**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 

# 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

<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>

# 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]{.underline}** 6

*Estimates of cost premium* 9

*Estimates of TEWI reduction* 10

**[Distributed compressor refrigeration system]{.underline}** 10

*Estimates of cost premium* 12

*Estimates of TEWI reduction* 12

**[Low charge multiplex systems]{.underline}** 12

*Estimates of cost premium* 14

*Estimates of TEWI reduction* 14

**[Advanced self-contained systems]{.underline}** 14

*Estimates of cost premium* 15

*Estimates of TEWI reduction* 15

**[HVAC integration approaches]{.underline}** 15

**Country report summaries** 18

**[Canada - CANMET Energy Technology Centre --Varennes (CETC--Varennes)
of Natural Resources Canada]{.underline}** 18

**[Canada - Hydro-Quebec Research Institut, Laboratoire des technologies
de l'énergie *(LTE*)]{.underline}** 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]{.underline}** 25

**[Sweden]{.underline}** 27

**[United Kingdom]{.underline}** 33

*Secondary Systems* 33

*Defrost* 34

*CHP/CCHP* 35

*Case and Store Environment Interaction* 36

*UK Overall Conclusions* 37

**[United States]{.underline}** 37

*System analysis summary* 38

*Payback estimates* 40

*Field test summary* 41

**Supermarket energy use screening level model** 42

**Overall Summary** 42

**[Canada]{.underline}** 43

**[Denmark]{.underline}** 43

**[Sweden]{.underline}** 44

**[United Kingdom]{.underline}** 44

**[United States]{.underline}** 45

**[Summary comments]{.underline}** 45

**References** 46

**Acknowledgements** 47

**Appendix A - Case Study Summaries** 48

**[Canada - LTE]{.underline}** 48

**[Denmark]{.underline}** 62

**[United States]{.underline}** 66

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

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)]{.underline} 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 CO~2~) 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.

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 m^2^ to 3000 m^2^ or somewhat larger.
Stores are typically larger in Canada and the US ranging from a minimum
of about 1000 m^2^ to 10000 m^2^. 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 air-refrigerant
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).

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.

![](media/image1.wmf){width="5.992361111111111in"
height="3.859027777777778in"}

# 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.

![](media/image2.jpeg){width="6.0in" height="4.500694444444444in"}

**Figure 2 -- Total equivalent warming impact (TEWI) for low temperature
supermarket refrigeration in North America**

![](media/image3.wmf){width="5.65in" height="4.419444444444444in"}

# 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 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]{.underline}**

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.

![](media/image4.wmf){width="5.65in" height="4.472916666666666in"}

## **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 CO~2~) condenser and
also chills a brine (propylene glycol), which is circulated to the
chilled food cases. The CO~2~ refrigerates the frozen food cases through
direct expansion evaporators. Other combinations of refrigerants are
possible for this system concept.

![](media/image5.wmf){width="5.121527777777778in"
height="3.0743055555555556in"}

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

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 CO~2~ 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 CO~2~ 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 NH~3~ 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 (NH~3~ 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 NH~3~ 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 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]{.underline}**

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.

![](media/image6.wmf){width="5.433333333333334in" height="6.0in"}

## **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 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.]{.underline}**

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.

![](media/image7.jpeg){width="6.0in" height="3.6in"}

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

![](media/image8.jpeg){width="6.0in" height="4.6194444444444445in"}

**Figure 8 -- Low-charge multiplex system compressor rack concept 2**

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.]{.underline}**

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 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]{.underline}**

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 air-cooled 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 pump
evaporator(s) and refrigeration system condenser(s) and the second
utilizes water-source 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.

![](media/image9.jpeg){width="5.6194444444444445in" height="3.75in"}

**Figure 9 -- Roof top heat pump unit**

![](media/image10.jpeg){width="5.629861111111111in"
height="3.9298611111111112in"}

**Figure 10 -- Plate heat exchangers (HX) for heat pump evaporators and
refrigeration desuperheaters and subcooler**

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 air-conditioner/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]{.underline}**

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 m^2^. Supermarkets are among the most energy intensive
buildings in the commercial sector with a total energy consumption of
more than 800 kWh-equivalent/m^2^/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 m^2^). The typical refrigeration
system contains approximately 750 kg of synthetic refrigerant equivalent
to 2,000 tonnes of CO~2~ in terms of greenhouse gas 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
non-technical 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 CO~2~/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 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 CO~2~ 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*)]{.underline}**

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 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 refrigerant-to-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 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.

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-   The total annual specific energy consumption (kWh/y/m^2^) 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 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 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
m^2^). 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 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.

Besides the phase out in 2006 for new plants HFCs are charged a CO~2~
tax. In Denmark emission of carbon dioxide (CO~2~) has been taxed since
May 15, 1992. Starting January 1, 2001 following the Kyoto protocol
Denmark extended the CO~2~ tax to include industrial greenhouse gasses
(HFCs, PFCs and SF~6~) based on the global warming potential (GWP) of
each gas. The tax is US\$ 0.0127 (DKK 0.10) pr. kg CO~2~ with CO~2~ 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 CO~2~). 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 CO~2~ was
used at the low temperature level (-32/-11°C). Direct expansion of CO~2~
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/
CO~2~ 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/ CO~2~ 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 CO~2~ 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 CO~2~.
Liquid CO~2~ is pumped directly to the fresh food cases while CO~2~ 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.

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]{.underline}**

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 CO~2~ as a
secondary refrigerant. This technology is implemented in more than 30
supermarkets throughout Sweden. CO~2~ 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.

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.

![](media/image11.wmf){width="5.011805555555555in"
height="3.86875in"}**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.

![](media/image12.wmf){width="5.89375in" height="3.61875in"}**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.

![](media/image13.wmf){width="5.881944444444445in"
height="3.660416666666667in"}**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.

![](media/image14.wmf){width="6.094444444444444in"
height="3.7423611111111112in"} **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 5^th^ 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.

**[United Kingdom]{.underline}**

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.

![](media/image15.png){width="5.06875in" height="2.792361111111111in"}

######## 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 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
CO~2~ 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 CO~2~ 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;
> CO~2~; 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, off-cycle 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 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.

*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).

![](media/image16.jpeg){width="6.0in" height="4.5in"}

**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.

For the chilled cabinet the energy consumption was more dependent on the
store conditions. A change of store temperature from 19 to 22^o^C
increased the energy use by up to 20%. At the higher temperature of
22^o^C 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, 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 m^2^ 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 m^2^ 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
low-charge 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.

+--------------+---------------+-----------+-------------+------------+
| # Tab        |               |           |             |            |
| le 1 - Predi |               |           |             |            |
| cted Energy  |               |           |             |            |
| Consumption  |               |           |             |            |
| for Low-Char |               |           |             |            |
| ge Refrigera |               |           |             |            |
| tion Systems |               |           |             |            |
+--------------+---------------+-----------+-------------+------------+
| System       | Heat          | Annual    | Energy      | \% Savings |
|              | Rejection     |           | Savings     | vs         |
|              |               | Energy    |             | baseline   |
|              |               |           | (kWh)       |            |
|              |               | (kWh)     |             |            |
+--------------+---------------+-----------+-------------+------------+
| Multiplex    | Air-Cooled    | 976,800   | \-          | \-         |
| (baseline)   |               |           |             |            |
|              | Condenser     |           |             |            |
+--------------+---------------+-----------+-------------+------------+
| Multiplex    | Evap.         | 896,400   | 80,400      | 8.2        |
|              | Condenser     |           |             |            |
+--------------+---------------+-----------+-------------+------------+
| Low-Charge   | Evap.         | 863,600   | 113,100     | 11.6       |
|              | Condenser     |           |             |            |
| Multiplex    |               |           |             |            |
+--------------+---------------+-----------+-------------+------------+
| Distributed  | Water-Cooled  | 866,100   | 110,700     | 11.3       |
|              | Condenser,    |           |             |            |
|              | Evap          |           |             |            |
|              | Rejection     |           |             |            |
+--------------+---------------+-----------+-------------+------------+
| Secondary    | Evap.         | 875,200   | 101,600     | 10.4       |
| Loop         | Condenser     |           |             |            |
+--------------+---------------+-----------+-------------+------------+
| Advanced     | Water-Cooled  | 1,048,300 | \-          | \-         |
| Se           | Condenser,    |           |             |            |
| lf-Contained | Evap          |           |             |            |
|              | Rejection     |           |             |            |
+--------------+---------------+-----------+-------------+------------+
| Secondary    | Water-Cooled  | 959,700   | 17,100      | 1.8        |
| Loop         | Condenser,    |           |             |            |
|              | Evap          |           |             |            |
|              | Rejection     |           |             |            |
+--------------+---------------+-----------+-------------+------------+
| 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 CO~2~ emission reductions compared to the
baseline multiplex system (multiplex with air-cooled 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.

+-----------+----------+-----+--------+---+------+-----+------+-----+
| **Table   |          |     |        |   |      |     |      |     |
| 2 - Total |          |     |        |   |      |     |      |     |
| E         |          |     |        |   |      |     |      |     |
| quivalent |          |     |        |   |      |     |      |     |
| Warming   |          |     |        |   |      |     |      |     |
| Impact    |          |     |        |   |      |     |      |     |
| (TEWI)    |          |     |        |   |      |     |      |     |
| for       |          |     |        |   |      |     |      |     |
| Su        |          |     |        |   |      |     |      |     |
| permarket |          |     |        |   |      |     |      |     |
| Refrig    |          |     |        |   |      |     |      |     |
| eration** |          |     |        |   |      |     |      |     |
+-----------+----------+-----+--------+---+------+-----+------+-----+
| System    | Co       | C   | P      | L | An   | T   |      |     |
|           | ndensing | har | rimary | e | nual | EWI |      |     |
|           |          | ge, |        | a |      | (m  |      |     |
|           |          | kg  | Refri  | k | kWh  | ill |      |     |
|           |          | /kW | gerant | ( |      | ion |      |     |
|           |          |     |        | % |      | kg  |      |     |
|           |          |     |        | ) |      | of  |      |     |
|           |          |     |        |   |      | CO~ |      |     |
|           |          |     |        |   |      | 2~) |      |     |
+-----------+----------+-----+--------+---+------+-----+------+-----+
|           |          |     |        |   |      | Dir | Ind  | To  |
|           |          |     |        |   |      | ect | irec | tal |
|           |          |     |        |   |      |     | t^a^ |     |
+-----------+----------+-----+--------+---+------+-----+------+-----+
| Multiplex | Ai       | 4   | R404A  | 3 | 976  | 13  | 9.52 | 23  |
|           | r-Cooled | .15 |        | 0 | ,800 | .62 |      | .14 |
|           | (b       |     | /R     |   |      |     | (9   |     |
|           | aseline) |     | -22^b^ | \ |      | (6. | .52) | (   |
|           |          |     |        | ( |      | 81) |      | 16. |
|           |          |     |        | 1 |      |     |      | 33) |
|           |          |     |        | 5 |      |     |      |     |
|           |          |     |        | \ |      |     |      |     |
|           |          |     |        | ) |      |     |      |     |
+-----------+----------+-----+--------+---+------+-----+------+-----+
|           | Eva      | 4   |        | 3 | 896  | 13  | 8.74 | 22  |
|           | porative | .15 |        | 0 | ,400 | .62 |      | .36 |
+-----------+----------+-----+--------+---+------+-----+------+-----+
| L         | Eva      | 2   | R404A  | 3 | 863  | 9   | 8.42 | 17  |
| ow-Charge | porative | .77 |        | 0 | ,600 | .08 |      | .50 |
|           |          |     | /R-22  |   |      |     |      |     |
| Multiplex |          |     | ^b^    |   |      |     |      |     |
+-----------+----------+-----+--------+---+------+-----+------+-----+
|           |          |     |        | 1 | 863  | 4   | 8.42 | 12  |
|           |          |     |        | 5 | ,600 | .54 |      | .96 |
+-----------+----------+-----+--------+---+------+-----+------+-----+
| Di        | Water    | 1   | R404A  | 5 | 866  | 1   | 8.44 | 9   |
| stributed | -Cooled, | .24 |        |   | ,100 | .00 |      | .44 |
| Co        | Evap.    |     |        |   |      |     |      |     |
| mpressors | Tower    |     |        |   |      |     |      |     |
+-----------+----------+-----+--------+---+------+-----+------+-----+
| Secondary | Eva      | 0   | R507   | 1 | 875  | 1   | 8.54 | 9   |
| Loop      | porative | .69 | ^c^    | 0 | ,200 | .13 |      | .67 |
+-----------+----------+-----+--------+---+------+-----+------+-----+
|           |          |     |        | 5 | 875  | 0   | 8.54 | 9   |
|           |          |     |        |   | ,200 | .56 |      | .10 |
+-----------+----------+-----+--------+---+------+-----+------+-----+
|           | Water    | 0   | R507   | 5 | 959  | 0   | 9.36 | 9   |
|           | -Cooled, | .27 | ^c^    |   | ,700 | .23 |      | .59 |
|           | Evap.    |     |        |   |      |     |      |     |
|           | Tower    |     |        |   |      |     |      |     |
+-----------+----------+-----+--------+---+------+-----+------+-----+
|           |          |     |        | 2 | 959  | 0   | 9.36 | 9   |
|           |          |     |        |   | ,700 | .09 |      | .45 |
+-----------+----------+-----+--------+---+------+-----+------+-----+
|           |          |     |        |   |      |     |      |     |
+-----------+----------+-----+--------+---+------+-----+------+-----+
| Results   |          |     |        |   |      |     |      |     |
| for       |          |     |        |   |      |     |      |     |
| Wa        |          |     |        |   |      |     |      |     |
| shington, |          |     |        |   |      |     |      |     |
| DC        |          |     |        |   |      |     |      |     |
| location  |          |     |        |   |      |     |      |     |
| -- 15     |          |     |        |   |      |     |      |     |
| year      |          |     |        |   |      |     |      |     |
| service   |          |     |        |   |      |     |      |     |
| life      |          |     |        |   |      |     |      |     |
+-----------+----------+-----+--------+---+------+-----+------+-----+

^a^Conversion factor = 0.65 kg CO~2~/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 air-conditioners/gas heaters. The cost premium
includes water piping for the heat pumps and 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-refrigerant-charge distributed refrigeration system was installed
that consisted of 10 compressor cabinets. Heat rejection for the
compressor cabinets was accomplished through water-cooled 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 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 m^3^ 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 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
low-charge refrigeration systems.

**[Canada]{.underline}**

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/m^2^/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
m^2^ 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]{.underline}**

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 CO~2~ 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 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]{.underline}**

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]{.underline}**

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 cycles are suited to relatively warm internal
temperatures and high relative humidities (22^o^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]{.underline}**

A 3720 m^2^ 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 CO~2~ 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]{.underline}**

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 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 SF~6~).

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.

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.

**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]{.underline}**

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   **Sale   **Low/Dual                               **Medium                                 
                Area**    Area**   Temperatures**                           Temperatures**                           

                \-        \-       \- 37 °C         \- 32   \- 29   -26 °C  -12 °C           -9 °C   -7 °C   -1 °C   7 °C
                                                    °C      °C                                                       

                ft^2^     ft^2^    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

"*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
  --------- ---------------- --------------------- -------------- --------- ---------

> **"*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       9 120          20         110
                             Service                                      

  *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        \-         \-
  --------- ---------------- ------------------ -------------- ---------- ----------

"*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;        8,4 A; 575 V/3 ph; 120     575 V/3 ph; 120 V/30 A
  Control            V/30 A                     
  ------------------ -------------------------- --------------------------

"*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  Both DT and MT Refrigeration
  (SDT) Temperatures (°C)                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; Space and Water Heating Systems
  Condenser; Geothermal                  

  Defrost Cycle Status and Operating     Fresh Meat Case and Cold Room;
  Time (sec/hr; hrs/day)                 Ice Cream
  -------------------------------------- --------------------------------

"*System2 -- phase1*": Characteristics of Dual (Low & Medium)
Temperature Compressors (THR - Total Heat Rejection; EER -- Energy
Efficiency Ratio)

  ---------------- --------- ------------------- --------------- ------------ ----------- ---------
  **Compressor**   **TSS**   **Model**           **Condensing\   **SUCTION\   **THR**     **EER**
                                                 temperature**   Capacity**               

  **-**            **°F**    **-**               **°F**          **Btu/h**    **Btu/h**   **-**

  COP/Discuss (22) -35       2DA3-060E-TFE-200   100             16 000       27 270      5,07
  L/T                                                                                     

  COP/Discuss (22) -25       3DB3-075E-TFE-200   100             37 000       55 490      6,76
  L/T                                                                                     

  COP/Discuss (22) -20       2DL3-040E-TFE-200   100             25 000       36 380      7,07
  L/T                                                                                     

  COP/Discuss (22) -20       3DS3-100E-TFE-200   100             56 000       82 320      7,1
  L/T                                                                                     

  COP/Discuss (22) -20       3DS3-100E-TFE-200   100             56 000       82 320      7,1
  L/T                                                                                     

  COP/Discuss (22) -20       4DA3-100E-TSE-200   100             59 000       88 510      6,93
  L/T                                                                                     

  COP/Discuss (22) 15        3DB3-100E-TFE-200   100             107 000      136 100     12,71
  M/T                                                                                     

  COP/Discuss (22) 15        3DS3-150E-TFE-200   100             142 000      181 100     12,32
  M/T                                                                                     

  COP/Discus (22)  15        4DA3-2000-TSE-200   100             150 000      190 800     12,54
  M/T                                                                                     

  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)

+---------------+---+--------------+---------+--------+--------+------+
| *             | * | **Model**    | **Con   | **SUC  | *      | **E  |
| *Compressor** | * |              | densing | TION** | *THR** | ER** |
|               | T |              | Temper  |        |        |      |
|               | S |              | ature** | **Capa |        |      |
|               | S |              |         | city** |        |      |
|               | * |              |         |        |        |      |
|               | * |              |         |        |        |      |
+---------------+---+--------------+---------+--------+--------+------+
| **-**         | * | **-**        | **°F**  | **B    | **B    | *    |
|               | * |              |         | tu/h** | tu/h** | *-** |
|               | ° |              |         |        |        |      |
|               | F |              |         |        |        |      |
|               | * |              |         |        |        |      |
|               | * |              |         |        |        |      |
+---------------+---+--------------+---------+--------+--------+------+
| COP/Discuss   | 2 | 2DA3-        | 100     | 83 000 | 103    | 1    |
| (22) M/T      | 0 | 075E-TFE-200 |         |        | 600    | 3,78 |
+---------------+---+--------------+---------+--------+--------+------+
| COP/Discuss   | 2 | 3DB3-        | 100     | 120    | 149    | 1    |
| (22) M/T      | 0 | 100E-TFE-200 |         | 000    | 300    | 3,82 |
+---------------+---+--------------+---------+--------+--------+------+
| COP/Discuss   | 2 | 3DL3-        | 100     | 159    | 199    | 1    |
| (22) M/T      | 0 | 150E-TFE-200 |         | 000    | 200    | 3,34 |
+---------------+---+--------------+---------+--------+--------+------+
| COP/Discuss   | 2 | 4DH3-        | 100     | 222    | 276    | 1    |
| (22) M/T      | 0 | 2500-TSE-200 |         | 000    | 800    | 3,69 |
+---------------+---+--------------+---------+--------+--------+------+
| Total         | 2 | \-           | \-      | 584    | \-     | \-   |
|               | 0 |              |         | 000    |        |      |
+---------------+---+--------------+---------+--------+--------+------+
|               | \ | \-           | \-      | \-     | 729    | \-   |
|               | - |              |         |        | 000    |      |
+---------------+---+--------------+---------+--------+--------+------+

"*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
  --------- ----------------------------------- -------------- ---------------

> "*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
  --------- ----------------------------------- -------------- ---------------

"*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    Each Unit
  Power (kW)                                 

  Suction Pressures (psig)                   Very Low, Low and Medium

  Common Discharge Pressures and Defrost     Each Rack
  Discharge (psig)                           

  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          Each Unit
  Exchangers Outlet (°C)                     

  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 Low and Dual Temperature
  (°C)                                         

  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;         Fresh Meat Case and Cold
  hrs/day)                                     Room; Ice Cream
  -------------------------------------------- --------------------------

> "System *CON*" : Display Cases and Walk‑in Boxes -- Low Temperature

  ------------- ----------------- -------------- ----------------- ---------------------------
  **Circuit**   **Description**   **Capacity**   **Operation       
                                                 Temperatures**    

  **-**         **-**                            **Evaporation**   **Condensing**

                                  **Btu/h**      **°F**            **°F**

  1             Different         25 500         \- 25             105
                Products                                           

  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        \-                \-
  ------------- ----------------- -------------- ----------------- ---------------------------

"*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                          
  ------------- ---------------------- -------------- ----------------- ----------------

"*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   Low and Medium Temperature
  (kWh)                                 Compressors

  Cooking Energy Consumption (kWh)      Each Electrical Oven

  Duration and Defrost Status (sec/hr;  Four Refrigeration Display Cases
  hrs/day)                              
  ------------------------------------- ---------------------------------

######## 

######## 

######## 

######## 

########  

###### Annual energy consumption

Total Store's Natural Gas Consumption vs. Sale Area

  ------------ ------------------- --------------- ------------------------- ------------------------ -------------------------- ----------- ----------------------
  **System**   **Sale Area**       **Total                                   **Equivalent Energy**                               **Total     
                                   Consumption**                                                                                 Natural Gas 
                                                                                                                                 Cost\       
                                                                                                                                 (CAN\$)**   

               $$\text{ft}^{2}$$   $m^{3}$/y       $m^{3}$/$\text{ft}^{2}$   GJ/y\                    GJ/$\text{ft}^{2}$/y\      \$/y        \$/$\text{ft}^{2}$/y
                                                                             (kWh/y)                  (kWh/$\text{ft}^{2}$/y)                

  Syst-1       28,302              45,179          1.59                      1.6\                     5.6\*$\text{10}^{- 5}$\    22,534      0.80
                                                                             (4.4\*$\text{10}^{6}$)   (155)                                  

  Syst-2       30,838              35,625          1.15                      1.27\                    4.12\*$\text{10}^{- 5}$\   17,860      0.58
                                                                             (3.5\*$\text{10}^{6}$)   (113)                                  
  ------------ ------------------- --------------- ------------------------- ------------------------ -------------------------- ----------- ----------------------

Total Store Annual Electrical Energy Consumption and Costs (Including
Demand Power Cost) vs. Sale Area (CAN\$)

  ------------ ------------------- -------------- ------------------------------ -------- ------------ ----------------------------- -------------
  **System**   **Sale Area**       **Total Store                                          **Total                                    
                                   Electrical                                             Store                                      
                                   Energy                                                 Electrical                                 
                                   Consumption\                                           Energy Cost\                               
                                   (including                                             (Power :                                   
                                   compressors                                            11,97                                      
                                   and heat                                               \$/kW)**                                   
                                   pumps)**                                                                                          

               $$\text{ft}^{2}$$   **kWh/y**      **kWh/**$\text{ft}^{2}$**.y\   **vs.    **\$/y\      **\$/**$\text{ft}^{2}$**.y\   **vs. CON.1\
                                                  **                             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

Annual Compressor Electrical Energy Consumption and Costs (Including
Power Demand Cost) -- vs. Sale Area (CAN\$)

  ------------ ------------------- ----------------- -------------- --------- ------------------------------- -------------- ---------- ------------------------------
  **System**   **Sale Area**       **Compressors**   **Total Energy                                           **Total                   
                                                     Consumption\                                             Electrical                
                                                     (Compressors                                             Energy Cost\              
                                                     only)**                                                  (Compressors              
                                                                                                              only)**                   

               $$\text{ft}^{2}$$   **-**             **kWh/y**      **% of    **kWh/**$\text{ft}^{2}$**.y**                  **\$/y**   **\$/**$\text{ft}^{2}$**.y**
                                                                    Total\                                                              
                                                                    Store                                                               
                                                                    Cons.**                                                             

  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.

**[Denmark]{.underline}**

**Fakta market in Beder, Denmark** (newest store in the Fakta chain as
of July 2001)

## Store details

> Size - 490 m^2^ sales area (720 m^2^ 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 CO~2~ 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 CO~2~ condenses in a cascade heat exchanger that
> also serves as one of two evaporators for the propane cycle. CO~2~ 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 CO~2~ circuit condenser); 10
> kW at -32°C evaporating temperature, CO~2~ circuit
>
> Primary refrigerant type and charge -- 10 kg propane; 6 kg CO~2~
>
> Secondary refrigerant type and charge -- 120 l propylene glycol
>
> Estimated annual losses -- propane circuit, 2% of charge; CO~2~
> 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
  -------------------------- --------------------------------------------

### CO~2~ system components

  ------------------------------- ---------------------------------------
  **Description**                 **Supplier**

  Compressor for CO~2~            Bitzer, X2KC-3.2, BSE 55

  Expansion valve                 AKV (nozzle 2-4)

  Safety pressure switches        KP 5

  Filter dryer                    DU 303
  ------------------------------- ---------------------------------------

## Display cases

+---------------------+--------+--------+---------+-------------------+
|                     | Air on | #####  |         | *Dimensions*      |
|                     |        | Refrig |         |                   |
|                     |        | eratio |         |                   |
|                     |        | n load |         |                   |
|                     |        |  \[W\] |         |                   |
+---------------------+--------+--------+---------+-------------------+
|                     | \[°C\] | Medium | Low     | LxBxH \[mm^3^\]   |
|                     |        |        |         |                   |
|                     |        | (+2°C) | (-20°C) |                   |
+---------------------+--------+--------+---------+-------------------+
| Storage (glass      | -2°C   | 4500   | \-      | 5100x3300x2360    |
| doors to shop)      |        |        |         |                   |
+---------------------+--------+--------+---------+-------------------+
| Storage, frozen     | -26°C  | \-     | 4200    | 3600x3300x2450    |
| food (doors)        |        |        |         |                   |
+---------------------+--------+--------+---------+-------------------+
| Storage, ice cream  | -28°C  |        | 2800    | 2100x3300x2450    |
| (doors)             |        |        |         |                   |
+---------------------+--------+--------+---------+-------------------+
| Vertical cabinet    | -2°C   | 14400  | \-      | 10 m in length    |
| (no doors)          |        |        |         |                   |
+---------------------+--------+--------+---------+-------------------+
| 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               |        | *      | *10000* |                   |
|                     |        | 21000* |         |                   |
+---------------------+--------+--------+---------+-------------------+

\* 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:                | 25°C (avg.)/ 22°C (min) / 30°C      |
|                                | (max)                               |
| Condensation Temperature:      |                                     |
|                                | -14°C (avg.)                        |
| Evaporation temperature:       |                                     |
|                                | 33 kW                               |
| Capacity (design):             |                                     |
+--------------------------------+-------------------------------------+
| CO2 system:                    | -11°C (avg.)                        |
|                                |                                     |
| Condensation Temperature:      | -32,5°C (avg.)                      |
|                                |                                     |
| Evaporation temperature:       | 10 kW                               |
|                                |                                     |
| Capacity (design):             |                                     |
+--------------------------------+-------------------------------------+
| Brine PG(40%):                 | -8°C                                |
|                                |                                     |
| Outlet temperature:            | -2°C                                |
|                                |                                     |
| Return temperature:            | 21 kW                               |
|                                |                                     |
| Capacity (design):             |                                     |
+--------------------------------+-------------------------------------+
|                                |                                     |
+--------------------------------+-------------------------------------+

## 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
  ------------------------------- ------ ----------- ----------- -----------

Refrigeration system energy consumption detail, July 2001

+---------------------------------------------+------------------------+
| Consumer                                    | Energy consumption     |
|                                             |                        |
|                                             | \[kWh for July\]       |
+---------------------------------------------+------------------------+
| Propane compressor 1                        | 3565                   |
+---------------------------------------------+------------------------+
| Propane compressor 2                        | 1240                   |
+---------------------------------------------+------------------------+
| CO~2~ 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.

![](media/image17.wmf){width="5.704861111111111in"
height="3.4159722222222224in"}

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.

**[United States]{.underline}**

# Test store -- Distributed compressor refrigeration system

> Size -- 4830 m^2^ 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.

<table>
<colgroup>
<col style="width: 13%" />
<col style="width: 9%" />
<col style="width: 2%" />
<col style="width: 7%" />
<col style="width: 0%" />
<col style="width: 20%" />
<col style="width: 6%" />
<col style="width: 9%" />
<col style="width: 2%" />
<col style="width: 7%" />
<col style="width: 21%" />
</colgroup>
<tbody>
<tr class="odd">
<td colspan="11"><p>Distributed Refrigeration System Compressors</p>
<p>(capacities are in kW for saturated evaporating and condensing
temperatures shown)</p>
<p>All compressor cabinets use refrigerant R-404A</p></td>
</tr>
<tr class="even">
<td colspan="4">Cabinet A Low Temperature</td>
<td colspan="2"></td>
<td colspan="4">Cabinet C Low Temperature</td>
<td></td>
</tr>
<tr class="odd">
<td colspan="2">Compressors</td>
<td colspan="4">Capacity: -32<sup>o</sup>C &amp; 41<sup>o</sup>C</td>
<td></td>
<td></td>
<td colspan="3">Capacity: -27<sup>o</sup>C &amp; 41<sup>o</sup>C</td>
</tr>
<tr class="even">
<td>A-1</td>
<td>ZF15K4</td>
<td colspan="2">4.8</td>
<td colspan="2"></td>
<td>C-1</td>
<td>ZF15K4</td>
<td colspan="2">6.3</td>
<td></td>
</tr>
<tr class="odd">
<td>A-2</td>
<td>ZF15K4</td>
<td colspan="2">4.8</td>
<td colspan="2"></td>
<td>C-2</td>
<td>ZF15K4</td>
<td colspan="2">6.3</td>
<td></td>
</tr>
<tr class="even">
<td>A-3</td>
<td>ZF15K4</td>
<td colspan="2">4.8</td>
<td colspan="2"></td>
<td>C-3</td>
<td>ZF15K4</td>
<td colspan="2">6.3</td>
<td></td>
</tr>
<tr class="odd">
<td>A-4</td>
<td>ZF18K4</td>
<td colspan="2">5.7</td>
<td colspan="2"></td>
<td>C-4</td>
<td>ZF15K4</td>
<td colspan="2">6.3</td>
<td></td>
</tr>
<tr class="even">
<td>A-5</td>
<td>ZF18K4</td>
<td colspan="2">5.7</td>
<td colspan="2"></td>
<td>C-5</td>
<td>ZF15K4</td>
<td colspan="2">6.3</td>
<td></td>
</tr>
<tr class="odd">
<td colspan="4">Cabinet B Low Temperature</td>
<td colspan="2"></td>
<td colspan="4">Cabinet D Low Temperature</td>
<td></td>
</tr>
<tr class="even">
<td></td>
<td></td>
<td colspan="4">Capacity: -32<sup>o</sup>C &amp; 41<sup>o</sup>C</td>
<td></td>
<td></td>
<td colspan="3">Capacity: -32<sup>o</sup>C &amp; 41<sup>o</sup>C</td>
</tr>
<tr class="odd">
<td>B-1</td>
<td>ZF13K4</td>
<td colspan="2">3.9</td>
<td colspan="2"></td>
<td>D-1</td>
<td>ZF15K4</td>
<td colspan="2">5.2</td>
<td></td>
</tr>
<tr class="even">
<td>B-2</td>
<td>ZF13K4</td>
<td colspan="2">3.9</td>
<td colspan="2"></td>
<td>D-2</td>
<td>ZF15K4</td>
<td colspan="2">5.2</td>
<td></td>
</tr>
<tr class="odd">
<td>B-3</td>
<td>ZF15K4</td>
<td colspan="2">4.8</td>
<td colspan="2"></td>
<td>D-3</td>
<td>ZF15K4</td>
<td colspan="2">5.2</td>
<td></td>
</tr>
<tr class="even">
<td></td>
<td></td>
<td colspan="4">Capacity: -26<sup>o</sup>C &amp; 41<sup>o</sup>C</td>
<td>D-4</td>
<td>ZF15K4</td>
<td colspan="2">5.2</td>
<td></td>
</tr>
<tr class="odd">
<td>B-4</td>
<td>ZF15K4</td>
<td colspan="2">6.1</td>
<td colspan="2"></td>
<td>D-5</td>
<td>ZF15K4</td>
<td colspan="2">5.2</td>
<td></td>
</tr>
<tr class="even">
<td>B-5</td>
<td>ZF18K4</td>
<td colspan="2">7.3</td>
<td colspan="2"></td>
<td>D-6</td>
<td>ZF18K4</td>
<td colspan="2">6.2</td>
<td></td>
</tr>
<tr class="odd">
<td>B-6</td>
<td>ZF18K4</td>
<td colspan="2">7.3</td>
<td colspan="2"></td>
<td></td>
<td></td>
<td colspan="2"></td>
<td></td>
</tr>
<tr class="even">
<td colspan="4">Cabinet E Medium Temperature</td>
<td colspan="2"></td>
<td colspan="4">Cabinet G Medium Temperature</td>
<td></td>
</tr>
<tr class="odd">
<td></td>
<td></td>
<td colspan="4">Capacity: -6.7<sup>o</sup>C &amp; 43<sup>o</sup>C</td>
<td></td>
<td></td>
<td colspan="3">Capacity: -8.3<sup>o</sup>C &amp; 43<sup>o</sup>C</td>
</tr>
<tr class="even">
<td>E-1</td>
<td>ZS38K4</td>
<td colspan="2">12.0</td>
<td colspan="2"></td>
<td>G-1</td>
<td>ZS38K4</td>
<td colspan="2">11.3</td>
<td></td>
</tr>
<tr class="odd">
<td>E-2</td>
<td>ZS38K4</td>
<td colspan="2">12.0</td>
<td colspan="2"></td>
<td>G-2</td>
<td>ZS38K4</td>
<td colspan="2">11.3</td>
<td></td>
</tr>
<tr class="even">
<td>E-3</td>
<td>ZS45K4</td>
<td colspan="2">14.4</td>
<td colspan="2"></td>
<td>G-3</td>
<td>ZS38K4</td>
<td colspan="2">11.3</td>
<td></td>
</tr>
<tr class="odd">
<td>E-4</td>
<td>ZS45K4</td>
<td colspan="2">14.4</td>
<td colspan="2"></td>
<td>G-4</td>
<td>ZS38K4</td>
<td colspan="2">11.3</td>
<td></td>
</tr>
<tr class="even">
<td>E-5</td>
<td>ZS45K4</td>
<td colspan="2">14.4</td>
<td colspan="2"></td>
<td>G-5</td>
<td>ZS38K4</td>
<td colspan="2">11.3</td>
<td></td>
</tr>
<tr class="odd">
<td>E-6</td>
<td>ZS45K4</td>
<td colspan="2">14.4</td>
<td colspan="2"></td>
<td>G-6</td>
<td>ZS45K4</td>
<td colspan="2">13.5</td>
<td></td>
</tr>
<tr class="even">
<td></td>
<td></td>
<td colspan="2"></td>
<td colspan="2"></td>
<td></td>
<td></td>
<td colspan="2"></td>
<td></td>
</tr>
<tr class="odd">
<td colspan="6">Cabinet F Low/Medium Temperature</td>
<td colspan="5">Cabinet H Low/Medium Temperature</td>
</tr>
<tr class="even">
<td></td>
<td></td>
<td colspan="4">Capacity: -6.7<sup>o</sup>C &amp; 43<sup>o</sup>C</td>
<td></td>
<td colspan="2"></td>
<td colspan="2">Capacity: -15<sup>o</sup>C &amp; 43<sup>o</sup>C</td>
</tr>
<tr class="odd">
<td>F-1</td>
<td>ZF09K4</td>
<td colspan="2"></td>
<td colspan="2"></td>
<td>H-1</td>
<td colspan="2">ZS26K4E</td>
<td>6.2</td>
<td></td>
</tr>
<tr class="even">
<td></td>
<td></td>
<td colspan="4">Capacity: -8.3<sup>o</sup>C &amp; 43<sup>o</sup>C</td>
<td>H-2</td>
<td colspan="2">ZS26K4E</td>
<td>6.2</td>
<td></td>
</tr>
<tr class="odd">
<td>F-2</td>
<td>ZS30K4</td>
<td colspan="3">9.2</td>
<td></td>
<td></td>
<td colspan="2"></td>
<td colspan="2">Capacity: 1.7<sup>o</sup>C &amp; 43<sup>o</sup>C</td>
</tr>
<tr class="even">
<td>F-3</td>
<td>ZS30K4</td>
<td colspan="3">9.2</td>
<td></td>
<td>H-3</td>
<td colspan="2">ZS38K4E</td>
<td>15.9</td>
<td></td>
</tr>
<tr class="odd">
<td>F-4</td>
<td>ZS30K4</td>
<td colspan="3">9.2</td>
<td></td>
<td>H-4</td>
<td colspan="2">ZS38K4E</td>
<td>15.9</td>
<td></td>
</tr>
<tr class="even">
<td>F-5</td>
<td>ZS38K4</td>
<td colspan="3">11.3</td>
<td></td>
<td>H-5</td>
<td colspan="2">ZS45K4E</td>
<td>19.1</td>
<td></td>
</tr>
<tr class="odd">
<td colspan="5">Cabinet I Medium Temperature</td>
<td></td>
<td colspan="5">Cabinet J Low/Medium Temperature</td>
</tr>
<tr class="even">
<td></td>
<td colspan="2"></td>
<td colspan="3">Capacity: -8.3<sup>o</sup>C &amp; 43<sup>o</sup>C</td>
<td></td>
<td colspan="2"></td>
<td colspan="2">Capacity: -6.7<sup>o</sup>C &amp; 43<sup>o</sup>C</td>
</tr>
<tr class="odd">
<td>I-1</td>
<td colspan="2">ZS38K4E</td>
<td>12.0</td>
<td colspan="2"></td>
<td>J-1</td>
<td colspan="2">ZF11K4E</td>
<td>3.6</td>
<td></td>
</tr>
<tr class="even">
<td>I-2</td>
<td colspan="2">ZS38K4E</td>
<td>12.0</td>
<td colspan="2"></td>
<td>J-2</td>
<td colspan="2">ZF13K4E</td>
<td>4.2</td>
<td></td>
</tr>
<tr class="odd">
<td>I-3</td>
<td colspan="2">ZS45K4E</td>
<td>14.4</td>
<td colspan="2"></td>
<td>J-3</td>
<td colspan="2">Zf13K4E</td>
<td>4.2</td>
<td></td>
</tr>
<tr class="even">
<td>I-4</td>
<td colspan="2">ZS45K4E</td>
<td>14.4</td>
<td colspan="2"></td>
<td></td>
<td colspan="2"></td>
<td></td>
<td></td>
</tr>
<tr class="odd">
<td>I-5</td>
<td colspan="2">ZS45K4E</td>
<td>14.4</td>
<td colspan="2"></td>
<td></td>
<td colspan="2"></td>
<td colspan="2">Capacity: -9.4<sup>o</sup>C &amp; 43<sup>o</sup>C</td>
</tr>
<tr class="even">
<td></td>
<td colspan="2"></td>
<td></td>
<td colspan="2"></td>
<td>J-4</td>
<td colspan="2">ZS38K4E</td>
<td>10.8</td>
<td></td>
</tr>
<tr class="odd">
<td></td>
<td colspan="2"></td>
<td></td>
<td colspan="2"></td>
<td>J-5</td>
<td colspan="2">ZS38K4E</td>
<td>10.8</td>
<td></td>
</tr>
<tr class="even">
<td></td>
<td colspan="2"></td>
<td></td>
<td colspan="2"></td>
<td>J-6</td>
<td colspan="2">ZS38K4E</td>
<td>10.8</td>
<td></td>
</tr>
</tbody>
</table>

> 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     Design Refrigeration
                          Doors, or Walk-in Floor Load (kW)
                          Area                    

  Cabinet A Sat. Suction                          
  Temp -32^o^C                                    

  Walk-in Freezer         94.1 m^2^               22.3

  Cabinet B Sat. Suction                          
  Temp -32^o^C                                    

  Walk-in Freezer         47 m^2^                 11.1

  Cabinet B Sat. Suction                          
  Temp -26^o^C                                    

  Reach-in Frozen Food    23 Doors                10.1

  Cabinet C Sat. Suction                          
  Temp -26^o^C                                    

  Reach-in Frozen Food    52 Doors                19.8

  Multi-Deck Frozen Food  2.4 m                   3.5

  Cabinet D Sat. Suction                          
  Temp -32^o^C                                    

  Multi-Deck Frozen Meat  17 m                    16.7

  Multi-Deck Frozen Fish  8.5 m                   7.0

  Walk-in Frozen Fish     13.4 m^2^               3.5

  Cabinet E Sat. Suction                          
  Temp -6.7^o^C                                   

  Multi-Deck Dairy        36.6 m                  47.1

  Multi-Deck Cheese       4.3 m                   5.5

  Walk-in Dairy           79.6 m^2^               18.2

  Cabinet F Sat. Suction                          
  Temp -8.3^o^C                                   

  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 m^2^               4.5

  Walk-in Raw             4.5 m^2^                1.4

  Cabinet F Sat. Suction                          
  Temp -29^o^C                                    

  Walk-in Deli Freezer    7.5 m^2^                1.8

  Cabinet G Sat. Suction                          
  Temp 8.3^o^C                                    

  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 m^2^               3.2

  Cabinet H Sat. Suction                          
  Temp 35^o^F                                     

  Meat Prep Room          6,144 ft^2^             153,600

  Cabinet H Sat. Suction                          
  Temp -15^o^C                                    

  Ice Maker                                       10.6

  Cabinet I Sat. Suction                          
  Temp -6.7^o^C                                   

  Multi-Deck Produce      35.4 m                  36.0

  Table Fish              4.6 m                   2.2

  Walk-in Produce         91.4 m^2^               19.5

  Cabinet J Sat. Suction                          
  Temp -9.4^o^C                                   

  Walk-in Meat            98.1 m^2^               23.2

  Walk-in Bakery          11.1 m^2^               5.8

  Cabinet J Sat. Suction                          
  Temp -29^o^C                                    

  Walk-in Freezer Meat    9.3 m^2^                2.2

  Walk-in Freezer Bakery  13.4 m^2^               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 m^2^ 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.

  --------------- --------------- ------------ ---------------------------------
  Multiplex                                    
  Refrigeration                                
  System                                       
  Compressors                                  

  Low Temp System                              
  Rack A                                       

  Main                                         
  Compressors                                  
  -29^o^C SST                                  
  R-404A                                       

  Comp1           4DL3-150E-TSK                

  Comp2           4DL3-150E-TSK                

  Comp3           4DL3-150E-TSK                

  Comp4           4DL3-150E-TSK                

                                               

  Satellite                                    
  Compressors                                  
  --34^o^C SST                                 
  R-404A                                       

  Comp5           3DF3-090E-TFD                

  Comp6           3DF3-090E-TFD                

  Comp7           2DL3-040E-TFD                

                                               

  Medium Temp                                  
  System Rack B                                

  Main                                         
  Compressors                                  
  --6.7^o^C SST                                
  R-22                                         

  Comp1           3DA3-0750-TFD                

  Comp2           3DA3-0750-TFD                

  Comp3           3DA3-0750-TFD                

  Comp4           2DD3-0500-TFD                

                                               

  Satellite                                    
  Compressors                                  
  --8.3^o^C SST                                
  R-22                                         

  Comp5           3DS3-1500-TFD                

  Comp6           3DF3-1200-TFD                

                                               

  Medium Temp                                  
  System Rack C                                

  Main                                         
  Compressors                                  
  --6.7^o^C SST                                
  R-22                                         

  Comp 1          3DS3-1500-TFD                

  Comp 2          3DA3-0750-TFD                

                                               

  Satellite                                    
  Compressors                                  
  1.7^o^C SST                                  
  R-22                                         

  Comp 3          3DS3-1500-TFD                

  Comp 4          3DS3-1500-TFD                

  Comp 5          3DF3-1200-TFD                

  Comp 6          2DL3-0750-TFD                

                                               
  --------------- --------------- ------------ ---------------------------------

## Display case line ups, storage and preparation room areas, and miscellaneous loads connected to compressor racks

  ----------------------- ----------------------- -----------------------
  Circuit                 Case Length, No. of     Design Refrigeration
                          Doors, or Walk-in Floor Load (kW)
                          Area                    

  Rack A Low                                      
  Temperature - Sat.                              
  Suction Temp -29^o^C                            
  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 m^2^               20.3

  Ice Maker                                       10.6

  Rack A Low                                      
  Temperature - Sat.                              
  Suction Temp -34^o^C                            

  Coffin Ice Cream        31.6 m                  21.9

  Coffin Shrimp           4.9 m                   1.7

  Walk-in Ice Cream       16.7 m^2^               5.0

  Rack B Medium                                   
  Temperature - Sat.                              
  Suction Temp -6.7^o^C                           
  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 m^2^                 3.1

  Floral Display          2.4 m                   3.0

  Rack B Medium                                   
  Temperature - Sat.                              
  Suction Temp -8.3^o^C                           

  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 -6.7^o^C                           
  Refrigerant - R22                               

  Walk-in Coolers         204.0 m^2^              57.9

  Rack C Medium                                   
  Temperature - Sat.                              
  Suction Temp 1.7 °C                             

  Meat Prep Room          93.8 m^2^               41.4

  Produce Prep Room       48.3 m^2^               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

## 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 (^o^C)          -26.6         -28.4        -27.3          -29.0

  SDT (^o^C)          32.2          28.2         16.1           21.7

  Return Temp (^o^C)  -11.4         -16.6        -8.7           -11.7

  Liquid Temp (^o^C)  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 (^o^C)          -8.8          -5.1         -10.3          -6.6

  SDT (^o^C)          33.6          28.5         22.3           19.4

  Return Temp (^o^C)  8.0           9.4          6.8            9.6

  Liquid Temp (^o^C)  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
  ------------------- ------------- ------------ -------------- ------------

TEWI Analysis for the Field Test Results -- May 2001 through February
2002

  ------------- ------------- -------- --------- --------- ---------- ------------
                Refrigerant            Energy    TEWI (kg             
                Leakage                          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 | Average      | Heat Pump   | Gas         |
|           | Recovered    | Space Heat   |             | Displaced   |
|           | (kW)         | Load (kW)    | Energy      |             |
|           |              |              | (kWh)       | (m^3^)      |
+-----------+--------------+--------------+-------------+-------------+
| November  | 72.7         | 95.6         | 17,236      | 8,076       |
| 01        |              |              |             |             |
+-----------+--------------+--------------+-------------+-------------+
|           |              |              |             |             |
+-----------+--------------+--------------+-------------+-------------+
| December  | 86.0         | 111.7        | 20,531      | 9,807       |
| 01        |              |              |             |             |
+-----------+--------------+--------------+-------------+-------------+
|           |              |              |             |             |
+-----------+--------------+--------------+-------------+-------------+
| January   | 88.5         | 115.7        | 21,004      | 10,080      |
| 02        |              |              |             |             |
+-----------+--------------+--------------+-------------+-------------+
|           |              |              |             |             |
+-----------+--------------+--------------+-------------+-------------+
| February  | 90.0         | 118.6        | 20,295      | 9,383       |
| 02        |              |              |             |             |
+-----------+--------------+--------------+-------------+-------------+
|           |              |              |             |             |
+-----------+--------------+--------------+-------------+-------------+
| 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.

# 

#