained system could be formulated using water-cooled 

condensers and a fluid loop for heat rejection to eliminate concerns of heat rejection in 

the store sales area. The compressor noise issue will still be a factor, but can be 

addressed by the use of scroll compressors. Until recently, scroll compressors were 

available only in a vertical configuration, which was not suitable for placement in display 

cases. Now, horizontal scroll compressors have been introduced, which could be 

employed for this purpose. These horizontal scrolls are capable of continuous unloading 

for capacity control and maintenance of a suction pressure set point. 

Estimates of cost premium. No referencable cost studies have been found for this 

system approach. The system can be assumed to be an extreme case of the distributed 

compressor approach, however, and therefore the cost premium over multiplex DX could 

be assumed to be similar as well. 

Estimates of TEWI reduction. No definite estimate has been made for this system 

type. However, as above it could potentially achieve the same TEWI reductions as the 

distributed system if efficient enough small compressors and other system components 

can be found or developed to keep the system energy use on a par with or better than that 

of the baseline DX system. 

HVAC integration approaches

The large amount of heat rejected by the refrigeration system in a typical 

supermarket offers an attractive resource for use in store space heating. Integration 

approaches examined during the course of this Annex have included straightforward heat 

reclaim, use of heat pumps integrated with the refrigeration system, and use of CHP or 

CCHP systems. 

 During a meeting of the Annex in Canada a supermarket with a heat reclaim 

system was visited. The store had two R22 multiplex refrigeration compressor racks (one 

each for medium temperature and low temperature loads) with two condensers in series. 

In winter, primary condensing is done in a plate heat exchanger with water on the other 

side. The water from the medium temperature rack heat exchanger is piped to coils in the 

HVAC air ducts and used for store heating. Heat from the low temperature rack is used 

for store water heating and for reheat for the store dehumidification system in summer. 

An auxiliary boiler is included in the HVAC system to cover times when heat from the 

refrigeration condenser cannot meet the store heating load. According to the installer, the 

propane boiler was not needed during the first winter (01/02) of operation. The 

refrigeration system saturated discharge temperature is allowed to float to a minimum of 

about 27 °C under most conditions. In periods of high space heat demand, the 

refrigeration system condensing temperature is raised as necessary (to a maximum of 35 

°C with some negative impact on refrigeration compressor energy use) to meet the store 

heating demand. In summer the refrigeration system uses a conventional rooftop aircooled condenser for heat rejection. Also according to the installer this system has about 

3% higher cost than a conventional refrigeration/HVAC system (Bastrash, 2002). 

The use of heat pumps represents an excellent way to utilize refrigeration reject 

heat for space heating. Two different heat pump integration approaches have been 

examined by individual Annex participants. One involves direct integration of the heat 
16

pump evaporator(s) and refrigeration system condenser(s) and the second utilizes watersource heat pumps (WSHP) where water-cooled condensers and water/glycol loops are 

used for refrigeration system heat rejection. The heat pumps can be installed in the 

glycol/water loop and use the rejected heat to provide space heating. Either method 

enables reclamation of a very large portion of the reject heat without requiring elevation 

of the condensing temperature of the refrigeration system as can happen with 

conventional heat reclaim approaches. Refrigeration system energy savings achieved by 

low head pressure operation can be realized along with the energy benefits seen through 

heat reclaim. 

An example of the first type was also seen during the Annex meeting in Canada. 

The refrigeration system is a traditional R22 multiplex system with two racks (low 

temperature and medium temperature) and both used a common set of two condensers in 

series. Figures 9 and 10 illustrate this heat pump integration approach. In winter, the 

discharge gas from the refrigeration compressors goes first through three plate heat 

exchangers which serve to desuperheat and precondense the gas. Final condensing is 

accomplished in a conventional rooftop air-cooled condeser. The other side of the three 

heat exchangers serve as the evaporators for three rooftop heat pumps which supply 

space heating to the store. A fourth rooftop heat pump is similarly integrated with the 

liquid line exiting the air-cooled condenser via a fourth plate heat exchanger. This heat 

pump subcools the liquid refrigerant leaving the refrigeration system condenser and uses 

the recovered heat for store space heating. In summer, the refrigeration system uses the 

air-cooled condenser exclusively and the heat pumps use their own condensers to reject 

heat from space cooling loads. The refrigeration system also includes a liquid refrigerant 

pump to boost the pressure of the refrigerant before it is sent to the cases. The pump is 

included because with the heat pump integration the refrigeration system head pressure is 

allowed to float to about 15 °C minimum saturated discharge temperature. According to 

the installer the system cost for that store is about 8% more than what a conventional 

refrigeration and HVAC system would have cost (Kantchev, 2002). Some details of this 

system are included in Canada’s country report where it is identified as the LMP system. 
17

Figure 9 – Roof top heat pump unit 

Figure 10 – Plate heat exchangers (HX) for heat pump evaporators and 

refrigeration desuperheaters and subcooler
18

The use of WSHPs in conjunction with water-loop heat rejection for the 

refrigeration system was examined by the US (see US country report, Volume 2). An 

analysis was performed for a supermarket HVAC system where conventional rooftop airconditioner/gas heating units and water-source heat pumps were examined and compared. 

The combination of a water-cooled distributed compressor refrigeration system with 

water-source heat pumps was predicted to save about 10-20% in overall operating 

(refrigeration plus HVAC energy) costs depending on local climate and utility rates when 

compared to the baseline air-cooled multiplex DX refrigeration system with conventional 

rooftop HVAC units. The added installation cost for the WSHP approach was estimated 

to be about $25,000 US (about 30%) over the conventional rooftop unit system. 

CHP/CCHP system concepts were studied in detail by the UK team. They 

concluded that CCHP has an improved load match compared to CHP, when used with 

absorption chilling for building cooling. Integral cascade vapor compression systems 

within each display case integrated with the absorption system provide food refrigeration. 

Primary energy savings are around 15%, and there is a TEWI reduction of more than 

50% compared with conventional supermarket systems. 

The impact that store ambient conditions can have on refrigeration system and 

case energy use was illustrated by part of the UK team effort. Environmental chamber 

tests of low and medium temperature case systems show that an increase in ambient 

temperature from 19 to 22 °C increases system energy use by 20%. Increasing RH from 

35 to 50% at an ambient of 22 °C increased the medium temperature case energy use by 

15% but had no effect on the freezer case. 

Country report summaries 

Following are summarized key results from the work programs carried out by 

each country under Annex 26. Full details of these projects can be found in the 

individual country reports. 

Canada - CANMET Energy Technology Centre –Varennes (CETC–Varennes) of 

Natural Resources Canada

The CETC–Varennes contribution to the Annex included a supermarket showcase 

with a fully integrated HVAC/Refrigeration system using secondary fluid on condenser 

and evaporator sides, the organization of an advisory committee, one workshop for the 

Canadian refrigeration industry and two research projects directly related to supermarket 

refrigeration. 

There are approximately 31,000 retail food stores in Canada 7,150 of which are 

considered large supermarkets – with a floor cover area of over 1,000 m2

. Supermarkets 

are among the most energy intensive buildings in the commercial sector with a total 

energy consumption of more than 800 kWh-equivalent/m2

/year (and increasing with 

time). Refrigeration represents about 50% of the electrical energy consumption of a 

supermarket, costing on average $150,000 CND per year for a typical large size 

supermarket (2000 m2

). The typical refrigeration system contains approximately 750 kg 

of synthetic refrigerant equivalent to 2,000 tonnes of CO2 in terms of greenhouse gas 
19

effect. Refrigerant losses are on average, 25% per year of the total charge. The 

refrigeration sector is currently facing several major challenges such as: the legislated 

replacement of synthetic refrigerants that have negative environmental impacts; the 

increasing cost of refrigerants; the poor efficiency of equipment that are not adapted to 

Canadian climatic conditions; new regulations for refrigerated product storage and 

transportation; and a lack of innovation and adequate expertise in Canada to address these 

issues. 

CTEC signed an agreement with a major Canadian supermarket chain to develop 

a supermarket technology showcase in a Montreal area in the end of 2003. The project’s 

objective is to develop innovative integrated HVAC&R technologies and practices 

suitable for Canadian climatic conditions, using compact and hermetic refrigeration 

systems using secondary fluid for the evaporator and condenser sides, and to demonstrate 

them in partnership with a major chain of supermarkets in Canada. These demonstrations 

will provide the opportunity to get a better understanding of their technical requirements 

such as design, installation, commissioning, and performance as well as their nontechnical barriers (knowledge transfer to energy consultants and technical operators, 

costs, procurements, etc.). The selected measures should result in reductions in operating 

and maintenance costs, refrigerant leakage and, indirectly, food spoilage. Secondary fluid 

systems will contain less refrigerant and fewer potential refrigerant leak locations, 

resulting in a substantial reduction in the emission of greenhouse gases. This project is 

part of a strategy to transfer technology and disseminate know-how. 

CTEC organized meetings of an Industry Advisory Panel consisting of 

supermarket and refrigeration sector representatives. The objective of these committees is 

to provide advice for case studies, laboratory experiments or demonstration projects in 

accordance with the Canadian supermarket sector needs. The committee’s invited 

membership includes companies representing retailers, compressor rack manufacturers, 

display case manufacturers, supermarket refrigeration control manufacturers, consulting 

engineers and government agencies. The Industry Advisory Panel participants understand 

and are pleased with the Annex 26 strategy to increase the energy efficiency/heat 

recovery of supermarket refrigeration systems, while reducing the overall refrigerant 

charge and potential leaks. 

More than 85 people attended the first Canadian Workshop on Refrigeration 

Systems on March 11, 2003 in Montreal. The event was sponsored by CETC-Varennes 

and the Office of Energy Efficiency (OEE), in collaboration with Agence de l’efficacité 

énergétique (AEE), Hydro-Québec, Manitoba Hydro and BC Hydro. With a possible 

reduction of 1.4 Mt equivalent CO2/year by 2010, refrigeration in supermarkets has a 

great potential of energy efficiency and greenhouse gas emission reductions. The 

refrigeration industry (equipment manufacturers, food retailers, design engineers, 

mechanical contractors, utilities) sees this as an opportunity. The event allowed the 

identification of proven and emerging refrigeration technologies and the validation of the 

supermarkets’ interests. In addition, six discussion groups shared ideas on how to lift 

barriers for adoption of the energy efficient refrigeration technologies. The information 

collected will provide CETC-Varennes and OEE with new ways to facilitate the adoption 

of the technologies developed 

A laboratory test bench was developed to test novel methods of providing 

refrigeration for the food marketing industry. In the first series of tests, CETC is expected 
20

to develop the knowledge to operate refrigeration systems at minimal head pressure and 

maintaining controllability of refrigeration system using an electronic expansion valve. 

Results from this work will validate or further improved the accuracy of refrigeration 

equipment models and will help to transfer the know-how to refrigeration experts. Some 

technical papers will be published soon. 

During the year 2000, CTEC started a new strategic project scheduled to last four 

years. The main objective is to construct and perform experiments on a liquid CO2

secondary fluid loop test bench. The secondary objectives are development of basic 

know-how (technical documentation, simulation models, experimental data) and 

demonstration of the technologies feasibility to potential partners. CTEC got a CFD 

program (Computational Fluid Dynamic) to model complex problems related to 

refrigeration technologies. A major partner from the carbon dioxide industry has joined 

the research team to realise this project. 

Canada - Hydro-Quebec Research Institut, Laboratoire des technologies de 

l’énergie (LTE)

The contribution of LTE covered the following areas: 

- Development of two improved methods for total heat recovery in cold climates. 

- Development of a high-speed defrost technique. 

- Effective use of liquid pumping and injection. 

- Further research on systems having low refrigerant charges and reduced 

supplementary costs and energy consumption. 

In Canada, supermarkets are one of the most energy-intensive types of commercial 

buildings, and the major source of electric energy use and demand comes from 

refrigeration systems (around 50 %), of which compressors and condensers account for 

60-70 %. Also, the majority of conventional supermarkets take partial advantage of the 

large amount of heat rejected, by using this heat for space or water heating. Generally, a 

conventional heat reclaim coil located in the central air handler allows recovery of 

approximately 35 % to 45 % of total heat rejection. However, this amount of heat is not 

sufficient to completely eliminate the use of fossil fuel combustion (natural gas, propane) 

during the coldest periods of the year. Moreover, at the present time, an issue facing the 

industry is the phase-out of ozone depleting refrigerants and the reduction of refrigerant 

charge and leakage, because a typical supermarket may experience refrigerant leakage of 

10 % to 30 % per year. 

The main objective of the LTE project was to experimentally document the benefits 

of two improved heat recovery systems developed in Canada, and to compare their 

performances with those of existing multiplex systems. Specific aims were to measure 

the energy usage for both low and medium-temperature zones and for total store, as the 

main system operating parameters and overall energy performances. An ultimate goal 

was the further development of options to reduce total equivalent warming impact 

(TEWI) of supermarkets by reducing the total refrigeration charge and leakage risks. LTE 

Laboratory worked with three retailers and two Canadian manufacturers on development 

and field testing of two commercial prototypes. The “System1-phase 1” (known as RSD) 

comprises a method to maximize the heat recovery capability for space and water 
21

heating, and a new, high-speed defrost concept. The “System2-phase 1” (known as LMP) 

proposes a new heat recovery and air-conditioning/dehumidifying method involving 

intermediate heat exchangers and multifunctional refrigerant-to-air heat pumps. These 

systems represented the first phase of their respective development processes. They have 

already been further improved, particularly in order to reduce total refrigerant charge and 

leak hazards. Three new supermarkets were extensively instrumented in order to compare 

a number of operating parameters, high-efficiency features such as heat recovery, 

defrosting, sub-cooling and liquid pumping, and to quantify energy performances 

(compressors and store energy consumption and demand, etc.). Two of the stores 

monitored, having approximately the same size and energy usage, (basically, electric 

refrigeration and natural gas cooking) were equipped with above-mentioned improved 

refrigeration systems, and a third one was used as a new baseline supermarket. In these 

stores, food is stored prior to transfer to the store in walk-in storage areas, and a variety 

of display cases are located at the periphery. A continuous monitoring of these systems 

was performed over a 12-month period (2000 – 2001). 

Total Heat Recovery (”System1-phase1”). This heat recovery method consists of 

bi-directional valves to supply discharge superheated vapour to condensers (in winter), or 

to a groundwater cooling device (in summer), depending on the outdoor temperature, or 

to heat reclaim coils for air or domestic hot water heating, when required. A parallel 

installed modulating valve equipped with an outside temperature sensor adjusts the 

discharge pressure depending on the outside temperature, and automatically re-directs a 

predetermined quantity of superheated vapour to the heat reclaim coils by means of a 

simple, economical and efficient method. If the system controller requires heating during 

the winter cold months, the modulating valve will direct some or all of the refrigerant to 

the heat recovery coils in order to cover up to 100 % of the space heating demand. The 

last heat reclaim coil and remote condensers are separately connected to a main receiver 

where the liquids flow at approximately the same pressure, independently of one another. 

Prior to being fed to the evaporators, the liquid is further sub-cooled by an internal heat 

exchanger equipped with special thermostatic expansion valves. 

Total Heat Recovery (”System2–phase1”). In conventional supermarket refrigeration 

systems, the condensing pressure and temperature are subject to ambient air temperature. 

During the cold periods of the year, even if there is a possibility of reducing condensing 

pressure, a high level is artificially maintained in order to provide sufficient pressure 

differential for proper operation of the expansion valves, and for heat reclaim. The 

”System2-phase1” heat recovery concept provides an original method to maximize the 

extraction of the condensing heat by using heat pumps. Refrigeration system superheated 

vapour is first directed to four intermediate plate heat exchangers that also serve as 

evaporators for heat recovery heat pumps where a minimum suction pressure, 

corresponding at about 10 °C refrigeration condensing temperature, is maintained. When 

store space heating is required, the plate heat exchangers and their respective refrigerantto-refrigerant heat pumps are activated. Sensible (superheat) and, if necessary, latent heat 

are removed from the hot discharge gas of the refrigeration system, which is then fed to 

the remote air condensers. The heat extraction from the superheated vapour is effected 

without excessively increasing the discharge pressure, thus allowing a reduction of the 

refrigeration compressor energy consumption as compared to conventional heat reclaim 

methods. In air conditioning mode, the plate heat exchangers aren’t used, and heat 
22

removed from the store is rejected outdoors through each heat pump’s own condenser. 

The heat recovery system is completely independent of the refrigeration system, and 

it is possible to extract total condensing heat without affecting its normal operating 

conditions. Each heat pump includes a heat exchanger for space heating, and the same 

coil is used for air re-heating in dehumidification mode. The heat pump’s condenser and 

air conditioning coil are not operational in heating mode, and special inlet pressure valves 

maintain a minimum suction pressure. Refrigeration system liquid returns from the 

remote condensers to a liquid delivery unit (including a refrigerant pump) where static 

pressure is increased before supplying the display cases, and a small quantity can be 

injected into the common discharge line during the summer. Low condensing pressures 

during the cold periods of the year are compensated for by the pressure increase provided 

by the liquid delivery system. 

Inside the heat pump dedicated to sub-cooling/heat reclaim, the evaporator removes 

heat from the condensed liquid of the main refrigeration system before it is sent to the 

display cases and storage coolers/freezers. When this heat pump operates in heat recovery 

mode, its own intermediate plate heat exchanger plays the role of an evaporator, and the 

discharge line goes to an air-heating coil for store space heating. It should be noted that 

this heat pump-heat reclaim concept is easily adaptable to any existing refrigeration 

system using a liquid delivery system. 

Results 

- With lower compressor energy consumption, the 12-month field monitoring 

proved that fossil fuel consumption for space heating in supermarkets has 

been completely eliminated in both cases. 

- “System1–phase1” raised the head pressures/condensing temperatures during 

the heating demand periods as high as they were in the baseline system with 

conventional heat reclaim devices, and also modified refrigerant flow and 

controls to eliminate current limitation of heat reclaim capability and giving 

priority to the heat reclaim mode. A modulating valve operating on the 

ambient temperature optimizes this process by varying head pressures as a 

function of the actual heating demand and the outdoor temperature variation. 

- The “System2-phase 1” was capable of reducing discharge pressures to an 

almost constant level (about 29 % lower than the other system), without 

compromising its own heat reclaim method. The discharge pressures were 

lowered to about 125 psig during the winter, but the thermodynamic level of 

the heat reclaim was increased by using heat pumps as a cascade stage. 

Savings in compressor energy consumption were however obtained, mainly 

by using four refrigerant-to-air heat pumps having average coefficients of 

performance of around 3.0. 

- The total annual specific energy consumption (kWh/y/m2

) of the refrigeration 

compressors represented 32.5 % (“System1–phase1”) and 34 % (“System2–

phase1”, including the four heat pumps), of to each store’s total electrical 

energy usage, compared to the current average in Canada, of 39 %. These 

performances were obtained even if the “System1 – phase1” was slightly 

penalized by the use of less efficient screw compressors, excessive fouling 

and lack of proper maintenance of groundwater heat exchangers. “System2–

phase1” was also penalized by the heat pump’s supplementary energy 
23

consumption, even when compensated for by lower head pressure operation in 

winter. Simultaneously, the total annual specific electrical energy 

consumption (vs. sales area) of improved supermarkets was reduced by 6 %, 

and the corresponding energy costs, including power demand cost, by 12.9 % 

and 17.9 % respectively, compared to the baseline monitored supermarket. 

High Speed Defrost. “System 1-phase1” system includes a high speed defrost 

concept capable of performing defrost cycles in a short period of time. In fact, the major 

disadvantages of conventional defrosting methods are their fairly lengthy cycles, because 

of a low-pressure differential across the evaporators, adverse effects on foodstuffs and 

compressors (energy cost increases, overheating, reduced technical life). As in 

conventional systems, when a refrigeration zone goes into a defrost cycle; superheated 

vapour is fed to the outlet of the evaporator coil. The return line of the defrost circuit is 

provided with a second valve which is open during the defrost cycle. This valve connects 

the return line to a special auxiliary receiver, which has a sufficient volume to take the 

full refrigerant charge and protect the compressors during and after each defrost cycle. 

The auxiliary receiver is connected to the low suction pressure header by means of a 

solenoid valve. By using an auxiliary receiver, a pressure differential can be created 

across the evaporator coils within a range of about 200 to 700 kPa, thus achieving quick 

defrost. In order to keep the main receiver supplied with sufficient quantities of liquid, it 

is necessary to periodically flush the auxiliary receiver during the refrigeration cycle. 

When the refrigerant temperature is below a predetermined value, normally –1 °C, the 

refrigerant liquid is directed to the main receiver. A pressure differential of about 200 kPa 

is created in order to have sufficient pressure to flush out most of the liquid accumulated 

in the auxiliary receiver. If the liquid is above –1 °C, it is sent to the remote condensers 

and then back into the main receiver. The auxiliary receiver is also provided with a level 

detector. When the liquid level reaches a predetermined value, the flushing circuit will 

operate during the next refrigeration cycle. A second level detector feeds an alarm that 

produces both a visual and an acoustic signal. 

Results 

- The new high-speed defrost method proposed by “System1-phase1” reduced 

defrost cycle time by about 85 % compared with a conventional hot gas 

defrost method. Immediate beneficial effects were a better quality and about 

50 % less perishable food losses. The actual maximum deformation of a 

representative case evaporator during a fast defrost cycle represents about 56 

% of the elastic limit of existing materials, but the long-term effects are not 

known, nor are they known for existing conventional systems. 

- Because of the high speed defrost, it was estimated that is possible to reduce 

the number of the refrigeration lines and consequently, the total refrigerant 

quantity by up to 15 %. 

Liquid Pumping and Injection. A simple, reliable liquid delivery technology 

developed in Canada can improve refrigeration efficiency, reduce compression ratios and 

work, and allow refrigeration systems to better operate with floating condensing 

pressures, depending on outdoor temperatures. This technology generally produces 

energy savings of up to 35 %, but creates difficulties in the area of heat recovery, 

especially when total condensing heat reclaim is required, because the air to be heated is 

warmer than the refrigerant to be cooled. Accordingly, when heat reclaim is required, the 
24

condensing pressure has to be artificially raised in order to achieve proper heat transfer, 

thus losing the benefits of the liquid delivery system. However, by combining liquid 

pumping with the new ”System2 – phase1” heat recovery method, it was expected that 

head pressures would not be artificially raised when heating is required. Refrigerant 

pumping pressurizes the liquid refrigerant enough to compensate for any pressure losses 

between the receiver and the expansion valve, and thus eliminates the formation of 

vapour. The liquid pump is a sealed unit, with no moving seals to wear out or leak. The 

”System2 – phase 1” new heat recovery concept allows the use of the same equipment 

(heat pumps) for air conditioning and sub-cooling purposes, and also improves the 

efficiency of the liquid pumping technology. It is well known that the compressor’s 

energy consumption and efficiency are a function of the ratio between the condensing 

and evaporating pressures. By lowering the condensing temperature to 16 °C or even 10 

°C, the compressor energy consumption and power diminish, and its refrigeration 

capacity increases. Moreover, because the pressure of the pump is usually higher than the 

compressor discharge, the injection of a small percentage of the liquid (6 – 10 %) back 

into the discharge line was possible. Thus, for ambient temperatures above 24 °C, the 

temperature of the discharge vapour was reduced, and the efficiency of the condenser 

increased. The liquid injection means lower velocity through the condenser, lower 

pressure drop, reduction of the discharge temperature, an increase in the area of the 

condenser used for condensing, and increased system capacity. 

Result 

- Liquid pumping technology used by “System2 - phase1” helped this new heat 

recovery technique by increasing the liquid pressure by about 180 kPa during 

the cold periods when the head pressure decreased to 860 kPa. 

Geothermal Cooling. During the hottest periods of the year, condensing heat 

from the ”System1–phase1” refrigeration system was rejected to the ground instead to the 

ambient air in order to reduce energy consumption and thus optimize the overall 

efficiency of the system. This concept allowed the reduction of the compressor’s head 

pressures and condensing temperatures, because the cooling fluid will be groundwater at 

an average temperature of about 8°C (46.4°F) instead the ambient air at more than 30°C 

(86°F). This objective was achieved by using supply and discharge groundwater wells. 

Two refrigerant-to-water heat exchangers for both low and medium temperature racks 

were employed. The main disadvantage of this energy efficiency method is the risk of 

clogging of the groundwater heat exchangers, which involves additional costs for 

periodical maintenance. However, the energy savings may generally justify the initial 

investment and supplementary maintenance. 

Further Developments. The major weakness of the proposed systems lies in the 

large of refrigerant required, and the associated environmental disadvantages (1,400 to 2,300 

kg for total sales area of 3,000 to 5,600 m2

). Even if the “System1 – phase1” has 

demonstrated that the developed high speed defrost method enables a reduction in 

refrigerant charge of up to 15 %, it isn’t considered sufficient. Therefore, the main objective 

of future systems improvement is to reduce total refrigerant quantity and leakage risks, and 

also to replace the HCFC-22 by alternative fluids, in order to develop new environmentally 

low-TEWI systems. 

(i) The “System1 – phase2” improved system, available since 2001, was developed 

by retaining both prior innovations (integral heat recovery and high speed defrost). It aims to 
25

replace the actual HCFC-22 fluid with a good replacement having no, or lower 

environmental impact (e.g., R-404a). The central refrigeration system was divided into 

several smaller, ultra-compact units that can be installed on the roof (distributed compressor 

approach). The previous conventional central refrigeration room was thus completely 

eliminated, and the space created could be used for commercial activities. This arrangement 

reduces total refrigeration piping length and number of fittings, thermal insulation and 

electric piping by 60%, and total refrigerant charge by 30 – 40%, because of the proximity 

of display cases and of heat reclaim coils, reduced diameters and length. Thus, the additional 

costs are lowered by 40 % compared to the previous system. These cabinets are insulated 

and equipped with their own heating and cooling devices (21°C in winter and 32°C in 

summer). Additional ventilation or make-up air and evacuation devices aren’t needed. 

“System1 – phase3” improved system aims also to retain the main previously analysed 

innovations, and also to reduce the refrigerant charge by 60 % or more compared with the 

first method. This concept retains the same small cabinets, but each high temperature side 

(condensers and heat reclaim) uses closed loops with circulation pumps and anti-freeze 

secondary fluid, storage tanks and controls. These condensing circuits can transfer heat 

directly to each heated space, or can use water-to-air heat pumps. Reduction of additional 

costs is estimated at 50 % compared to the first improved system. 

(ii)“System2 – phase2” aims to maintain heat recovery by heat pumps, but to 

improve the liquid delivery system (horizontal receiver installed outdoors with improved 

level control), accelerate the defrost cycles by increasing the refrigerant flow rate, and 

reduce the additional cost by about 10 % compared to the previous concept. On the other 

hand, “System2 -phase3” aims to develop a simplified heat recovery method without heat 

pumps, a zoned defrost concept, and to reduce the refrigerant charge by up to 30 % and 

the additional cost by 80 % compared to the first improved system. 

Summary information on the phase 1 test systems is included in Appendix A. 

Full details on the tested and improved systems may be found in the Canada country 

report. 

Denmark

There is a great deal of concern in Denmark about the use of HFCs due to 

concerns about their impact on the environment in the long term, especially the 

contribution to global warming. In the opening speech to the IIR conference on 

“Applications for Natural Refrigerants”, Aarhus, Denmark (1996) the Danish Minister for 

Environment and Energy, Svend Auken, in his conclusion said: “It is therefore my 

sincere hope that in ten years’ time, not a single fridge, freezer or cooling plant is being 

built in Denmark that requires HFCs or other greenhouse gases.” Recently an 

announcement (in Danish; Bekendtgørelse, BEK nr. 552 of July 2., 2002) has been issued 

which forbid the use of certain industrial warming potential gasses, which includes the 

HFC ́s. From January 1, 2006 the use of HFC’s is no longer allowed in new products. 

Exemptions are the use in refrigeration plants, heat pumps, air conditioning plants and 

dehumidifiers with refrigerant charge between 0.15 kg and 10 kg and compact heat 

recovery plant factory assembled with charge less than 50 kg. According to this rule 

conventional multiplex refrigeration systems cannot be used in supermarkets after 

January 2006. 
26

Besides the phase out in 2006 for new plants HFCs are charged a CO2 tax. In 

Denmark emission of carbon dioxide (CO2) has been taxed since May 15, 1992. Starting 

January 1, 2001 following the Kyoto protocol Denmark extended the CO2 tax to include 

industrial greenhouse gasses (HFCs, PFCs and SF6) based on the global warming 

potential (GWP) of each gas. The tax is US$ 0.0127 (DKK 0.10) pr. kg CO2 with CO2 by 

definition having the GWP of 1 with a maximum tax of US$ 50.95 (DKK 400) pr. kg 

(Danish law #49 of December 21, 2000). 

In view of what is stated above it is obvious why the Danes are searching for 

natural refrigerant substitutes for HFCs. Several supermarket case studies have been 

done in Denmark with a primary focus on increasing energy efficiency and applying 

natural refrigerants (like hydrocarbons, ammonia, and CO2). In one small Danish 

supermarket the old refrigeration plant was replaced with a cascade plant. Propane was 

used as the high temperature refrigerant (-14/+30°C) while CO2 was used at the low 

temperature level (-32/-11°C). Direct expansion of CO2 is used in the freezer cases while 

a brine circuit with propylene glycol is used in the coolers. Total energy consumption 

decreased by 10% with the new plant, however it must be noted that the old plant was 

worn out. Based on the results from this initial demonstration a second propane/ CO2

cascade plant was built in a medium-size supermarket and just started operation in late 

2001. The test supermarket is part of a chain of about 250 similar shops so it has been 

possible to compare the propane/carbon dioxide plant with newer traditional plants with 

the same cooling capacity. Energy use of the demonstration store has been compared to 

that of eight similar stores in the same chain and found to be about the same as that of the 

most efficient of the newest conventional stores. It is estimated that the additional cost for 

a propane/ CO2 cascade plant for a medium sized Danish supermarket (30 kW freezing 

load and 60 kW cooling load) will amount to approximately 10% or less of the total 

installation. Details of this second demonstration are included in Appendix A. 

A third demonstration plant has recently been started in a new supermarket in 

which carbon dioxide is used in both the low temperature and medium temperature cases. 

This approach eliminates the secondary propylene glycol circuit with its thermodynamic 

efficiency penalty. The plant has 190 kW cooling capacity and 60 kW freezing capacity, 

and the condenser heat is utilized to supply the shop with all of its space heating and 

sanitary hot water needs. The refrigeration plant is a cascade plant with R404A in the 

high temperature stage and CO2 in the low temperature stage. R404A is used in this plant 

because the equipment is installed in the basement of the store where it is difficult to get 

permission to install flammable or toxic refrigerants. Both the freezer cases and fresh 

food cases are cooled by direct evaporation of CO2. Liquid CO2 is pumped directly to the 

fresh food cases while CO2 is direct expanded into the freezer cases. The amount of 

R404A in this demonstration is only about 10% of that needed by conventional direct 

expansion systems for stores of this size. Taking into account the Danish tax on R404A 

this reduction reduces the investment in refrigerant by about US$ 37,000. (DKK 

290,000). The total investment in the refrigeration plant is approximately 10% higher 

than a conventional HFC-plant taking into account the reduced refrigerant cost. This 

demonstration has been underway since February 2002. A comparison with similar shops 

in the area with conventional refrigeration systems shows significantly lower energy 

consumption. In November 2002 a second plant built in according with this concept has 

been set in operation. 
27

In the future a project with title, ”Energy efficient design and control of the 

secondary side of an indirect refrigeration system with natural refrigerant” will be 

performed. 

Sweden

 Swedish activities for Annex 26 have been aimed at improving supermarket 

energy efficiency in several different ways. These have included development of a 

detailed computer model and analyses of full stores to examine different secondary loop 

system designs and heat recovery approaches including impact of climate effects 

particularly relative humidity. Several full-scale demonstration projects have been 

undertaken examining different secondary loop systems and display cases with advanced 

features such as automatic night covers. 

Increased sales in supermarkets, stricter environmental legislation for CFC and 

HCFC refrigerants and a major consideration of the use of energy and the effect on the 

environment have influenced the supermarket sector during the last year in Sweden. Sales 

in supermarkets have increased 4.6 % during year 2001 in comparison with the year 2000 

[30]. The consumption of deep-frozen products in supermarkets during 2001 was 231788 

metric tons while during 2000 it was 216427 metric tons, which is an increase of 7%. The 

increase in sales in supermarkets places a demand on more cabinets, cold rooms and 

refrigeration systems that increase the energy consumption from food stores. 

Supermarkets are using large amounts of energy; approximately 3% of the electric 

energy consumed in Sweden is used in supermarkets (1,8 TWh/year). A breakdown of 

the energy usage shows that, typically, 47% is used for medium and low temp 

refrigeration, 27% for illumination, 13% for fans and climate control, 3% for kitchen, 5% 

for outdoor usage and 5% for other uses. 

Refrigeration systems, display cases, indoor climate control and illumination are 

the areas with the greatest potential for improvement. Since the energy systems of a 

supermarket are relatively complex, improvements in one subsystem affects other 

systems, thus making an analysis of potential improvements non-additive. Typical 

efficiency improvements may involve refrigeration systems, heat recovery, more efficient 

illumination, more efficient display cases with night lids, more efficient control, floating 

condensation etc. 

Several new system solutions such as completely, partially and cascade secondary 

loop (or indirect) systems have been developed and introduced in recent years in Sweden 

to lower the refrigerant charge and, at the same time, minimize potential refrigerant 

leakage. 

Growing interest in secondary loop systems has led to the development of some 

new secondary refrigerants based on potassium formate and potassium acetate alone or 

mixed. Another very promising development is CO2 as a secondary refrigerant. This 

technology is implemented in more than 30 supermarkets throughout Sweden. CO2

systems require much lower tube diameters and the pressure drop is negligible when 

compared to conventional systems. A third promising development in this field is ice 

slurries. These systems offer additional advantages with enhanced thermal capacity and a 

“built-in” thermal storage in the system without increased pressure drop if ice-slurry with 

the right consistence is produced. 
28

Secondary systems and the minimization of refrigerant charge may lead to an 

unwanted trade off in overall energy efficiency. Theoretical calculations confirm this due 

to the obvious extra temperature differences introduced in the system. An evaluation 

using a concept like TEWI (Total Equivalent Warming Impact) may be used to estimate 

the overall environmental impact. 

In 1998 the Department of Energy Technology of the Royal Institute of 

Technology of Sweden started a project in co-operation with different companies and the 

Swedish National Energy Administration. Modelling and field measurements of 

supermarket energy systems have been undertaken in the research project “The energy 

efficient supermarket” for more than four years. The results have been part of the 

Swedish contribution to the IEA Annex 26. 

The overall aim of the project “The energy efficient supermarket” is to develop a 

sound simulation model where different system solutions can be compared in detail with 

focus on energy usage, environmental impact (TEWI) and LCC (Life Cycle Cost). 

The program, “CyberMart” is built in modules dealing with subsystems such as 

in- and outdoor climate, display cases, cooling and freezing rooms, refrigeration 

machinery, the building envelope etc. CyberMart is a day-to-day simulation program that 

allows the user to see the variation of different variables such as compressor power, 

refrigeration capacity and temperatures in the supermarket during one year. 

The model is currently under validation with four different supermarkets in 

Sweden in cooperation with COOP and ICA, two major Swedish supermarket chains. 

The model development activities are now treating issues such as illumination, indoor 

climate and heat recovery, comfort cooling etc. Interesting issues to study are the 

potentials in heat recovery in the winter, moisture control/dehumidification and floating 

condensation. Additional Field measurements intended to validate data are currently 

being set up. A window for calculation of energy consumption is shown in Figure11. 

The diagram shows the results from the simulation of the compressor power of the 

intermediate and low temperature systems during one year. 
29

Figure 11 - Results from CyberMart

The influence of the outdoor temperature on the indoor relative humidity of air is 

an important factor to take in consideration when dimensioning refrigeration and heat 

recovery systems in supermarkets. Lower outdoor temperature and moisture affect the 

compressor power and the dry cooler fluid approach temperature. 
30

Figure 12 - Indoor temperature, outdoor temperature and indoor relative humidity 

in Hedemora and Sala 

Measurement of indoor temperatures, outdoor temperatures and indoor relative 

humidity from two supermarkets in the cities of Hedemora and Sala are shown in Figure 

12. The distance between the cities is about 60 km. The results confirm the dependence 

of relative humidity on outdoor temperature. The indoor relative humidity in both 

supermarkets follows the variation of outdoor temperature during the period of study. 

Sala - Hedemora 

Aug 2000 - Feb 2001

-30

-20

-10

0

10

20

30

40

50

60

70

aug-00 aug-00 sep-00 okt-00 okt-00 nov-00 nov-00 dec-00 jan-01 jan-01 feb-01

Time

temp[°C], RH[%]

Indoor Temp. Hedemora

Outdoor temp. Hedemora

RH. Hedemora

Indoor Temp. Sala

RH. Sala

Outdoor Temp. Sala
31

Figure 13 - Temperatures, relative humidity and compressor power in Sala 

The variations of temperatures, moisture and compressor power during week 32 

in August 1999 in the supermarket in the city of Sala are shown in Figure 13. In the 

figure it is possible to see the influence of night covering of the display cases and 

defrosting on the compressor power. The cabinets are defrosted with electrical defrost. 

The compressor power is reduced between 10 and 20 % by the night covering of the 

cabinets. The covering occurs automatically when the supermarket closes at 21.00 and 

ends at 8.00 in the morning. The positive effect on energy saving of night covering is 

dependent on the quality of the curtains. Night covering is an efficient method to reduce 

infiltration and radiation loss in cabinets. 

The influence of high indoor temperatures on compressor power has been studied 

in the supermarket in the city of Hedemora. The high investment cost of the air 

conditioning system and a short period of higher outdoor temperatures during the summer 

have affected the decision of install AC in many supermarkets in Sweden. 

Sala 

Week32

0

10

20

30

40

50

60

70

1999-08-09 1999-08-10 1999-08-11 1999-08-12 1999-08-13 1999-08-14 1999-08-15 1999-08-16

Time

temp[°C], power[kW], RH[%]

Indoor temp.

Outdoor temp.

Comp. Power

RH
32

 Figure 14 - Temperatures in a deep freeze cabinet and indoor and outdoor 

temperatures during four days in July in Hedemora 

Figure 14 presents indoor temperatures, outdoor temperatures and air return 

temperatures in four sections of a deep-freeze cabinet during four days in July 2001. The 

outdoor temperatures are above 30°C during a long part of the day and this affects the 

indoor temperatures. The values of indoor temperatures during the day are above 25°C, 

which is the dimensioning indoor temperature of the display cases. The maximum indoor 

temperature is 28°C, which occurs on the 5th of July at around 17:00 o’clock. The air 

return temperatures in three sections are above –20°C and one of them has air return 

temperatures higher than -15°C during the day. If the temperature of the products is 

assumed 1°C higher than the air return temperature, which is a reasonable supposition, 

then the temperature of the product should be about –14°C at an air return temperature of 

-15°C. According to the National Food Administration, which regulates and supervises 

the food area in Sweden, the maximum temperature of freeze products is –18°C. The date 

for the minimum durability, or use-by date, of frozen food is calculated at a temperature 

of –18°C. Higher product temperature implicates a risk of the multiplication of bacteria in 

the frozen food that might cause illness to the consumers. The results from Figure 12 

confirm the necessity of air conditioning during the warmer days in the summer in 

Sweden. 

The economy and overall energy efficiency of supermarkets, in cold climates, 

benefit from heat recovery. In theory, the necessary heat can always be supplied from the 

condensers. Practical experiences show that in real systems only 40 % – 70 % of the 

necessary heat is recovered. There are many reasons for this but the most important are 

on/off regulation, low cooling load during cold days, poorly designed heat recovery 

systems and non-communication control systems for refrigeration and HVAC. 

Hedemora 

3 - 6 July 2001

-30

-20

-10

0

10

20

30

40

50

00:00 12:00 00:00 12:00 00:00 12:00 00:00 12:00 00:00

Time

Temp [°C], RH[%]

Indoor Temp.

Outdoor Temp.

DFsec1

DFsec2

DFsec3

DFsec4

RH
33

United Kingdom

Storing and displaying refrigerated goods accounts for a considerable amount of 

energy use in the UK. Figure 15 shows the breakdown of the end use of refrigeration 

energy. Of the total of 24,000 GWh, refrigerated retail display cabinets overall consume 

some 7,740 GWh per year in the UK. This provides a great incentive to improve the 

energy efficiency of refrigerated supermarket display units. In addition to this there is 

also the requirement to reduce the emission of environmentally damaging refrigerants 

that contribute to the total equivalent warming impact. 

Figure 15 - Refrigeration energy usage in The United Kingdom by end use 

Reported here are a summary of the Phase 2 studies and the overall analysis of the 

current position regarding the available technologies. Full details of the Phase I 

technology review and Phase 2 detailed studies are included in the UK country report. 

The second phase of the UK contribution covered four research areas: 

• secondary refrigerant systems 

• alternative defrost approaches 

• interactions between the store environment and the cabinet case 

• combined heat and power and combined heating, cooling and 

power applications in supermarkets. 

This summary records the main points that were concluded from each of the four 

studies carried out. Each area of research was carried out by an organisation specialising 

in refrigeration technology. The areas where further research may be useful have also 

been identified. 

Secondary Systems. Secondary refrigerant systems offer the potential of reduced 

total equivalent warming impact (TEWI) as a consequence of reduced refrigerant leakage 

of high ozone depleting potential (ODP) refrigerants. However, it is recognised that there 
34

may be additional energy requirements to operate this type of system because of greater 

temperature gradients and increased pumping power. 

Field monitoring data from a typical supermarket was used to validate a 

refrigeration simulation model, and a detailed investigation of the thermal and fluid 

properties of secondary refrigerant fluids was made. This enabled simulations to be 

carried out to assess the feasibility of a wide range of secondary systems and refrigerants 

in a typical supermarket. A key finding of the simulation was that the energy use of the 

secondary systems simulated would typically be greater than for other systems and as a 

consequence they would deliver higher operational CO2 emissions. In terms of overall 

TEWI, the secondary system using R290 (propane) would reduce the TEWI by 8% over a 

15-year life compared to an R22 direct system with 15% refrigerant leakage. The benefit 

of a reduced TEWI of a secondary system is lessened as the leakage rate of the equivalent 

direct system decreases. 

The project has shown that there is sufficient knowledge of secondary fluid 

properties to be able to design an effective system. However, careful optimisation will be 

needed for a secondary system to become more attractive both financially and in terms of 

TEWI, so further research is recommended into the optimisation of secondary systems. 

Further research needs. Secondary systems using pumpable ice have been tried in 

the UK but further work is needed to establish if it is a viable option. Research is 

also needed to address the issues of store layout and pipework arrangements with 

the specific aim of reducing the capital cost of the secondary system. 

The use of CO2 as the secondary refrigerant needs to be researched in the 

UK context. The current situation is not clear with respect to the health and safety 

issues and the performance of the systems in operation. Cascade systems have 

been used in the USA and may have some energy benefits and the possibility of 

using these in the UK should be investigated. 

If changes in legislation force the adoption of lower Global Warming 

Potential (GWP) refrigerants then retailers may wish to investigate secondary 

systems further. The options for improvement include: system design (store 

layout to reduce pipework capital costs); pumpable ice; CO2; and cascade 

systems. Reference should be made to experience in North America and 

Scandinavia where some of these approaches have been adopted. 

Defrost. The energy used for defrosting evaporator coils in supermarkets can 

amount to 30% of the operating energy. Alternative means of reducing the defrost energy 

have been researched by laboratory studies of different defrosting techniques and control 

strategies. Studies were carried out in both high and low temperature applications and the 

following defrost techniques were considered: 

• off-cycle 

• hot/cool gas 

• electric. 

However, not all techniques can be used in all circumstances, for example, offcycle is not appropriate in low temperature applications, leaving a choice of gas or 

electric defrost. 

The findings showed that, for high temperature applications, the cheapest and 

least energy consuming means of defrost is ‘off-cycle’. This method of defrost is widely 
35

used in the industry. For low temperature applications, the use of gas defrosting is more 

energy efficient than electric defrost but the considerable extra capital costs lead to long 

payback periods, up to 28 years. Consequently, electric defrost is most commonly used 

for low temperature applications in the UK. 

The length and frequency of the defrost cycle has a significant effect on the 

energy use of a refrigerated cabinet. The normal procedure of a fixed number of cycles 

per day is wasteful of energy because the time period between defrosts may be either too 

long - thereby reducing the frosted evaporator efficiency before defrost - or too short 

which provides unnecessary defrost events. In addition, the length of defrost can be 

excessive and this wastes further energy, and risks adverse effects on the product quality 

and life. Optimised defrost cycles could reduce energy use by between 25 and 50%, 

where electric defrost is employed. Work is under way to develop controls to achieve 

this. 

The optimum defrosting cycle will depend on the cabinet type and store 

conditions and a means of control that relates to these parameters would lead to 

significant energy savings. 

Further research needs. A number of issues for further research were revealed by 

these studies. Liquid defrosting for low temperature applications should be further 

researched as it is potentially a more effective and energy efficient technique. 

 Methods of determining the need for defrosting should be investigated and 

these should take into account the avoidance of simultaneous defrosting of a 

number of units leading to high electrical loads. Evaporator design has an 

influence on the formation of ice and should also be researched in order to avoid 

or reduce the incidence of frosting. 

CHP/CCHP. Combined heat and power and combined cooling, heat and power 

are potentially efficient methods of providing the energy for supermarkets. A simulation 

model has been developed that allows the energy and heat flows in a supermarket to be 

modelled and will predict energy and cost savings arising from a range of measures taken 

to improve efficiency of supply or use of energy. The model was tested against a store 

using CHP and then used to evaluate a range of options. 

The studies showed that currently available equipment can be used satisfactorily 

but needs to be carefully optimised for the specific store in which it is being installed. 

The cost effectiveness of CHP/CCHP is sensitive to the fuel costs and in particular the 

ratio of the gas and electricity costs. Additionally the number of hours for which the 

plant operates in a year is a key determinant of the cost effectiveness. 

Further research needs. Determining the heat and power load profiles of a wider 

range of stores would assist in developing a clearer view of the potential for CHP 

in supermarkets. 

The design of the refrigeration system with the use of CHP and CCHP 

needs to be investigated as novel approaches may provide a better match of heat 

and power usage. For example, gas engine prime movers for vapour compression 

machines can be attractive and the potential for their use should be examined. 

The simulation model that has been developed in the context of this 

research could be made available for retailers and other interested parties to carry 

out feasibility studies on future stores. 
36

Case and Store Environment Interaction. This project was based on the premise 

that the internal environment of the store, in which the refrigerated display cabinet is 

located, can have a significant bearing on the energy use and performance. The 

temperature and humidity of the air and its movement around the cabinet are all likely to 

influence the energy use. A store environment was measured and the conditions were 

recreated in the laboratory. Tests were then carried out with these different internal 

environmental parameters to investigate their impact on energy use. Two different cases 

were tested in the laboratory – an integral (self-contained open front chiller unit and an 

open well (coffin-type) freezer with a row of half-height glass door shelves above the 

open well (see Figure 16). 

Figure 16 - Example of freezer case type tested by UK team in environmental 

chamber to determine impact of store temperature, relative humidity and 

ventilation mode (conventional or displacement) on performance 

The results showed that for the freezer tested varying the store’s temperature, 

relative humidity and ventilation mode and rate had little impact on the energy 

consumption of the unit. It may be surmised that although the rate of frost formation on 

the evaporator was dependent on the ambient relative humidity, whilst the defrost 

initiation and duration were preset, no significant impact on the energy use for defrost 

was observed. 
37

For the chilled cabinet the energy consumption was more dependent on the store 

conditions. A change of store temperature from 19 to 22o

C increased the energy use by 

up to 20%. At the higher temperature of 22oC the relative humidity of the store 

environment increased the energy consumption by 15%, when raised from 35%rh to 50% 

rh. Increased ventilation around the chiller could also lead to some increase in energy use. 

Overall, relative humidity is not as much of a significant factor in energy use as 

was initially expected from the literature review. This is probably because the rh values 

in the UK are typically much lower than in the USA, where the other research had been 

carried out. 

An additional finding was that reduction of the voltage supply to the cabinet had a 

large effect on the electricity use. Reducing the voltage from 240 to 220 volts resulted in 

a saving in energy of approximately 10% without any loss in performance. This effect 

has also been observed in stores, and current practice is to reduce voltage supply to the 

whole store in order to make savings. 

Further research needs. The use of high velocity air supply in aisles in order to 

mix the stagnant cool air needs to be investigated particularly to ascertain the 

impact on customer comfort. This needs to be studied in the context of customers’ 

expectations of comfort levels in stores and particularly adjacent to refrigerated 

display areas. 

Research is required into controlling anti-sweat heaters by methods other 

than time control. The control of the anti-sweat heaters is an area in which there is 

a potentially large energy saving to be made. 

The design and manufacture of cabinets is a major issue that warrants 

further examination. The tendency is to provide standard solutions that are not 

being developed in line with the aims of improving energy efficiency. To 

overcome this needs the active participation of manufacturers in the realistic 

testing and measurement of in-store operation in future trials. 

UK Overall Conclusions. These projects have confirmed that many of the 

practices the retailers currently have in place are the most cost effective e.g. the use of 

DX systems at low leakage rates, and electric and off-cycle defrosting. However, some of 

the more expensive and/or innovative technologies could further reduce the TEWI of 

supermarket refrigeration. 

A number of issues remain to be addressed in order to achieve these potential 

savings in all applications. There also appears to be great scope for considering the whole 

system – from power supply to refrigerated goods display in a store environment – as a 

single system to be optimised; rather than a collection of individual components each 

acting in isolation. Of particular note are: 

• Defrost methods and control strategies – which has great retrofit potential 

• Secondary system optimisation 

• Store environmental conditions for comfort and cabinet energy use 

• Variable speed compressors. 

United States

 The US contribution to the Annex included an analysis of several low-charge 

refrigeration system options, a field test of a distributed compressor refrigeration system, 
38

and development of a spreadsheet-based refrigeration system model. Summaries of the 

analyses and the field test results are given below. The spreadsheet model is described 

briefly below and more fully in the US country report. Full details of the analyses and 

field test projects are given in the US country report. Appendix A provides some 

summary information on the field tests as well. 

 There are over 100,000 retail food markets in the United States. Of these almost 

33,000 are supermarkets – stores with over $2 million in sales annually and generally 

having much more than 1000 m2 of sales floor area. Total annual electricity consumption 

for refrigeration alone in all US supermarkets combined is about 35-40 billion kWh.

System analysis summary. A 3720 m2 supermarket based on the layout in Figure 

1 was simulated and energy consumption estimates were made for a baseline air-cooled 

multiplex refrigeration system which uses R22 for the medium temperature compressor 

rack and R404A for the low temperature compressor rack and four advanced, low charge 

systems (Walker and Baxter, 2003). The advanced systems investigated included 

distributed compressor (R404A), secondary loop (R507, primary refrigerant), low-charge 

multiplex (same refrigerants as baseline), and advanced self-contained (R404A). The 

secondary loop system modeled included four separate temperature loops operating at 

temperatures of -30, -18, -7, and -1 °C with a potassium formate brine in the lowest 

temperature loop and propylene glycol brine in the higher temperature loops. The lowcharge multiplex system modeled was based on the system described in Figure 7. 

 Total refrigeration load was 328 kW (82 kW for frozen foods and 246 kW for 

fresh foods). The baseline system charge was 1360 kg or 4.15 kg/kW load. Both the 

distributed compressor system (with secondary loop for heat rejection and evaporative 

cooling tower) and the secondary loop system (with evaporative condensing) achieved 

similar results with annual energy savings of 11.3% and 10.4%, respectively, compared 

to the baseline for a Washington, DC location. The low-charge multiplex refrigeration 

system (with an evaporative condenser) showed annual energy savings of 11.6% 

compared to the baseline system. No energy savings were estimated for the advanced, 

self-contained system. Refrigerant charge reductions were 67% for the distributed 

compressor system, 85% for the secondary loop system, and 33% for the low-charge 

multiplex system. Energy savings results are summarized in Table 1. 
39

Table 1 - Predicted Energy Consumption for Low-Charge Refrigeration Systems 

System Heat Rejection Annual 

Energy 

(kWh) 

Energy Savings 

(kWh) 

% Savings vs 

baseline 

Multiplex 

(baseline) 

Air-Cooled 

Condenser 

976,800 - - 

Multiplex Evap. Condenser 896,400 80,400 8.2 

Low-Charge 

Multiplex 

Evap. Condenser 863,600 113,100 11.6 

Distributed Water-Cooled 

Condenser, Evap 

Rejection 

866,100 110,700 11.3 

Secondary Loop Evap. Condenser 875,200 101,600 10.4 

Advanced SelfContained Water-Cooled 

Condenser, Evap 

Rejection 

1,048,300 - - 

Secondary Loop Water-Cooled 

Condenser, Evap 

Rejection 

959,700 17,100 1.8 

Results for supermarket at Washington, DC location 

Use of evaporative heat rejection (evaporative condenser or cooling tower) was 

the principal driver for energy savings for all of the advanced systems compared to the 

baseline case. The energy consumption of the multiplex baseline system would be about 

8.2% lower if equipped with an evaporative condenser instead of the air-cooled 

condenser. 

An environmental assessment of these refrigeration systems was also made 

through a TEWI analysis for a 15-year life. The lowest TEWIs were achieved by the 

distributed compressor system and the secondary loop systems with estimated CO2

emission reductions compared to the baseline multiplex system (multiplex with aircooled condensing) of 13- 14 million kg, or about 57 - 60%. The low-charge multiplex 

system had estimated TEWI reductions of about 43% compared to the baseline system 

for a 15% annual refrigerant loss rate. Replacing the air-cooled condenser with an 

evaporative condenser on the baseline system resulted in an estimated TEWI reduction of 

about 3% due to the energy savings from operation at lower condensing temperatures. 

TEWI results are summarized in Table 2. 
40

a

Conversion factor = 0.65 kg CO2/kWh 

b

1/3 R404A (low temp), GWP = 3260; 2/3 R22 (medium temp), GWP = 1700 c

R507, GWP = 3300 

One observation from this study is that further efforts to reduce total global 

warming impacts for the distributed and secondary loop systems would benefit more 

from reduction in energy usage (through efficiency increases or load reductions) than 

from further reduction in direct impact (from refrigerant losses). This based on the fact 

that the direct (refrigerant loss) portion of the estimated TEWI for these systems only 

comprises about 10% of the total. 

Payback estimates. A simple payback analysis was conducted based on the 

results from the analysis project. Estimated installed cost premiums for the distributed 

compressor, secondary loop, and low-charge multiplex refrigeration systems were 

$60,000 (about 15%), $147,000 (about 35%), and $0, respectively compared to the 

baseline air-cooled multiplex DX system. These estimates were based on actual 

construction budgets supplied by the engineering departments of two supermarket chains. 

It should be noted that the actual installed cost of any refrigeration system will vary 

greatly, depending upon many factors, such as: purchasing arrangements between the 

supermarket and refrigeration equipment vendors; whether or not display cases are 

purchased in conjunction with the refrigeration system; special system features or 

configuration requested by the supermarket; and unique installation requirements of each 

site. 

The cost premium for the water-source heat pumps used for heat recovery in 

refrigeration systems that use a secondary fluid heat rejection loop was estimated to be 

$25,000 (about 30%) over a conventional system consisting of roof top airconditioners/gas heaters. The cost premium includes water piping for the heat pumps and 

Table 2 - Total Equivalent Warming Impact (TEWI) for Supermarket Refrigeration 

TEWI (million kg of 

CO2) 

System Condensing Charge, 

kg/kW 

Primary 

Refrigerant 

Leak 

(%) 

Annual 

kWh 

Direct Indirecta Total 

Multiplex Air-Cooled 

(baseline) 

4.15 30 

(15)

976,800 13.62 

(6.81) 

9.52 

(9.52) 

23.14 

(16.33)

 Evaporative 4.15 

R404A 

/R-22b 

30 896,400 13.62 8.74 22.36 

Low-Charge 30 863,600 9.08 8.42 17.50 

Multiplex 

Evaporative 2.77 R404A 

/R-22 b 15 863,600 4.54 8.42 12.96 

Distributed 

Compressors 

Water-Cooled, 

Evap. Tower 

1.24 R404A 5 866,100 1.00 8.44 9.44 

Secondary 10 875,200 1.13 8.54 9.67 

Loop 

Evaporative 0.69 R507 c

5 875,200 0.56 8.54 9.10 

 Water-Cooled, 5 959,700 0.23 9.36 9.59 

Evap. Tower 

0.27 R507 c

2 959,700 0.09 9.36 9.45 

 

Results for Washington, DC location – 15 year service life 
41

over-sizing of the refrigeration heat rejection system to allow heat pump heat rejection 

during space cooling. 

Simple paybacks for refrigeration operation only were estimated for each of the 

low-charge systems at four different locations in the US. Operating cost savings for the 

refrigeration systems include the savings obtained for reduced energy use and refrigerant 

leakage. Water costs are included for all systems employing evaporative heat rejection. 

The payback period for the low-charge multiplex system is immediate, since no installed 

cost premium exists. Paybacks for the distributed compressor system ranged from 3.4 to 

7.0 years, while the payback for the secondary loop system ranged from 8.3 to 16.8 years. 

In addition simple payback was estimated for a distributed refrigeration system in 

combination with water-source heat pumps for store HVAC compared to a baseline 

multiplex refrigeration system with air-cooled condensers and rooftop air-condensers 

with gas heat for HVAC. These pay back periods ranged from 4.2 years to 7.8 years 

depending upon local outdoor climate conditions. The combined payback for 

refrigeration and HVAC savings was less than that for refrigeration savings alone for 

cold climate locations because of increased space heating savings. Combined paybacks 

were lower for refrigeration system alone in the warmer climate locations because space 

heating is not as significant at these locations. 

Field test summary. A field test was conducted to compare the performance of a 

conventional multiplex system and a distributed compressor refrigeration system. Two 

supermarkets in the vicinity of Worcester, Massachusetts were the sites utilized for the 

field test. One store was equipped with a multiplex refrigeration system that had three 

compressor racks and air-cooled condensers. At the second store, a low-refrigerantcharge distributed refrigeration system was installed that consisted of 10 compressor 

cabinets. Heat rejection for the compressor cabinets was accomplished through watercooled condensers piped to a fluid loop that used dry fluid coolers. The second store also 

had water-source heat pumps for space heating and cooling that were piped into the fluid 

loops. Both sites were instrumented to determine energy consumption, refrigeration and 

heating, ventilating and air conditioning (HVAC) loads, and numerous system state 

points to characterize operation. 

Energy consumption of the two refrigeration systems was compared for winter and 

summer time periods from May 2001 through February 2002. The data show that overall 

energy consumption for the tested multiplex system was about 8% lower than that of the 

distributed compressor system tested in winter and 14% lower in summer – in contrast to 

the analytical projections. For summer operation the multiplex system efficiencies 

(COPs) were 34.7 and 18.5% higher than those of the distributed system for low and 

medium temperature refrigeration, respectively. For winter operation the multiplex 

system COPs were higher than those of the distributed system by 12.1 and 22.2% for low 

and medium temperature refrigeration, respectively. These results are disappointing 

because the distributed compressor system did not perform as well as the analytical 

results suggested it could. This does not mean the advanced system is fatally flawed, but 

there are several reasons for its poor performance relative to the baseline system tested. 

First, the tested distributed system used dry cooling towers for heat rejection rather than 

wet towers as assumed in the analysis. Second, the technique used to provide subcooling 

for the scroll compressors in the low temperature cabinets of the distributed system 

simply did not work. Finally, the baseline multiplex system tested was more efficient 
42

than the one modeled in the analytical case study. This had the effect of degrading the 

relative performance of the advanced vs. baseline system in the field test as compared to 

the analysis. 

A one-year TEWI analysis of each test system was conducted. The charge sizes 

were estimated at 1360 kg of refrigerant for the multiplex system and 454 kg for the 

distributed system. Actual refrigerant leakage rates for the systems were not known so 

rates of 20% and 5 % were used in the analysis for the multiplex and distributed systems, 

respectively. The analysis showed that the distributed system had about a 40% lower 

TEWI than the multiplex system over the combined summer and winter test periods 

despite its higher energy consumption. 

Field test results for the water-source heat pumps show that the amount of heat 

recovered from the refrigeration system was about 25.9% of the total heat rejection. This 

amount is considerably less than the capability of the heat pumps, which can recover as 

much as 59.8% of the rejected heat from the refrigeration system. The ambient 

conditions during this time period were very mild so that the amount of space heat 

needed for the store was much less than normally seen. Despite limited operation, the 

water-source heat pumps were able to displace approximately 37,346 m3

 of natural gas 

usage at the test store for the winter period evaluated. The value of this displacement is 

dependent upon the utility rates for electricity and gas. For this particular site, the 

estimated energy cost savings seen over this winter period were $3,171. 

 

Supermarket energy use screening level model 

The model discussed herein was developed by the US as part of its role as the 

Annex operating agent. 

 The model consists of a set of Excel® spreadsheets and was developed to provide 

tools to assist supermarket planners and engineers to compare the relative energy and 

refrigerant requirements for alternative designs of their refrigeration systems. The 

spreadsheets have been completed for multiplexed direct expansion and secondary loop 

refrigeration systems; an additional spreadsheet may be developed in the future for 

distributed refrigeration systems. 

 A users’ manual for the model is included Volume 2 as part of the US country 

report. Users must specify or select information about the display cases and walk-in 

coolers/freezers, secondary loop piping and fluids, and type of condenser. All 

calculations of fluid viscosity, friction factors, pressure drops, compressor power, and 

secondary fluid loop pumping power are built into the model. Model validation exercises 

have been performed using data from the baseline multiplex DX system tested in 

Massachusetts. Results of the validation are included in the users manual in the US 

country report. 

Overall Summary 

The five participating countries in Annex 26 have carried out a significant amount 

of research on supermarket refrigeration and heat recovery systems – an estimated $5 
43

million US total effort. A few major conclusions and observations are noted (by country 

along with some summary comments) in this section from the analyses and testing 

programs conducted. Before proceeding, however, it must be noted that these 

conclusions are subject to the assumptions used in the analyses and the particular 

locations and installations of the field test systems. These specific results should not be 

considered to be generalizable to all store sizes and locations. However, they do provide 

a good relative indication of the energy savings and TEWI reduction potential of the lowcharge refrigeration systems. 

Canada

 Hydro Quebec’s laboratory (LTE) is field-testing two advanced systems and a 

baseline multiplex DX system. One advanced approach uses a multiplex DX system with 

heat reclaim for space and water heating and ground water to supplement heat rejection. 

The other has heat pumps integrated to provide space heating for the store and subcooling 

for the refrigeration system. In winter, the discharge gas from the refrigeration 

compressors goes first through three plate heat exchangers that serve to desuperheat and 

precondense the gas. The three heat exchangers also serve as evaporators for rooftop 

heat pumps that supply space heating to the store. Using heat pumps for heat recovery 

places no minimum limit on refrigeration system condensing pressure, as is the case for 

traditional heat recovery approaches. A fourth rooftop heat pump is integrated with the 

liquid line exiting the air-cooled condenser via a fourth plate heat exchanger. This heat 

pump subcools the refrigerant leaving the condenser and uses the recovered heat for store 

space heating. Initial baseline tests in 1999-2000 showed that both advanced approaches 

achieved about 6% lower specific energy consumption (kWh/m2

/yr) compared to the 

baseline store. 

In addition CETC-Varennes is partnering with Loblaws (Canadian supermarket 

chain) to install a refrigeration showcase in a new 9000 m2 store in 2003. Refrigeration 

for all cases will be provided by a secondary loop system (the frozen food cases will have 

self-contained refrigeration systems with horizontal scroll compressors and will reject 

heat to the main secondary loop system). Refrigeration heat recovery will be used for 

store heating. 

Denmark

A propane/carbon dioxide demonstration plant has been built in a medium-size 

supermarket. Propane is used as the high temperature refrigerant (-14/+30°C) while 

carbon dioxide is used at the low temperature level (-32/-11°C). Carbon dioxide is used 

directly to perform the cooling in the freezers while a brine circuit with propylene glycol 

is used in the coolers. Energy consumption in the test store is similar to that of other new 

stores in the same chain of markets with conventional systems. The additional cost for a 

propane/carbon dioxide cascade plant for this size store is estimated at approximately 

15% of the total installation. A second test store (190 kW chilled food load, 60 kW 

frozen food load) with a cascade system using R404A as the high temperature refrigerant 

and R744 (carbon dioxide) for the lower stage has been monitored also. The CO2 is used 

to cool both frozen food and chilled food cases, thus the amount of R404A required is 

only about 120 kg or about 10% of what a conventional DX system would require 

resulting in initial refrigerant cost savings of about $37,000US due to high Danish taxes 
44

on HFCs. The total system cost was about 10% more than a conventional DX but 

showed 15-20% lower energy use than similar stores with R404A DX systems. The store 

used recovered heat from the refrigeration system for heating and required no back up 

during the test period. 

Sweden

Sweden’s work for the Annex is part of their national program Eff-Sys under a 

project “Energy Efficient Solutions for Supermarkets in Theory and Practice.” They 

have developed a computer model (Cybermart) for system predesign and have carried out 

field measurements in four supermarkets to validate the model. Sweden’s analyses 

indicate that well designed advanced secondary loop systems do not compromise energy 

efficiency compared to conventional DX systems. The analyses also show that two-loop 

secondary systems with subcooling of the low temperature loop are more efficient than 

cascade systems. Major conclusions from the field test results include the following. 

- Standard secondary loop systems with subcooling of the low temperature loop 

primary refrigerant by the medium temperature brine have efficiency advantages 

over cascade systems. 

- Night covering of display cases reduces energy consumption by 10-20% 

- Practical experience with heat recovery to date show that only about 40-70% of 

stores’ heating needs are supplied by refrigeration heat recovery, primarily due to 

very low refrigeration loads during winter and unoptimized control systems. 

- Air-conditioning is needed in Swedish stores even if the cooling season is very 

short to avoid food quality problems. When the store ambient exceeds 25 °C food 

in display cases will exceed code minimum temperatures and shelf life is halved. 

United Kingdom

 Four research activities have been completed. The first is an evaluation of 

combined heat and power, and combined cooling, heating, and power schemes for 

supermarkets. CCHP has an improved load match compared to CHP, when used with 

absorption chilling. Cascade vapor compression systems integrated with the absorption 

system provide case refrigeration. Primary energy savings are around 15%, and there is a 

TEWI reduction of more than 50%. A second study involved comparison of various 

secondary systems with standard DX systems. Results from that study indicated that 

secondary loop systems use about 30% more energy mostly due to secondary pumping 

power. TEWI for secondary systems was estimated to be 8% lower than a DX baseline 

system, assuming the reference case had a 15% annual refrigerant leakage rate and a 15 

year service life. Capital costs were estimated to be about 20% higher for the secondary 

approach. The third study was an investigation of the effect of various store conditions 

on case performance. Environmental chamber tests of low and medium temperature case 

systems show that an increase in ambient temperature from 19 to 22 °C increases system 

energy use by 20%. Increasing RH from 35 to 50% at an ambient of 22 °C increased the 

medium temperature case energy use by 15% but had no effect on the freezer case. Large 

savings were observed when the voltage was dropped from 240V to 220V, with no 

negative impact on product temperature. Finally analytical and experimental 

investigations of defrost methods and alternative control strategies were carried out. 

Studies looking at the frequency and duration of defrost cycles showed that the current 
45

cycles are suited to relatively warm internal temperatures and high relative humidities 

(22o

C and 60% RH), and for different internal conditions, a lower frequency would be 

sufficient. For electric defrost, optimized cycles and controls could reduce energy use by 

25-50%, and better termination controls on completion of defrosting would reduce energy 

use still further. 

United States

A 3720 m2 supermarket was simulated and TEWI and energy consumption 

estimates were made for a baseline air-cooled multiplex refrigeration system and 

advanced systems. Total refrigeration load was 328 kW with a refrigerant charge of 4.15 

kg/kW load. The distributed compressor, low-charge multiplex, and secondary loop 

systems (with four independent secondary loops) all achieved estimated annual energy 

savings of about 11%. Use of evaporative heat rejection was the principal driver for 

these energy savings - baseline system energy consumption was 8.2% lower with an 

evaporative condenser. The lowest TEWIs were achieved by the distributed system and 

the secondary loop systems with CO2 emission reductions of about 13 - 14 million kg, or 

57 - 60%. The low-charge multiplex system had estimated TEWI reductions of about 

24% or 43% depending upon the refrigerant loss assumption. An analysis of an 

integrated water source heat pump and distributed compressor refrigeration system 

showed about 13% operating cost savings compared to a baseline air-cooled multiplex 

refrigeration system with conventional rooftop HVAC units. 

Summary comments

Analyses carried out under the Annex 26 project and individual country programs 

have shown that both energy savings (over 10%) and TEWI reductions (up to 60%) are 

possible with low-charge refrigeration systems as compared to the most prevalent type 

baseline -- a multiplex DX system with air-cooled condensers. Savings are possible with 

distributed compressor systems, secondary loop systems, and low-charge multiplex 

systems. The secondary loop system option was the most thoroughly investigated 

advanced low-charge alternative in this Annex (all participants studied this option to 

some extent). Energy consumption comparisons with the baseline ranged from up to 

30% greater energy use to about 10% savings. Use of evaporative heat rejection 

approaches (condensers or cooling towers) to reduce condensing temperatures is a key to 

obtaining maximum energy savings for all of the systems studied. Evaporative 

condensers (or cooling towers) will impose greater maintenance efforts and costs, 

however. Proper design and implementation of advanced low-charge systems is essential 

if energy savings are to be realized. 

In general further efforts to reduce total global warming impacts (TEWI) of the 

advanced low-charge systems examined here would benefit more from reduction in 

energy usage (through efficiency increases or load reductions) than from further 

reduction in direct impact (from refrigerant losses). 

Recovery of refrigeration system rejected heat was shown to be able to provide 

from about 40% to all of space and water heat needs for the test stores examined in this 

Annex. The amount of waste heat that is effectively applied to the space and water heat 

requirements at a given site will depend upon the size of the coincident refrigeration load 

and the refrigeration/HVAC control system’s ability to effectively manage the heat 
46

recovery process. The integration of heat pumps with refrigeration systems was shown to 

be an excellent means for recovering refrigeration waste heat and reducing overall store 

energy use. Heat-pump-based heat recovery does not require the refrigeration system 

condensing temperature to be maintained artificially high to facilitate heat recovery. 

Available cost studies show installed cost premiums for secondary loop systems 

range from 0% to about 35% while distributed compressor systems show about a 15% 

premium. The low-charge multiplex approach is estimated to cost no more than current 

multiplex systems. It should be noted that the actual installed cost of any supermarket 

refrigeration system will be highly dependent upon many factors, not the least of which 

are the complex relationships between supermarket companies and equipment suppliers. 

These factors include negotiated purchasing arrangements, inclusion of display cases 

with system purchase, special system features, and unique installation requirements (site 

specific). If/when any of the advanced systems studied begin to obtain increased market 

share, the price differential with the current baseline can be expected to decrease 

somewhat due to the pressures of these market forces. 

References 

Danish law #49, December 21, 2000. Act amending the Act on taxes on certain 

chloroflouro-carbons and halons (CFC tax). (Tax on the industrial greenhouse gasses 

HFCs, PFCs and SF6). 

Bashtrash, S., 2002. Personal communication to Annex 26 participants during 

supermarket system tour near Montreal, April. 

Horton, W. T., 2002. Modeling of secondary loop refrigeration systems in supermarket 

applications, HL 2002-3, Report # 52571-1, Purdue University, W. Lafayette, IN, April. 

Hrnjak, P., 1996. “Secondary Coolants for Low Temperature Supermarket 

Refrigeration”, presentation given at EPRI Supermarket Refrigeration Design Review 

Panel Meeting, ComEd Commercial Center, Oak Brook, IL. May 3. 

Hrnjak, P., 2000. “Heat transfer issues in laminar flow of single-phase secondary 

refrigerants through the pipes,” Proceedings of the IEA Annex 26 Workshop, Stockholm, 

Sweden, October. P. Lundqvist, ed. 

Kantchev, J., 2002. . Personal communication to Annex 26 participants during 

supermarket system tour near Montreal, April. 

Kazachki, G., E. Bayoglu, and C. Gage, 1997. “Comparative Evaluation of Heat-Transfer 

Fluids for Secondary Loop Systems”, presentation given at the International Conference 

on Ozone –Protection Technologies, Baltimore, MD, November 13. 
47

Lundqvist P., ed 2000. Proceedings of the IEA Annex 26 Workshop: Advanced 

Supermarket Refrigeration/Heat Recovery Systems, IEA Heat Pump Centre, Sittard, The 

Netherlands. (Cd-rom format only), ISBN# 90-73741-39-4. 

Melinder, A., 2000. “Update on Secondary Refrigerants for Indirect Systems”, 

Proceedings of the IEA Annex 26 Workshop, Stockholm, Sweden, October. P. 

Lundqvist, ed. 

Sand J. R., Fischer S. K., and Baxter V. D., 1997. Energy and global warming impacts of 

HFC refrigerants and emerging technologies, Alternative Fluorocarbons Environmental 

Acceptability Study and U. S. Department of Energy, Washington, DC, USA. 

Sherwood, G. J., 1999. Design and analysis of secondary refrigeration systems, MS 

Thesis, University of Minnesota, March. 

Vana, J., 2000, “Comparison of secondary and multiplex DX equipment,” Proceedings 

of the IEA Annex 26 Workshop, Stockholm, Sweden, October. P. Lundqvist, ed. 

Walker D. H. and Baxter V. D., 2003. “Analysis of advanced, low-charge refrigeration 

systems for supermarkets,” ASHRAE Trans. V. 109, pt. 1. 

Acknowledgements 

The Annex Operating Agent gratefully acknowledges the support and 

contributions of the lead investigators from each Annex 26 participant: V. Minea and D. 

Giguère of Canada; Prof. H. J. H. Knudsen of Denmark, Prof. P. Lundqvist of Sweden, 

and A. Crompton and J. Palmer of the United Kingdom. Thanks are also expressed to US 

contributors S. Fischer and D. Walker. The US Department of Energy, Office of 

Building Technology, State and Community Programs sponsored U. S. activities for IEA 

Annex 26 under contract DE-AC05-00OR22725 with UT-Battelle, LLC. 
48

Appendix A – Case Study Summaries

This appendix is included to give the reader some summary details for a few of the stores 

that were the objects of case studies under Annex 26. For full details on the studies one 

should consult the individual country reports in Volume 2. 

Canada – LTE

Test store 1 (“System1-phase1”) -- METRO – Messier, Montreal 

Test store 2 (“System2-phase1”) -- IGA – Crevier, Montreal 

Conventional store (“CON”) -- METRO – Bordeleau, Montreal 

- “System1-phase1” – multiplex system with reciprocating compressors for low 

temperature loads and screw compressors for medium temperature loads. New 

total heat reclaim technique for space and hot water heating. High speed hot gas 

defrost method. 

- “System2-phase1” – improved multiplexed system with reciprocating 

compressors. Cascade heat pumps coupled with the refrigeration system to 

recover rejected heat for store heating in winter. Dedicated heat pump for 

mechanical sub-cooling. Refrigerant liquid pumping. 

- “CON” – multiplex, conventional heat reclaim and defrost. 

Store and system design details 

Store Areas and Refrigeration Capacities as function of Saturated Suction Temperatures 

System Total 

Area 

Sale 

Area 

Low/Dual Temperatures Medium Temperatures 

 - - - 37 °C - 32 °C - 29 °C -26 °C -12 °C -9 °C -7 °C -1 °C 7 °C 

 ft2 ft2 kW kW kW kW kW kW kW kW kW 

System1 38,000 28,302 7.15 28.34 8.54 13.63 - 82.77 134.4 11.58 35.43 

System2 40,900 30,838 4.61 9.02 56.33 - 3.61 118.0 155.3 9.50 - 

CON 18,000 12,200 - 35.57 - - - 45.70 78.61 - - 

- store operating hours (hours open per day): 14 hours/day. 

- total traffic or sales volume: average traffic - 10 000 per week 

- estimated total primary refrigerant charge (estimate): R-22, 1500 – 1800 kg 

- estimate of annual loss of primary refrigerant: 15% of charge 
49

”System1-phase1”: Display Cases and Walk-in Boxes – Low Temperature Rack 

(SST - Saturated Suction Temperature; CST - Condensing Saturated Temperature)

No Dimensions Description Capacity SST CST 

- ft - Btu/h °F °F 

1 8’ Frozen Tomb 10 860 -25 105 

2 12’ Frozen Tomb 13 380 -25 105 

3 15 doors Ice Cream 25 500 -25 105 

4 12 doors Grocery 18 000 -15 105 

5 14 doors Meat 21 000 -15 105 

6A 5 doors Fish 7 500 -15 105 

6B 3 doors Bakery 4 500 -20 105 

7A 26’x15’x8’ Grocery 12 500 -25 105 

7B 26’x15’x8’ Grocery 12 500 -25 105 

8 20’x13’x8’ Meat 15 900 -20 105 

9 11’x8’x8’ Bakery 8 750 -20 105 

10 12’ Frozen Tomb 3 960 -25 105 

11 Spare 10 % 18 000 -25 105 

13A 8’+2 (bouts) Ice Cream 12 200 -35 105 

13B 8’+2 (bouts) Ice Cream 12 200 -35 105 
50

”System1-phase1”: Display Cases and Walk-in Boxes – Medium Temperature 

(SST - Saturated Suction Temperature; CST - Condensing Saturated 

Temperature)

No. Dimensions Description Capacity SST CST 

- ft - Btu/h °F °F 

20 28 Fresh Meat 44 800 15 110 

21 36 Fresh Meat 57 600 15 110 

22A 14 Delicatessen 15 400 15 110 

22B 14 Delicatessen 15 400 15 110 

23A 14 Cheese 10 850 15 110 

23B 14 Cheese 10 850 15 110 

24A 8 Pizza 9 600 15 110 

24B 8 Spare 10 600 15 110 

24C 8 Spare 2 400 20 110 

24D 8 Spare 2 400 20 110 

25 36 Delicatessen 52 800 15 110 

26A 8 Fich 9 600 15 110 

26B 8 Bakery 9 600 15 110 

27 12 Start-up 9 300 15 110 

28 Spare 10 % - 26 000 15 110 

29 A/C AC offices 79 800 45 110 

30A 30’x20’x10’ Meat 17 550 20 110

30B 30’x20’x10’ Meat 17 550 20 110

31A 10’x9’x8’ Fish 8 200 20 110

31B 10’x9’x8’ Delicatessen 8 200 20 110

31C 12’x8’x8’ Bakery 8 000 20 110

32A 8 Meat 3 040 20 110 

32B 24 Delicatessen Service 9 120 20 110 

33A 26’x18’x8’ Milk 11 300 20 110

33B 26’x18’x8’ Milk 11 300 20 110

34A 36 Milk 47 700 20 110 

34B 12 Wine 15 900 20 110 

35 38 Milk 50 360 20 110 

36 40 Milk 53 000 20 110 

37 44 Beer 63 800 20 110 

38A 8 Fruits 3 800 20 110 

38B 20 Fruits 26 500 20 110 

39 48 Fruits 37 440 20 110 

40 24’x15’x8’ Fruits 20 150 20 110

41A 608 Preparing 19 760 30 110

41B 608 Preparing 19 760 30 110

42 Spare 10 % - 41 000 20 110 

43 - Sub-cooling 41 084 40 110 

TOTAL - - 901 514 - - 
51

”System1-phase1: Characteristics of the Low Temperature Compressors 

Compressor SST Model Power Capacity THR RLA kW 

- °F - hp Btu/h Btu/h A - 

Copeland -25 3DA3-0600-TFE 6 28 661 44 545 10,5 4,6 

Copeland -25 4DA3-1000-TSE 10 48 307 75 860 17,5 8 

Copeland -25 4DL3-1500-TSE 15 64 685 101 098 20,9 10,7 

Copeland -25 3DS3-1000-TFE 10 46 399 71 891 16,8 7,4 

Copeland -35 3DF3-0900-TFE 9 29 470 49 160 16,5 5,7 

TOTAL - - 50 217 522 342 554 82,2 36,4 

”System1-phase1: Characteristics of the Medium Temperature Compressors 

Compressor TSS Model Power Capacity THR RLA kW 

- °F - hp Btu/h Btu/h A - 

Carlyle 15 06TAH078C2EA 40 323 173 408 548 51,4 37,2 

Carlyle 15 06TAH078C2EA 40 323 173 408 548 51,4 37,2 

Carlyle 15 06TAH078C2EA 40 323 173 408 548 51,4 37,2 

TOTAL - - 120 969 519 1 225 644 154,2 111,6 

”System1-phase1: Characteristics of the Air-Cooled Condensers 

 Low Temperature Medium Temperature 

Model CLD057 FDFCT-180-R 

Capacity 424 800 Btu/h 1 800 000 Btu/h 

AC Circuit; Control 8,4 A; 575 V/3 ph; 120 V/30 A 575 V/3 ph; 120 V/30 A 
52

”System 1-phase 1”: Example of Monitored Parameters 

Type of Parameter Observation 

Store Total Power (kW) - 

Compressor’s Power (kW) Both DT and MT Compressor 

Total Electrical Power (kW) LT and MT Compressor 

Suction Pressures (psig) Both DT and MT Refrigeration Line 

Discharge Pressures (psig) Both DT and MT Refrigeration Line 

Suction Temperatures (°C) Both DT and MT Refrigeration Line 

Discharge Temperatures (°C) Both DT and MT Refrigeration Line 

Liquid Temperatures (°C) Both DT and MT Refrigeration Line 

Liquid Pressures (psig) Both DT and MT Refrigeration Line 

Cold Water Temperatures (°C) Entering/Leaving Storage Tank 

Groundwater Temperature (°C) Entering/Leaving Plate Heat Exchanger 

Control Valves Status (on/off) Defrost, Heat Recovery, Flushing, etc. 

”System 1-phase 1”: Example of Calculated Thermodynamic and Energy Parameters 

Type of Parameter Observation 

Mass Enthalpies (kJ/kg) Both DT and MT Line (Suction, Discharge, Liquid) 

Total Refrigerant Flow-rate (kg/s) Both DT and MT Refrigeration Line 

Title of Two-phase Refrigerant (%) Both DT and MT Refrigeration Line 

Refrigeration Power (kW) Both DT and MT Refrigeration Line 

Condensing Power (kW) Both DT and MT Refrigeration Line 

Refrigeration Load (kWh) Both DT and MT Refrigeration Line 

Energy Rejected from Condensers (kWh) Both DT and MT Refrigeration Line 

Refrigeration Efficiency Both DT and MT Refrigeration Line 

Saturated Suction (SST) and Discharge (SDT) Temperatures (°C) Both DT and MT Refrigeration Line 

Compressor Consumed Energy (kWh) Both DT and MT Refrigeration Line 

Energy Consumption (kWh) Both DT and MT Refrigeration Line 

Operating Time: Heat Reclaim, Defrost; Condenser; Geothermal Space and Water Heating Systems 

Defrost Cycle Status and Operating Time (sec/hr; hrs/day) Fresh Meat Case and Cold Room; Ice Cream 
53

 “System2 – phase1”: Characteristics of Dual (Low & Medium) Temperature Compressors (THR - 

Total Heat Rejection; EER – Energy Efficiency Ratio) 

Compressor TSS Model Condensing 

temperature 

SUCTION 

Capacity 

THR EER 

- °F - °F Btu/h Btu/h - 

COP/Discuss (22) L/T -35 2DA3-060E-TFE-200 100 16 000 27 270 5,07 

COP/Discuss (22) L/T -25 3DB3-075E-TFE-200 100 37 000 55 490 6,76 

COP/Discuss (22) L/T -20 2DL3-040E-TFE-200 100 25 000 36 380 7,07 

COP/Discuss (22) L/T -20 3DS3-100E-TFE-200 100 56 000 82 320 7,1 

COP/Discuss (22) L/T -20 3DS3-100E-TFE-200 100 56 000 82 320 7,1 

COP/Discuss (22) L/T -20 4DA3-100E-TSE-200 100 59 000 88 510 6,93 

COP/Discuss (22) M/T 15 3DB3-100E-TFE-200 100 107 000 136 100 12,71 

COP/Discuss (22) M/T 15 3DS3-150E-TFE-200 100 142 000 181 100 12,32 

COP/Discus (22) M/T 15 4DA3-2000-TSE-200 100 150 000 190 800 12,54 

Total -35 - - 16 000 - - 

 -25 - - 37 000 - - 

 -20 - - 195 000 - - 

 15 - - 399 000 - - 

Total Unit - - - - 880 200 - 

 “System2 – phase1”: Characteristics of Medium Temperature Compressors (THR - 

Total Heat Rejection; EER – Energy Efficiency Ratio) 

Compressor TSS Model Condensing 

Temperature

SUCTION 

Capacity 

THR EER 

- °F - °F Btu/h Btu/h - 

COP/Discuss (22) M/T 20 2DA3-075E-TFE-200 100 83 000 103 600 13,78 

COP/Discuss (22) M/T 20 3DB3-100E-TFE-200 100 120 000 149 300 13,82 

COP/Discuss (22) M/T 20 3DL3-150E-TFE-200 100 159 000 199 200 13,34 

COP/Discuss (22) M/T 20 4DH3-2500-TSE-200 100 222 000 276 800 13,69 

Total 20 - - 584 000 - - 

 - - - - 729 000 - 
54

 “System2 – phase1”: Display Cases and Walk-in Boxes –Low Temperature 

No. Description Capacity SST 

- - Btu/h °F 

13 Ice Cream 15 720 - 35 

11 Pastry and Frozen Meat 5 280 - 25 

4 Ice Cream 5 doors 25 500 - 25 

10 Bakery and Frozen Crustaceous 16 680 - 20 

8 Frozen Meat and Fish 21 720 - 20 

3 Frozen Products 10 860 - 20 

5 Frozen Products 23 160 - 20 

6 Frozen Meat 17 640 - 20 

7 Frozen Bakery 20 760 - 20 

9 Frozen Meat and Fish 27 000 - 20 

2 Frozen Products 24 000 - 20 

1 Frozen Products 22 500 - 20 

12 Spare 7 894 - 20 
55

 “System2 – phase1”: Display Cases and Walk-in Boxes – Medium Temperature 

(SST : Saturated Suction Temperature) 

No. Description Capacity SST 

- - Btu/h °F 

41 Combined Service. “BAS” 12 320 10 

41 Shelves of Prepared Meals 7 200 15 

41 Flowers 6 240 20 

52 Prepared Meals 5 080 15 

52 Delicatessen and mets préparés 12 420 20 

31 Fresh Meat 30 280 15 

30 Fresh Meat 44 800 15 

32 Fresh Meat, Fish and Crustaceous 33 960 15 

33 Delicatessen 52 800 15 

34 Delicatessen 38 400 15 

38 Prepared Meals 43 200 15 

37 Pizza 31 200 15 

38 Prepared Meals 11 670 20 

35 Wine 26 500 20 

35 Delicatessen 26 500 15 

40 Vinaigrettes 37 100 20 

46 Fruits and Vegetables 43 680 20 

36 Fruits and Vegetables 18 720 20 

42 Fruits and Vegetables 44 160 20 

50 Milk Products 50 340 20 

49 Milk Products 31 800 20 

48 Milk Products 47 700 20 

53 Milk Products 71 520 20 

51 Beer 68 900 20 

44 Milk 22 650 20 

43 Fresh Meat and Fish 39 090 20 

45 Bakery; Fruits and Vegetables 36 990 20 

47 Preparation Area 32 400 30 
56

 “System2 – phase1”: Example of Monitored Parameters 

Type of Parameter Observation 

Store Total Power (kW) - 

Compressor Power (kW) Each Low and Dual Compressor 

Rack Power (kW) LT and MT Compressor 

Heat Recovery and Sub-Cooling Heat Pump Power (kW) Each Unit 

Suction Pressures (psig) Very Low, Low and Medium 

Common Discharge Pressures and Defrost Discharge (psig) Each Rack 

Suction Temperatures (°C) Very Low, Low and Medium 

Common Discharge Temperatures (°C) DT/MT 

Defrost Discharge Temperatures (°C) DT/MT 

Liquid Defrost Return Temperatures (°C) DT/MT 

Liquid Temperatures (°C) DT/MT 

Liquid Pressures (psig) DT/MT 

Liquid Pump Pressures (psig) Entering/Leaving each LP 

Control Valves Status (on/off) Defrost, Heat Recovery, Flushing, etc. 

Hot Gas Temperature at Plate Heat Exchangers Outlet (°C) Each Unit 

Natural Gas Valve Status (on/off) Each Unit 

 “System2 – phase1”: Example of Calculated Parameters 

Type of Parameter Observation 

Mass Enthalpies (kJ/kg) LT/DT (Suction, Discharge and Liquid0 

Compression Mass Enthalpy Variation (kJ/kg) LT/DT 

Saturated Suction and Discharge Temperatures (°C) Low and Dual Temperature 

Compressor Energy Consumption (kWh) Each Low and Dual Compressor 

Total Energy Consumption (kWh) Each Low and Dual Compressors 

Heat Pumps Operating Time (sec/hr; hrs/day) Space and Water Heating Systems 

Defrost Status and Duration (sec/hr; hrs/day) Fresh Meat Case and Cold Room; Ice Cream 
57

 “System CON” : Display Cases and Walk-in Boxes – Low Temperature 

Circuit Description Capacity Operation Temperatures 

- - Evaporation Condensing 

 Btu/h °F °F 

1 Different 

Products 

25 500 - 25 105 

2 Ice Cream 22 100 - 25 105 

3 Different Product 11 500 - 25 105 

4 Fish 25 680 - 25 105 

5 Meat 2 640 - 25 105 

7 Meat 19 260 - 25 105 

8 Meat 9 000 - 25 105 

9 Pastry 5 700 - 25 105 

Total - 121 380 - - 
58

 “System CON” : Display Cases (15 °F - Medium Temperature) 

Circuit Description Capacity Operation Temperatures 

- - Evaporation Condensing 

 Btu/h °F °F 

19 Fresh Meat 57 600 15 110 

11 Delicatessen 16 140 15 110 

12 Fish 9 600 15 110 

13 Mets cuisinés 24 000 15 110 

14 Delicatessen 14 400 15 110 

15 Cheese 10 800 15 110 

20 Meat 23 400 15 110 

Total - 155 940 - - 

 “System CON” : Display Cases (20 °F - Medium Temperature) 

Circuit Description Capacity Operation Temperatures 

- - Evaporation Condensing 

 Btu/h °F °F 

17 Pastry 9 600 20 110 

18 Farm Delicatessen 10 500 20 110 

18 Milk 63 600 20 110 

20 Beer 21 200 20 110 

21 Jus/Vegetables 37 100 20 110 

21 Refrigerated Table 5 600 20 110 

22 Fruits/Vegetables 40 300 20 110 

23 Milk 11 500 20 110 

24 Delicatessen 4 500 20 110 

25 Fruits/Vegetables 10 000 20 110 

26 Beer 19 000 20 110 

27 Meat 24 505 20 110 

28 Fruits/Vegetables 10 800 20 110 

Total - 268 205 
59

 “System-CON”: Monitored Parameters 

Type of Parameter Observation 

Compressor Power (kW) LT and MT Compressors 

Electrical Oven Power – Cooking (kW) Ovens #1, 2 and 3 

Suction Temperature (°C) Low and Medium Temperature Racks 

Common Discharge Temperature (°C) Each Rack 

Hot Gas Temperature – Heat Reclaim (°C) Entering and Leaving HR Coil 

Defrost Valves Status (on/off) Freezing Products; Fresh Meat 

Based on the measured parameters, several thermodynamic, energy and performance calculations 

were performed (

). 

 “System-CON”: Thermodynamic, Energy and Performance Calculated Parameters 

Type of Parameter Observation 

Energy Consumption of Compressor (kW) Low and Medium Temperature Compressors 

Total Refrigeration Consumed Energy (kWh) Low and Medium Temperature Compressors 

Cooking Energy Consumption (kWh) Each Electrical Oven 

Duration and Defrost Status (sec/hr; hrs/day) Four Refrigeration Display Cases 
60

Annual energy consumption 

Total Store’s Natural Gas Consumption vs. Sale Area 

Syste

m 

Sale 

Area Total Consumption Equivalent Energy 

Total Natural Gas 

Cost 

(CAN$) 

2 ft 3 m /y 3

m / 2 ft GJ/y 

(kWh/y) 

GJ/ 2 ft /y 

(kWh/ 2 ft /y)

$/y $/ 2 ft /y 

Syst-1 28,302 45,179 1.59 1.6 

(4.4* 6 10 )

5.6* 5 10−

(155) 

22,534 0.80 

Syst-2 30,838 35,625 1.15 1.27 

(3.5* 6 10 )

4.12* 5 10−

(113) 

17,860 0.58 

Total Store Annual Electrical Energy Consumption and Costs (Including Demand 

Power Cost) vs. Sale Area (CAN$) 

System Sale 

Area 

Total Store Electrical Energy 

Consumption 

(including compressors and heat 

pumps) 

Total Store Electrical Energy Cost 

(Power : 11,97 $/kW) 

2 ft kWh/y kWh/ 2 ft .y vs. 

CON.1

(%) 

$/y $/ 2 ft .y vs. CON.1 

(%) 

Syst-1 28,302 2,770,873 97.9 + 2.4 157,.704 5.6 - 12.5 

Syst-2 30,838 2,768,775 89.8 - 5.7 163,411 5.3 - 17.2 

CON.1 12,200 1,095,462* 95.5* - 78 440* 6.4* - 

* Without Electrical Cooking 
61

Annual Compressor Electrical Energy Consumption and Costs (Including Power Demand 

Cost) – vs. Sale Area (CAN$)

Syste

m 

Sale 

Area 

Compressors Total Energy Consumption 

(Compressors only) 

Total Electrical Energy 

Cost 

(Compressors only) 

2 ft - kWh/y % of Total

Store 

Cons. 

kWh/ 2 ft .y $/y $/ 2 ft .y 

Syst-1 28,302 Low Temp. 264,906 9.6 9.4 15,077 0.53 

 Medium T. 637,299 23 22.5 36,262 1.28 

 TOTAL 

compressors 

902,205 32.5 31.9 51,339 1.81 

Syst-2 30,838 Low/Dual T. 423,291 15.3 13.7 24,982 0.81 

 Medium T. 206,298 7.5 6.69 12,172 0.39 

 TOTAL 

compressors 

629,589 22.7 20.4 37,154 1.2 

 Heat Pumps 313,805 11.3 10.2 18,521 0.6 

 TOTAL 943,394 34.0 30.6 55,675 1.8 

CON. 12,200 TOTAL 

compressors 

355,721 30.6* 29.2 23,995 1.97 

* Excluding electrical cooking. 
62

Denmark

Fakta market in Beder, Denmark (newest store in the Fakta chain as of July 2001) 

Store details 

Size - 490 m2 sales area (720 m2

 total area) 

Operating hours/day – 12 hours/day 

Lighting load – 12.18 kW indoor; 1 kW outdoor 

Refrigeration system 

Type – This is a cascade system using propane as the refrigerant for the higher 

temperature loads (–14/25°C) and CO2 as the refrigerant for the lower temperature 

loads (–32/–10°C). The propane cycle rejects heat through an air-cooled 

condenser at the roof of the supermarket while the CO2 condenses in a cascade 

heat exchanger that also serves as one of two evaporators for the propane cycle. 

CO2 is used directly (direct expansion) in the low-temperature (LT) cabinets and 

cold stores of the supermarket, while the propane circuit chills propylene glycol 

brine in its second evaporator. The glycol is pumped in a closed system to the 

high-temperature (HT) cabinets and cold stores. 

Design capacity - 33 kW at -14°C evaporating temperature, propane circuit (21 

kW for HT cabinets, 12 kW for CO2 circuit condenser); 10 kW at -32°C 

evaporating temperature, CO2 circuit 

Primary refrigerant type and charge – 10 kg propane; 6 kg CO2

Secondary refrigerant type and charge – 120 l propylene glycol 

Estimated annual losses – propane circuit, 2% of charge; CO2 circuit, 15% of 

charge 

Propane system components 

Description Supplier 

2 compressors Bitzer, 4T-8.2P 

Air-cooled condenser ECO FCE 071C63 

Plate heat exchanger R290/PGl: 21 kW, Swep, B25x70 

Plate heat exchanger R290/R744: 10 kW, Swep V27x80 HP 

Plate heat exchanger R290/R290: 2 kW, Swep, B12x70 

Plate heat exchanger R290/PG: 3 kW, Swep B27x50 

Electronic expansion valve Danfoss ETRE – 30 kW 

Electronic expansion valve Siemens Staefa – 14 kW 

CO2 system components 

Description Supplier 

Compressor for CO2 Bitzer, X2KC-3.2, BSE 55 

Expansion valve AKV (nozzle 2-4) 

Safety pressure switches KP 5 

Filter dryer DU 303 
63

Display cases 

Air on Refrigeration load [W] Dimensions 

 

[°C] 

Medium 

(+2°C) 

Low 

(-20°C) 

LxBxH [mm3] 

Storage (glass doors to shop) -2°C 4500 - 5100x3300x2360 

Storage, frozen food (doors) -26°C - 4200 3600x3300x2450 

Storage, ice cream (doors) -28°C 2800 2100x3300x2450 

Vertical cabinet (no doors) -2°C 14400 - 10 m in length 

Horizontal case -27°C 2200 3750x1000x960 

Horizontal case -27°C 800 1985x1000x960 

Horizontal case -2°C 1400 3750x1000x960 

Horizontal case -2°C 700 1985x1000x960 

Total 21000 10000 

* Type of defrost method: Electric resistance 

HVAC system 

Type – ventilation only; 2.62 kW power input 

System operating data 

Saturated discharge and suction temperatures 

Propane system: 

Condensation Temperature: 

Evaporation temperature: 

Capacity (design): 

25°C (avg.)/ 22°C (min) / 30°C (max) 

-14°C (avg.) 

33 kW 

CO2 system: 

Condensation Temperature: 

Evaporation temperature: 

Capacity (design): 

-11°C (avg.) 

-32,5°C (avg.) 

10 kW 

Brine PG(40%): 

Outlet temperature: 

Return temperature: 

Capacity (design): 

-8°C 

-2°C 

21 kW 

 

System energy use and weather conditions 

Total store energy consumption, July 2001 

 pcs kW/pcs Hours/day kWh/ July 

Refrigeration plant 1 - 24 8682 

In store lighting 210 0,058 12 4531 

Out store lighting 10 0,1 12 372 

Ventilation store 1 2,62 14 1137 

Tellers 3 0,4 12 446 

Bottle machine 1 1 1 31 

Various 1 2 12 372 

Total 15571 
64

Refrigeration system energy consumption detail, July 2001 

Consumer Energy consumption 

[kWh for July] 

Propane compressor 1 3565 

Propane compressor 2 1240 

CO2 compressor 605 

Cases and cabinet: 

Lighting 107 

Ventilation 487 

Storage rooms: 

Lighting 391 

Ventilation 744 

Brine pump: 126 

Heaters (Edge): 1116 

Defrost (cases, cabinet and storage) 84 

Ventilation condenser 140 

Electronics 78 

Total 8682 

The average outdoor temperature during the July 2001 period was 17.9 oC. Store indoor 

temperature averaged 22.5 oC. Energy consumption of the refrigeration system in the 

Fakta Beder store compared to eight similar Fakta supermarkets with conventional 

R404A direct expansion refrigeration systems is illustrated in Figure A1. The time 

period plotted begins in July 2000 and ends in February 2002. 
65

0

1000

2000

3000

4000

5000

6000

7000

8000

9000

10000

July

September

November

January

Marts

May

July

September

November

January

Energy consumption [kWh]

Lower 25%

Upper 25%

New Fakta in Beder

Figure A1 – Comparison of energy use of Fakta Beder store (with cascade 

refrigeration system) to that of eight other Fakta stores (with conventional R404A 

DX refrigeration systems); July 2000 through February 2002. 
66

United States

Test store – Distributed compressor refrigeration system 

Size – 4830 m2 sales area 

Operating hours – 168 hrs/week 

Refrigeration system

The system consists of several small compressor cabinets distributed throughout the store 

and located close to the loads they serve. Heat rejection is done via a water/glycol loop 

that connects the compressor cabinets to roof-mounted cooling towers (dry type). Water 

source heat pumps are integrated into the condenser cooling loop and recover the 

refrigeration waste heat in winter to heat the store space. 

Distributed Refrigeration System Compressors 

(capacities are in kW for saturated evaporating and condensing temperatures shown) 

All compressor cabinets use refrigerant R-404A 

Cabinet A Low Temperature Cabinet C Low 

Temperature 

Compressors Capacity: -32oC & 41oC Capacity: -27oC & 41oC 

A-1 ZF15K4 4.8 C-1 ZF15K4 6.3 

A-2 ZF15K4 4.8 C-2 ZF15K4 6.3 

A-3 ZF15K4 4.8 C-3 ZF15K4 6.3 

A-4 ZF18K4 5.7 C-4 ZF15K4 6.3 

A-5 ZF18K4 5.7 C-5 ZF15K4 6.3 

Cabinet B Low Temperature Cabinet D Low 

Temperature 

 Capacity: -32oC & 41oC Capacity: -32oC & 41oC 

B-1 ZF13K4 3.9 D-1 ZF15K4 5.2 

B-2 ZF13K4 3.9 D-2 ZF15K4 5.2 

B-3 ZF15K4 4.8 D-3 ZF15K4 5.2 

 Capacity: -26oC & 41oC D-4 ZF15K4 5.2 

B-4 ZF15K4 6.1 D-5 ZF15K4 5.2 

B-5 ZF18K4 7.3 D-6 ZF18K4 6.2 

B-6 ZF18K4 7.3 

Cabinet E Medium 

Temperature 

Cabinet G Medium 

Temperature 

 Capacity: -6.7oC & 43oC Capacity: -8.3oC & 43oC 

E-1 ZS38K4 12.0 G-1 ZS38K4 11.3 

E-2 ZS38K4 12.0 G-2 ZS38K4 11.3 

E-3 ZS45K4 14.4 G-3 ZS38K4 11.3 

E-4 ZS45K4 14.4 G-4 ZS38K4 11.3 

E-5 ZS45K4 14.4 G-5 ZS38K4 11.3 

E-6 ZS45K4 14.4 G-6 ZS45K4 13.5 
67

 

Cabinet F Low/Medium Temperature Cabinet H Low/Medium Temperature 

 Capacity: -6.7oC & 43oC Capacity: -15oC & 43oC 

F-1 ZF09K4 H-1 ZS26K4E 6.2 

 Capacity: -8.3oC & 43oC H-2 ZS26K4E 6.2 

F-2 ZS30K4 9.2 Capacity: 1.7oC & 43oC 

F-3 ZS30K4 9.2 H-3 ZS38K4E 15.9 

F-4 ZS30K4 9.2 H-4 ZS38K4E 15.9 

F-5 ZS38K4 11.3 H-5 ZS45K4E 19.1 

Cabinet I Medium 

Temperature 

Cabinet J Low/Medium Temperature 

 Capacity: -8.3oC & 43oC Capacity: -6.7oC & 43oC 

I-1 ZS38K4E 12.0 J-1 ZF11K4E 3.6 

I-2 ZS38K4E 12.0 J-2 ZF13K4E 4.2 

I-3 ZS45K4E 14.4 J-3 Zf13K4E 4.2 

I-4 ZS45K4E 14.4 

I-5 ZS45K4E 14.4 Capacity: -9.4oC & 43oC 

 J-4 ZS38K4E 10.8 

 J-5 ZS38K4E 10.8 

 J-6 ZS38K4E 10.8 

Primary refrigerant type and charge – R-404A; estimated at about 410 kg total 

Secondary refrigerant type and charge – ethylene glycol on condenser fluid loop; 

total charge unknown 

Estimated annual losses – primary refrigerant annual loss estimated at 5% of total 

charge 

Display case line ups, storage and preparation room areas, and miscellaneous loads 

connected to compressor cabinets

Circuit Case Length, No. of Doors, or 

Walk-in Floor Area 

Design Refrigeration Load 

(kW) 

Cabinet A Sat. Suction Temp -32oC 

Walk-in Freezer 94.1 m2 22.3

Cabinet B Sat. Suction Temp -32oC 

Walk-in Freezer 47 m2 11.1

Cabinet B Sat. Suction Temp -26oC 

Reach-in Frozen Food 23 Doors 10.1

Cabinet C Sat. Suction Temp -26oC 

Reach-in Frozen Food 52 Doors 19.8

Multi-Deck Frozen Food 2.4 m 3.5

Cabinet D Sat. Suction Temp -32oC 

Multi-Deck Frozen Meat 17 m 16.7

Multi-Deck Frozen Fish 8.5 m 7.0

Walk-in Frozen Fish 13.4 m2 3.5

Cabinet E Sat. Suction Temp -6.7oC 

Multi-Deck Dairy 36.6 m 47.1

Multi-Deck Cheese 4.3 m 5.5
68

Walk-in Dairy 79.6 m2 18.2

Cabinet F Sat. Suction Temp -8.3oC 

Service Deli 25.6 m 7.6

Multi-Deck Cheese 6.1 m 8.4

Isle Deli 4.9 m 11.3

Walk-in Deli 14.2 m2 4.5

Walk-in Raw 4.5 m2 1.4

Cabinet F Sat. Suction Temp -29oC 

Walk-in Deli Freezer 7.5 m2 1.8

Cabinet G Sat. Suction Temp 8.3oC 

Single-Deck Meat 7.3 m 3.3

Multi-Deck Meat 20.7 m 26.9

Floral Display 7.3 m 9.1

Multi-Deck Deli 8.5 m 12.4

Single-Deck Fish 4.9 m 5.2

Walk-in Fish 10.2 m2 3.2

Cabinet H Sat. Suction Temp 35oF 

Meat Prep Room 6,144 ft2 153,600

Cabinet H Sat. Suction Temp -15oC 

Ice Maker 10.6

Cabinet I Sat. Suction Temp -6.7oC 

Multi-Deck Produce 35.4 m 36.0

Table Fish 4.6 m 2.2

Walk-in Produce 91.4 m2 19.5

Cabinet J Sat. Suction Temp -9.4oC 

Walk-in Meat 98.1 m2 23.2

Walk-in Bakery 11.1 m2 5.8

Cabinet J Sat. Suction Temp -29oC 

Walk-in Freezer Meat 9.3 m2 2.2

Walk-in Freezer Bakery 13.4 m2 3.2

Multi-Deck Bakery 4.3 m 4.6

All Compressor Cabinets employ R-404A as the refrigerant 

Baseline store – Conventional multiplex direct expansion refrigeration system 

Size – 4830 m2 sales area 

Operating hours – 168 hrs/week 

Refrigeration system

The system consists of three compressor racks - one for the low temperature load with 

seven compressors and two for the medium temperatures loads with six compressors 

each. All compressors are reciprocating type and the rack with highest suction 

temperature (Rack C) supplies mechanical subcooling to the other two racks. 
69

Multiplex Refrigeration 

System Compressors 

Low Temp System Rack A 

Main Compressors -29oC SST R-404A 

Comp1 4DL3-150E-TSK 

Comp2 4DL3-150E-TSK 

Comp3 4DL3-150E-TSK 

Comp4 4DL3-150E-TSK 

 

Satellite Compressors –34oC SST R-404A 

Comp5 3DF3-090E-TFD 

Comp6 3DF3-090E-TFD 

Comp7 2DL3-040E-TFD 

 

Medium Temp System 

Rack B 

Main Compressors –6.7oC SST R-22 

Comp1 3DA3-0750-TFD 

Comp2 3DA3-0750-TFD 

Comp3 3DA3-0750-TFD 

Comp4 2DD3-0500-TFD 

 

Satellite Compressors –8.3oC SST R-22 

Comp5 3DS3-1500-TFD 

Comp6 3DF3-1200-TFD 

 

Medium Temp System 

Rack C 

Main Compressors –6.7oC SST R-22 

Comp 1 3DS3-1500-TFD 

Comp 2 3DA3-0750-TFD 

 

Satellite Compressors 1.7oC SST R-22 

Comp 3 3DS3-1500-TFD 

Comp 4 3DS3-1500-TFD 

Comp 5 3DF3-1200-TFD 

Comp 6 2DL3-0750-TFD 

 
70

Display case line ups, storage and preparation room areas, and miscellaneous loads 

connected to compressor racks

Circuit Case Length, No. of Doors, or 

Walk-in Floor Area 

Design Refrigeration Load 

(kW) 

Rack A Low Temperature - Sat. Suction Temp -29oC Refrigerant - R404A 

Reach-in Frozen Food 76 Doors 33.4

Multi-Deck Frozen Meat 12.2 m 17.7

Multi-Deck Frozen Fish 5.5 m 8.0

Multi-Deck Bakery 3.7 m 5.3

Walk-in Freezers 85.9 m2 20.3

Ice Maker 10.6

Rack A Low Temperature - Sat. Suction Temp -34oC 

Coffin Ice Cream 31.6 m 21.9

Coffin Shrimp 4.9 m 1.7

Walk-in Ice Cream 16.7 m2 5.0

Rack B Medium Temperature - Sat. Suction Temp -6.7oC Refrigerant - R22 

Reach-in Dairy 46 Doors 18.3

Single-Deck Produce 17.1 m 15.6

Multi-Deck Produce 7.3 m 9.4

Service Meat & Deli 19.5 m 6.9

Multi-Deck Deli 14.6 m 20.5

Coffin Cheese & Deli 4.9 m 7.3

Tables Fish 6.1 m 2.9

Cooler Floral 13 m2 3.1

Floral Display 2.4 m 3.0

Rack B Medium Temperature - Sat. Suction Temp -8.3oC 

Bakery Display 4.6 m 3.7

Multi-Deck Meat 20.7 m 26.9

Isle Cheese & Deli 17.1 m 18.6

Isle Ready Meals 15.8 m 16.8

Rack C Medium Temperature - Sat. Suction Temp -6.7oC Refrigerant - R22 

Walk-in Coolers 204.0 m2 57.9

Rack C Medium Temperature - Sat. Suction Temp 1.7 °C 

Meat Prep Room 93.8 m2 41.4

Produce Prep Room 48.3 m2 11.4

Mechanical Subcooling 106.0

Primary refrigerant type and charge – 

R-404A in low temperature system; estimated at about 454 kg total 

R-22 in medium temperature system; estimated 908 kg total 

Estimated annual losses –refrigerant annual loss estimated at 20% of total charge 
71

System operating data & measured energy use 

Energy Consumption Comparison between Multiplex and Distributed Refrigeration 

Systems – May 2001 through February 2002 

 Energy Consumption (kWh/day) 

 Distributed Multiplex Difference % 

Difference 

May-Aug 

LT 

Compressor 

1306.4 1290.2 16.2 1.2 

MT 

Compressor 

1594 1201.2 392.8 24.6 

Heat Reject 702.7 608.4 94.3 13.4 

Total 3603.2 3100.5 502.7 14.0 

Nov - Feb 

LT 

Compressor 

863.2 957.9 -94.7 -11.0 

MT 

Compressor 

951.3 635.9 315.4 33.2 

Heat Reject 316.1 364.4 -48.3 -15.3 

Total 2130.5 1958.2 172.3 8.1 

 Average State Points, Refrigeration Loads, and efficiencies for the Multiplex and 

Distributed Refrigeration Systems – May 2001 through February 2002 

 May - Aug Nov – Feb 

Low Temp Distributed Multiplex Distributed Multiplex 

SST (oC) -26.6 -28.4 -27.3 -29.0 

SDT (oC) 32.2 28.2 16.1 21.7 

Return Temp (oC) -11.4 -16.6 -8.7 -11.7 

Liquid Temp (oC) 26.2 14.3 13.8 10.2 

Ref Load (kW) 100.0 133.5 90.6 113.1 

Comp Power (kW) 54.4 53.8 36.0 39.9 

COP 1.84 2.48 2.52 2.83 

 

Med Temp Distributed Multiplex Distributed Multiplex 

SST (oC) -8.8 -5.1 -10.3 -6.6 

SDT (oC) 33.6 28.5 22.3 19.4 

Return Temp (oC) 8.0 9.4 6.8 9.6 

Liquid Temp (oC) 29.4 20.1 19.5 14.6 

Ref Load (kW) 213.9 191.4 158.6 129.8 

Comp Power (kW) 66.4 50.1 39.6 26.5 

COP 3.22 3.82 4.01 4.90 
72

TEWI Analysis for the Field Test Results – May 2001 through February 2002 

 Refrigerant 

Leakage 

Energy TEWI (kg CO2) 

 (kg) (kWh) Direct Indirect Total 

 R-404A R-22 

May - August 

Multiplex 30.3 60.3 381,300 202,055 247,845 449,900 

Distributed 7.7 443,169 24,895 288,060 312,954 

Difference 136,046 

 

Nov - Feb 

Multiplex 30.3 60.3 234,960 202,055 152,724 354,779 

Distributed 7.7 255,600 24,895 164,034 188,929 

Difference 165,850 

 

Multiplex leak rate estimated at 20%/yr 

Distributed leak rate estimated at 5%/yr 

Field Test Results for the Water-Source Heat Pumps 

Space Heating Performance (November 2001 - February 2002) 

 Average Heat 

Recovered (kW) 

Average Space 

Heat Load (kW) 

Heat Pump 

Energy (kWh) 

Gas Displaced 

(m3

) 

November 01 72.7 95.6 17,236 8,076 

 

December 01 86.0 111.7 20,531 9,807 

January 02 88.5 115.7 21,004 10,080 

February 02 90.0 118.6 20,295 9,383 

Total 79,066 37,346 

Amount of heat recovered estimated at 25.9% of total refrigeration system heat rejection 

during the period. The heat pumps could recover up to about 60% of the rejected heat.