IIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 

Field Testing of an Advanced Low-Charge Supermarket Refrigeration 

System 

David H. Walker 

Foster-Miller, Inc. 

Waltham, MA 02451 

USA 

Van D. Baxter 

Oak Ridge National Laboratory 

Oak Ridge, TN 37831 

USA 

Abstract 

A field test was conducted to compare the performance of multiplex and distributed supermarket refrigeration 

systems. 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 3 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 watercooled condensers piped to a fluid loop that used dry fluid coolers. The second store also had water-source heat 

pumps for space heating and cooling that were piped into the fluid loops. Both sites were instrumented to determine 

energy consumption, refrigeration and heating, ventilating and air conditioning (HVAC) loads, and numerous 

system state points to characterize operation. Data were collected for approximately 18 months. A comparison of 

the performance of the two refrigeration systems was made based on the information gathered. The results showed 

that the multiplex refrigeration system used less energy and showed better EER than the distributed system. A 

TEWI comparison between the two refrigeration systems showed significant reductions in atmospheric CO2 

generation by the distributed system despite added energy use. Test results also showed that the water-source heat 

pumps produced energy cost savings for store space heating by recovery of the refrigeration reject heat. 

1. Introduction/Background 

Present supermarket systems employ direct expansion of refrigerant in the display cases and walk-in coolers 

with remotely operated multiplexed compressor racks and condensers. Because of the large floor area of modern 

supermarkets, the amount of refrigerant needed for operation is very large, on the order of 3,000 – 5,000 lb. With 

growing concern over the environmental impact of refrigerant leakage from these systems, several new system 

configurations have been designed and constructed that require less refrigerant for their operation. Examples of 

such systems include: 

- Distributed – Display cases and walk-in’s are piped to compressors located in cabinets on the sales floor or 

around the perimeter of the store. Heat rejection may be done through the use of water-cooled condensers and a 

fluid loop or with air-cooled condensers on the store roof above each compressor cabinet. 

- Secondary loop – Refrigeration to the display cases and walk-in’s is provided by a secondary fluid loop that is 

refrigerated by a central chiller system. Heat rejection is done through air-cooled or evaporative condensers, or 

a fluid loop can be employed to minimize the refrigerant charge. 

- Advanced self-contained – each display case or group of several cases has a condensing unit installed in the 

case. Heat rejection is done through the use of a fluid loop. 

While each of these systems uses significantly less refrigerant than standard multiplex compressor systems, the 

energy consumption of these systems is also very important in determining their impact on the environment. A 

complete assessment of the environmental impact of a refrigeration system can be made through a TEWI (Total 

Equivalent Warming Impact) evaluation that takes into account both primary and secondary global warming caused 

by direct refrigerant leakage and electric generation, respectively [1]. 

The energy consumption of low-charge refrigeration systems can be greater than that of multiplex systems 

because of the methods used to limit their refrigerant charges. For distributed refrigeration, low noise-level scroll 

compressors must be employed in order to locate the compressor cabinets in the sales area. Scroll compressors 

generally have lower EER values than reciprocating compressors today. If a fluid loop is used for heat rejection, 
IIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 

this will add an extra temperature difference that will increase the operating condensing temperature, and it will also 

result in adding pump energy consumption for fluid circulation in the loop. For the secondary loop system, the 

pumping energy associated with the secondary loops adds significantly to the total refrigeration energy. 

System design methods and enhancements exist that help offset the added energy associated with low-charge 

operation. Both distributed and secondary loop systems can be close-coupled to their evaporator loads which helps 

maintain higher suction pressure and tends to lower the return gas temperature. Scroll compressors can be operated 

at very low head pressures, because they have no suction or discharge valves. Refrigerant subcooling can be 

employed by scroll compressors through mid-scroll vapor injection. For the secondary loop system, refrigerant 

subcooling can be obtained through the use of hot brine defrost where the brine used for defrost is heated by 

subcooling the liquid refrigerant. The use of evaporative heat rejection helps to lower the condenser temperature 

penalty seen with fluid loops. 

An analytical investigation of low-charge refrigeration systems showed that with proper design and operation 

both distributed and secondary loop refrigeration systems could use less energy than a baseline air-cooled multiplex 

system. [2,3]. The estimated energy savings ranged from 6.3 - 12.4 % for distributed and 3.7 - 10.2 % for 

secondary loop refrigeration, respectively. A TEWI analysis of the distributed and secondary loop systems showed 

significant reductions (up to 60%) in CO2 generation versus the baseline air-cooled multiplex refrigeration system. 

Another energy saving method investigated was the use of water-source heat pumps in conjunction with fluid 

loops for heat rejection. In this approach, the heat of rejection from the refrigeration system can be reclaimed by the 

water-source heat pumps to provide space heating for the store. The advantage of this method over standard 

refrigeration heat reclaim is that the condensing temperature of the refrigeration system can be floated to minimum 

levels allowable without limiting heat recovery by the heat pump. A larger fraction of the reject heat can be utilized 

with the water-source heat pumps than can be recovered with conventional heat reclaim. 

In order to verify these analysis results, it was decided that a field test of a low-charge refrigeration system 

coupled with water-source heat pumps for HVAC should be undertaken. Through the cooperation and assistance of 

Price Chopper Supermarkets (Golub Corporation, Schenectady, NY) and Massachusetts Electric, such a field test 

was conducted involving two Price Chopper supermarkets in the Worcester, MA area. The test supermarket located 

in Marlborough, MA. was equipped with multiplex compressor racks with air-cooled condensers. The store 

consisted of 52,000 ft2

 of floor space and was a new construction site that opened in 1997. The second test 

supermarket was located in Webster, MA and also had approximately 52,000 ft2 of floor space. During renovation 

and remodeling of the Webster store, the distributed refrigeration system was installed along with 3 water-source 

heat pumps for store HVAC. Both stores were thoroughly instrumented and were monitored for approximately 2 

years. During that time period energy and performance data were gathered for the refrigeration and HVAC systems 

at both sites. 

2. Description of the Test Refrigeration Systems 

Table 1 provides a listing of the refrigerated fixtures that were connected to the multiplex refrigeration system. 

Three racks are employed with each rack having two suction groups. Rack 1 is the low temperature refrigeration 

system that employs R-404A as the refrigerant. Racks 2 and 3 are used for medium temperature refrigeration and 

both racks use R-22 as the refrigerant. The high temperature suction group of Rack 3 supplies mechanical 

subcooling for both Racks 1 and 2. Mechanical subcooling is normally used on low temperature refrigeration only, 

making the subcooling for the medium temperature Rack 2 somewhat unique. Separate air-cooled condensers were 

used for each compressor rack. The low temperature condenser was sized with a design temperature difference of 

10o

F while the medium temperature condensers were sized to operate with a 15oF temperature difference. During 

low ambient temperature operation, fan cycling was used to maintain condensing temperature at approximately 

70o

F. This same condenser temperature set point was employed with all 3 rack systems. The design refrigeration 

loads were 422,825 and 899,953 Btu/h for the low and medium temperature refrigeration, respectively. The design 

subcooling load for Racks 1 and 2 was 361,670 Btu/h. 

The details of the distributed refrigeration system are given in Table 2. The refrigeration load is divided among 

10 compressor cabinets (A – J). Cabinets A, B, C, and D are dedicated to low temperature refrigeration, while 

Cabinets E, G, and I are medium temperature refrigeration. Cabinets F, H, and J each have two suction groups, one 
IIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 

for low temperature refrigeration and the other for medium temperature. The low temperature cabinets A, B, C, and, 

D were equipped with subcooling by mid-scroll vapor injection. The design refrigeration loads for the distributed 

system were 404,845 and 1,010,936 Btu/h for low and medium temperature refrigeration, respectively. 

Heat rejection for the distributed refrigeration system was accomplished by fluid-cooled condensers located at 

each compressor cabinet. Two fluid loops using a propylene glycol/water solution were connected to these 

condensers and heat was rejected by fluid coolers located on the roof of the supermarket. The fluid coolers were dry 

units, rather than evaporative. The choice of dry coolers was made by Price Chopper, because the store is located a 

considerable distance from their headquarters and the added maintenance of evaporative units could not be 

addressed by their in-house maintenance people. Price Chopper felt that improperly maintained evaporative fluid 

coolers could result in refrigeration system reliability issues for the store. 

The use of dry fluid coolers was detrimental to the performance of the distributed system for several reasons. 

The added temperature difference associated with the operation of the fluid loops for heat rejection could not be 

offset by the dry coolers, since the coolers reject heat to the ambient dry-bulb temperature. This factor was 

particularly evident during summer operation when the condensing temperatures of the distributed cabinets were 

considerably higher than those of the multiplex racks. The fan power needed for the dry fluid coolers was also 

considerably larger than that associated with the multiplex air-cooled condensers, primarily because of the inclusion 

of the water-source heat pumps in the fluid loops. The fluid coolers had to be sized to provide heat rejection for the 

refrigeration and the water-source heat pumps when operating in space cooling mode. 

Water-source heat pump operation also impacted the sizing of the fluid loop pumps since added fluid flow was 

needed for space cooling heat rejection. Each fluid loop had a circulating pump. The power of each pump was 

measured at 6.5 kW. Pumping power for these loops was higher than expected because the loops were filled with a 

50/50 mixture of propylene glycol and water for freeze protection. The concentration of glycol was higher than 

needed for this particular location, but the concentration was chosen to insure reliable operation of the fluid loops 

during winter. 

Three water-source heat pumps were installed at the distributed system test store. The heat pumps were 

integrated into the fluid loops so that the refrigeration heat of rejection could be used by the heat pumps for store 

space heating and the fluid loops could be used for heat rejection from the heat pumps during store space cooling. 

Table 3 provides descriptions of the water-source heat pump units that were installed. The 2 large heat pumps were 

constructed to utilize a dual-path approach for space cooling operation. The heat pump was equipped with 2 sets of 

compressors and evaporator coils; one set of compressors and a coil were used for cooling return air from the store, 

while the other set of compressors and coil provided cooling and dehumidification to outside ventilation air. The 

cooling capacity of the heat pump was divided such that the outside air system provided approximately 25 tons of 

cooling while the return air system provided approximately 20 tons. In space heating mode, only the return air 

system was used. The third heat pump was considerably smaller than the 2 dual-path units and was operated as a 

single-path system where outside air was mixed with return air prior to passing through the heat pump coil. 

Operating data gathered from the heat pumps showed that they were very effective in recovering the rejected 

heat from the refrigeration system for store space heating. The dual-path approach for space cooling was found not 

to be appropriate for this particular site for space cooling. The dehumidification requirements for the site were much 

less than the dual-path system was designed for, and it was found that both the return and outside air systems had to 

be run in order to meet the total space cooling load of the supermarket. The outside air system required considerably 

more energy for cooling because of the higher temperature and humidity of the outside air. Energy consumption for 

space cooling was found to be very high for this particular site. A conventional single-path arrangement for all 3 

heat pumps would have been more appropriate. 

Space heating operation of the water-source heat pumps also affected the minimum operating temperature of the 

fluid loops, which, in turn, impacted the minimum condensing temperature that could be achieved by the distributed 

refrigeration system. Originally, it was decided that the minimum condensing temperature for the distributed system 

should be set at 50o

F, which meant that the fluid loops had to be operated at a minimum temperature of 40oF. 

During the first winter of operation it was found that the heat pumps required a fluid temperature of 50o

F in order to 

meet the largest space heating load at lowest ambient temperature. At the largest space heating load, the heat pumps 

needed both heat recovery from the fluid loop and auxiliary heating in order to meet this load. For this reason, the 
IIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 

fluid loops were maintained at a minimum temperature of 50oF during winter operation, which resulted in a 

minimum condensing temperature for the distributed refrigeration system on the order of 60oF. 

Table 3 - Operating Characteristics of the Water-Source Heat Pumps 

Heat Pumps WSHP-1 and WSHP-2 

 Space Cooling 

Entering Air Temp (oF ) Entering Water Temp(oF ) - 100 

 Dry- Bulb 73 Wet-Bulb 60 

 Cooling Capacity 

(Btu/h) 

Compressor Power 

(kW) 

Water Flow 

(gpm) 

Outside Air 293,173 26.4 68 

Return Air 252,202 19.4 66 

Space Heating 

Entering Air Temp (oF ) - 70 Entering Water Temp (oF ) - 70 

 Heat Absorbed 

(Btu/h) 

Compressor Power 

(kW) 

Water Flow 

(gpm) 

Return Air 278,672 17.2 66 

Heat Pump WSHP-3 

Space Cooling 

Entering Air Temp (oF ) Entering Water Temp(oF ) - 100 

 Dry- Bulb 73 Wet-Bulb 60 

 Cooling Capacity 

(Btu/h) 

Compressor Power 

(kW) 

Water Flow 

(gpm) 

 65,634 6.5 20 

Space Heating 

Entering Air Temp (oF ) - 70 Entering Water Temp (oF ) - 70 

 Heat Absorbed 

(Btu/h) 

Compressor Power 

(kW) 

Water Flow 

(gpm) 

 87,282 6.3 20 

Both sites were equipped with similar energy management systems (EMS) that were used for control of the 

refrigeration. The display cases and walk-in coolers at both sites employed case controllers for control of evaporator 

refrigerant flow and discharge air temperature. The case controllers employ 4 temperature readings in their control 

algorithms that are located at the refrigerant inlet and outlet of the coil and at the air discharge and return. 

Refrigerant flow was regulated at the evaporator by either an electronic expansion valve or by a combination of 

solenoid valve and electronic suction pressure regulator. 

2.1 Test Instrumentation and Data Collection

Both the multiplex and distributed refrigeration systems were thoroughly instrumented in order to assess and 

compare their operating performances. Table 4 lists the instrumentation associated with the multiplex refrigeration 

system. Each rack and each condenser was equipped with a watt transducer for power measurement; the watt 

transducer readings were used for all energy consumption comparisons. Data were collected at 5-minute intervals 

and stored in the EMS for later retrieval. 

The refrigeration load of each suction group was calculated based upon the suction and discharge pressures 

measured at the rack, the compressor on-off digital signals, and through the capacity curves for each compressor as 

specified by the manufacturer [reference Copeland Compressor data]. The procedure to determine the refrigeration 

load consisted of first finding the rated capacity of each compressor at the measured saturated suction and discharge 

temperatures (determined from the suction and discharge pressure readings). The rated capacities were then 

adjusted to take into account the actual return gas and liquid refrigerant temperatures. The refrigeration load for the 

rack suction group was then determined from 
IIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 

Qsg = ∑RFiCapi

where 

Qsg is the refrigeration load of the rack suction group (Btu/h) 

RFi is the on/off status of the compressor (on=1, off=0) 

Capi is the refrigeration capacity of the compressor for the measured operating conditions 

The load calculation was performed using the 5-minute data. The results were later averaged to determine the 

hourly and daily refrigeration loads. 

The refrigerant liquid subcooling loads for racks 1 and 2 were found by calculating a refrigerant mass flow rate from 

the load calculations and the enthalpy difference between the liquid and suction manifolds of each rack. The amount 

of subcooling could then be calculated using the liquid temperatures at the condenser outlet and the liquid manifold. 

The compressor power for subcooling was included in the power measurement of Rack 3. The power used for 

subcooling was estimated from the compressor performance curves based upon the subcooling load and the state 

points of the compressors. 

Table 4 – Multiplex Refrigeration Instrumentation (Each Compressor Rack) 

Rack Suction Pressure (one per suction group) 

Rack Discharge Pressure 

Return Gas Temperature (one per suction group) 

Liquid Temperature at Condenser Outlet 

Liquid Manifold Temperature (after subcooler) 

Rack Power (kW) 

Compressor On/Off Digital (one per compressor) 

Condenser Fan Power 

Table 5 describes the instrumentation installed on the distributed refrigeration system for the field test. Watt 

transducers were installed on all compressor cabinets and on the fluid coolers. Refrigeration state point 

measurements consisted of suction and discharge pressures and return gas temperature. For cabinets with two 

suction groups separate measurements were made for each suction pressure and return gas temperature 

Table 5 – Distributed Refrigeration Instrumentation (Each Compressor Cabinet) 

Cabinet Suction Pressure (one per suction group) 

Cabinet Discharge Pressure 

Return Gas Temperature (one per suction group) 

Liquid Manifold Temperature 

Cabinet Power (kW) 

Fluid Cooler Fan Power (one per fluid cooler) 

The large number of compressors made it too costly to install digital readings on each compressor. The 

refrigeration load for each cabinet was estimated based upon the rated capacity of compressors at the measured 

operating conditions. The rated power for each compressor was also estimated and the refrigeration load was 

adjusted by the ratio of the measured to the rated compressor power. The refrigeration loads of the compressor 

cabinets estimated using the instrument reading measured at 5-minute intervals. The refrigeration loads were 

calculated for each 5-minute interval and later averaging of these load estimates determined the hourly and daily 

refrigeration load values. 

Table 6 lists the instrumentation used for the evaluation of the water-source heat pumps. The operating mode of 

the heat pumps (heating or cooling) was determined by monitoring the heat pump power, the return air coil 

temperatures, and the water outlet temperature. Space heating or cooling was indicated by the coil temperature, with 

a temperature above 150o

F indicating heating and a temperature below 50oF indicating cooling. Compressor cycling 

could also be detected by a change in coil temperature accompanied by a change in total heat pump power. Heat 
IIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 

recovered from the fluid loop was determined by the change in fluid temperature and assuming a constant flow rate 

through the heat exchanger of 66 gpm. 

Table 6 – Instrumentation Associated with the Water-Source Heat Pumps 

Inlet Fluid Temperature to Water-Refrigerant HX 

Outlet Fluid Temperature leaving Heat Pump Water-Refrigerant HX 

Outlet Fluid Temperature leaving Water-Refrigerant HX for Outside Air Cooling 

Refrigerant-Air Coil Temperature - Return Air Compressor 1 

Refrigerant - Air Coil Temperature - Return Air Compressor 2 

Return Air Temperature 

Supply Air Temperature 

Heat Pump Power 

Outside ambient temperature and relative humidity readings were installed at both sites. After several months, 

it became apparent that the ambient temperature reading at the distributed refrigeration site was not reading 

correctly. The ambient temperature probe was located on the supermarket roof and the readings were strongly 

influenced by the sun. Also, the relative humidity readings at both sites stopped functioning after approximately 6 

months of operation. It was decided that a better approach for ambient data was to use hourly temperature and 

humidity readings that were obtained from NOAA for the Worcester Airport. Comparison of the airport temperature 

readings with readings from the multiplex refrigeration site showed good agreement and the airport relative 

humidity measurements were considered to be more representative. The NOAA ambient readings were used for 

analysis of refrigeration performance data for both sites. Both stores were equipped with inside dry-bulb 

temperature and relative humidity measurements that were used in conjunction with the EMS for HVAC control. 

All display cases and walk-in coolers at both sites were equipped with discharge air temperature sensors that 

were used by the EMS for refrigeration and fixture temperature control. Discharge air temperature readings for all 

cases and walk-in coolers were collected regularly as part of the field test data. 

All instrumentation at both sites was connected to the store EMS and readings were taken at 5-minute intervals. 

The EMS at both sites had adequate data storage capability so that they could be used as data acquisition systems for 

the field testing. 

2.2 Discussion of Field Test Results

Data collection was begun at both test stores in November 1999 and is continuing through the present. Test 

results for the first year of operation and comparison showed that the water-cooled condensers of the distributed 

system were under-sized, particularly for the low temperature compressor cabinets. The water-cooled condensers on 

all compressor cabinets were replaced with larger heat exchangers prior to May 2001. The results shown below are 

based upon data collected after the condenser replacement. 

Comparison between the two systems was done on the basis of low temperature (LT) and medium temperature 

(MT) refrigeration. For the multiplex system, the low temperature refrigeration data consisted of measurements 

taken from the low temperature compressor rack. For the distributed system, data for the low temperature 

refrigeration consisted of measurements taken on 4 compressor cabinets (Cabinets A-D) and on the 3 lowtemperature suction groups associated with Cabinets F, J, and H. The medium temperature data for the multiplex 

system were taken from the remaining 2 compressor racks. For the medium temperature refrigeration load of the 

multiplex system, the refrigeration provided for mechanical subcooling was estimated and removed, since this 

cooling was not used directly by the display cases or walk-in coolers. The compressor power associated with 

mechanical subcooling was included with the total power for the medium temperature refrigeration. The medium 

temperature refrigeration data for the distributed system consisted of measurements taken for the remaining 

compressor cabinets and the medium temperature suctions groups of the 3 split cabinets. The power and energy data 

for heat rejection of either system were combined for both medium and low temperature refrigeration, because each 

fluid loops of the distributed system service both the low and medium temperature compressor cabinets. 

Table 7 and Figure 1 show the energy consumption of the two refrigeration systems for two time periods, the 

first consisting of May through August 2001, and second from November 2001 through February 2002. Data for the 
IIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 

months of September and October 2001 were not included, because problems incurred with the EMS at both sites 

during that period did not allow adequate data to be collected for representative comparison. 

Table 7 - Energy Consumption Comparison between Multiplex and Distributed Refrigeration 

 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 

The energy data show that the multiplex compressors consumed less energy for both low and medium 

temperature refrigeration during summer operation. For the winter, the distributed system showed lower energy 

consumption for low temperature refrigeration, and higher energy consumption for medium temperature 

refrigeration. For heat rejection, the multiplex system had lower energy consumption during summer and higher 

during winter. This finding is significant in that the energy for heat rejection for the distributed system included 

both fan and pump energy, and the minimum rejection temperature for the fluid loops was 50o

F, versus 70oF for the 

multiplex condensers. It is likely that the reduction in heat rejection energy can be attributed to the operation of the 

water-source heat pumps for space heating. In general, the energy associated with heat rejection accounted for 15 - 

20% of the total refrigeration energy for both the multiplex and distributed refrigeration systems. 

0.0

500.0

1000.0

1500.0

2000.0

2500.0

3000.0

3500.0

May

June

July

August

Nov

Dec

Jan

Feb

Energy Consumption (kWh/day)

Dist -Comp

Mult - Comp

Dist - HR

Mult - HR

Figure 1 - Average daily energy consumption for the Multiplex and Distributed refrigeration systems 

Table 8 provides a description of the operating state points, refrigeration loads, and energy efficiency ratios 

(EER's) measured for each refrigeration system for summer and winter operation. The operating state points (i.e. 
IIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 

saturated suction and discharge temperatures, and return gas and liquid refrigerant temperatures) have been 

combined for each system in terms of low and medium temperature refrigeration. The average state point values 

were calculated on the basis of refrigeration load by 

∑ = ∑

i

i i

avg Q

SPQ SP

where 

 SPavg = the average state point value 

 SPi = the state point value for a particular suction group 

 Qi = the refrigeration load associated with the suction group 

The refrigeration load measurements indicated that the low temperature refrigeration loads are significantly 

different for the multiplex and distributed systems and the load for the distributed system is lower than that of the 

multiplex system. The medium temperature load for the distributed system is higher than that measured for the 

multiplex system. 

Normalization of the test results was done by comparing EER values. This comparison shows that the 

multiplex system had significantly higher EER values for low and medium temperature refrigeration for both 

summer and winter operation. For summer operation the multiplex system EER values were 34.7 and 18.5% higher 

than the EER values of the distributed system for low and medium temperature refrigeration, respectively. For 

winter operation the multiplex system EER values were higher than those of the distributed system by 12.1 and 

22.2% for low and medium temperature refrigeration, respectively. Figure 2 shows the EER values for both systems 

as a function of average daily dry-bulb temperature. The results in this plot show that the EER values for the 

multiplex system were consistently higher than those of the distributed system. 

Table 8 - Average State Points , Refrigeration Loads, and EER's for the Multiplex and Distributed Refrigeration 

Systems 

 May - Aug Nov - Feb 

Low Temp Distributed Multiplex Distributed Multiplex 

SST ( oF ) -15.8 -19.2 -17.1 -20.2 

SDT ( oF ) 90.0 82.8 61.0 71.0 

Return Temp ( oF ) 11.5 2.2 16.3 11.0 

Liquid Temp ( oF ) 79.1 57.8 56.9 50.4 

Ref Load (Btu/h) 340,800 455,600 309,000 385,900 

Comp Power (kW) 54.4 53.8 36.0 39.9 

EER (Btu/W-h) 6.34 8.54 8.63 9.67 

 

Med Temp Distributed Multiplex Distributed Multiplex 

SST ( oF ) 16.2 22.9 13.4 20.2 

SDT ( oF ) 92.5 83.3 72.1 67.0 

Return Temp ( oF ) 46.4 48.9 44.3 49.3 

Liquid Temp ( oF ) 85.0 68.1 67.1 58.3 

Ref Load (Btu/h) 729,900 653,200 541,000 442,800 

Comp Power (kW) 66.4 50.1 39.6 26.5 

EER (Btu/W-h) 11.16 13.22 13.71 16.75 
IIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 

0.00

2.00

4.00

6.00

8.00

10.00

12.00

14.00

16.00

18.00

20.00

0 20 40 60 80 100

Average Daily Dry-Bulb Temperature (F)

Compressor EER (Btu/W-h)

Dist LT Nov-Feb

Dist MT Nov-Feb

Dist LT May-Aug

Dist MT May-Aug

Mult LT Nov-Feb

Mult MT Nov-Feb

Mult LT May-Aug

Mult MT May-Aug

Figure 2 - EER Comparison between the Multiplex and Distributed Refrigeration 

A TEWI analysis of each test system was conducted and the results of this analysis are given in Table 9. The 

charge sizes for the two systems were estimated at 3,000 lb. of refrigerant for the multiplex system and 1,000 lb. for 

the distributed system. Actual refrigerant leakage rates for the systems are not known. The estimated leakage rates 

employed in the analysis are 20 and 5 % for the multiplex and distributed systems, respectively. The analysis 

showed that the distributed system had a lower TEWI for both the summer and winter periods despite the higher 

indirect value due to the higher energy consumption. 

Table 9 - TEWI Analysis for the Field Test Results 

 Refrigerant Leakage Energy TEWI (kg CO2) 

 (lb) (kWh) Direct Indirect Total 

 R-404A R-22 

May - August 

Multiplex 67 133 381,300 202,055 247,845 449,900 

Distributed 17 443,169 24,895 288,060 312,954 

Difference 136,046 

 

Nov - Feb 

Multiplex 67 133 234,960 202,055 152,724 354,779 

Distributed 17 255,600 24,895 164,034 188,929 

Difference 165,850 

 

Multiplex leak rate estimated at 20%/yr 

Distributed leak rate estimated at 5%/yr 

The field test results for the water-source heat pumps are given in Table 10 for the winter period from 

November 2001 through February 2002. The table lists the average recovery rate of heat from the fluid loops. The 

amount of heat recovered was estimated at 25.9% of the total heat rejection of the refrigeration system. This amount 
IIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 

is considerably less than the capability of the heat pumps, which can recover as much as 663,200 Btu/h, or 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. For much of the time, the 

large heat pumps had only one compressor operating. The small heat pump had a run fraction of approximately 

30% during this time period. 

Table 10 - Field Test Results for the Water-Source Heat Pumps 

Space Heating Performance (November 2001 - February 2002) 

 Average 

Recovered 

Heat (Btu/h) 

Average 

Space Heat 

(Btu/h) 

Heat Pump 

Energy 

(kWh) 

Gas Displaced 

(therms) 

November 01 248,400 326,100 17,236 2,966 

 

December 01 293,500 381,000 20,531 3,602 

January 02 301,900 394,700 21,004 3,702 

February 02 307,200 404,700 20,295 3,446 

Total 79,066 13,716 

Despite limited operation, the water-source heat pumps were able to displace approximately 13,716 therms of 

natural gas. The value of this displacement is dependent upon the utility rates for electric and gas. Figure 3 shows 

the estimated energy cost savings for a range of utility rates. For this particular site, the average commercial electric 

rate is approximately $.09/kWh and the gas rate is $.75/therm. The estimated energy cost savings seen over this 4-

month period were $3,171. 

-4,000

-2,000

0

2,000

4,000

6,000

8,000

10,000

12,000

0.05 0.06 0.07 0.08 0.09 0.1 0.11 0.12

Electric Rate ($/kWh)

Energy Cost Savings ($)

$0.50/therm

$0.75/therm

$1.00/therm

Figure 3 - Water-Source Heat Pump Energy Cost Savings (Nov '01 - Feb '02) 
IIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 

2.3 Conclusions and Recommendations

The distributed refrigeration system consumed much more energy and operated at much lower EER values 

relative to the baseline system than expected based on the projections from the earlier analytical studies [2,3]. This 

increased energy use can be attributed to several factors: 

♦ Use of dry fluid coolers for heat rejection - Lower energy consumption should be expected if evaporative heat 

reject was used for the fluid loops. This is particularly true for summer operation. With evaporative heat 

rejection the fluid loop temperature could be dropped below the ambient dry-bulb temperature and would help 

to lower the condensing temperatures seen at the compressor cabinets. Evaporative heat rejection would also 

require less fan power than dry heat rejection, which would help to lower energy consumption. It should be 

noted that evaporative heat rejection can be used with any supermarket refrigeration system but usually is not 

due primarily to the increased maintenance requirements. 

♦ The performance of the scroll compressors - Comparison of manufacturer’s data for the scroll and reciprocating 

compressors suggests that the EER of the scroll compressors is less than that of the reciprocating compressors at 

the same saturated suction and discharge temperatures. This observation is true for both low and medium 

temperature compressors. The methods needed to overcome this performance difference include operation of 

the scroll compressors at lower saturated discharge temperature and at higher saturated suction temperature. 

The differences in these values obtained by the distributed system tested were not adequate to overcome the 

performance difference between these two compressor types. It is also desirable to maintain as low a return gas 

temperature as possible to eliminate excess refrigeration load. The return gas temperature values were about the 

same for both the multiplex and distributed refrigeration systems despite the shorter piping runs used by the 

distributed system. 

♦ Extensive mechanical subcooling of the multiplex systems coupled with ineffective vapor injection subcooling 

for the distributed system compressors - The multiplex system used mechanical subcooling extensively for both 

the low and medium temperature compressor racks to maintain low liquid refrigerant temperature. This added 

subcooling produced more of a performance enhancement for the multiplex system than had been originally 

expected. Subcooling by mid-scroll vapor injection was installed on the scroll compressors in the low 

temperature compressor cabinets of the distributed system. The subcooling obtained was found to be very 

limited because the injection ports on the scroll compressors were sized for liquid injection and were too small 

to allow adequate vapor flow. 

Some of the performance difference between the two systems can also be attributed to the use of R-22 for 

medium temperature refrigeration on the multiplex system and R-404A on the distributed system. This difference 

will disappear in later installations after the phase-out of R-22. 

TEWI analysis for the two systems indicates that the distributed system has less environmental impact with 

lower total CO2 production than the multiplex system. This result was obtained even though the indirect CO2

production of the distributed system was greater due to higher energy consumption. 

The distributed system that was tested did not have compressor cabinets located in the sales area of the 

supermarket. The compressor cabinets were located either in backrooms around the perimeter of the store, or above 

several of the walk-in coolers. This fact suggests that other compressor types, such as reciprocating could have been 

used in the cabinets in this installation without concern about excessive noise in the sales area. Small rooftop 

refrigeration units may also be an effective alternative for a distributed refrigeration system. The use of more 

efficient reciprocating compressors could help to reduce the energy difference between the distributed and multiplex 

systems. 

Operation of the water-source heat pumps in conjunction with the distributed refrigeration system was shown to 

be effective at recovering reject heat from the refrigeration system for space heating. Energy cost savings were 

found for the test site due to this heat recovery by the heat pumps. It should be noted that this recover approach 

could be applied to any supermarket refrigeration system that employs a fluid loop for heat rejection and is not 

limited to the distributed system. 

Future project work calls for more analysis of the field test data to further quantify the differences between the 

multiplex and distributed systems. Field test data will be used to validate system models for both refrigeration 
IIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 

systems. The models will then be used to determine ways to change operation or design of the distributed system to 

improve its energy performance. 

3. References 

1. Sand, James R, Steven K. Fischer, Van D. Baxter, Energy and Global Warming Impacts of HFC Refrigerants 

and Emerging Technologies, Oak Ridge National Laboratory, sponsored by Alternative Fluorocarbons 

Environmental Acceptability Study (AFEAS), U.S. Department of Energy, 1997. 

2. Walker, D.H., “Low-charge refrigeration for supermarkets”, IEA Heat Pump Center Newsletter, Sittard, the 

Netherlands, Volume 18, No. 1/2000. 

3. Walker, David H., "Development and Demonstration of an Advanced Supermarket Refrigeration/HVAC 

System", Final Analysis Report, ORNL Subcontract Number 62X-SX363C, Foster-Miller, Inc., Waltham, MA 

02451, September, 2001. 

4. Acknowledgements 

The authors would like to acknowledge the support and direction provided by Dr. Esher Kweller of the US 

DOE. The US Department of Energy, Office of Building Technology, State and Community Programs sponsored 

this work under contract DE-AC05-00OR22725 with UT-Battelle, LLC. 
IIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 

Table 1 - Description of the Multiplex Refrigeration Test Store 

Circuit Case Length, No. of Doors, or 

Walk-in Floor Area 

Design Refrigeration Load (Btu/hr) 

Rack 1 Low Temperature - Sat. Suction Temp -20oF Refrigerant - R404A 

Reach-in Frozen Food 76 Doors 114,000 

Multi-Deck Frozen Meat 40 ft 60,520 

Multi-Deck Frozen Fish 18 ft 27,234 

Multi-Deck Bakery 12 ft 18,156 

Walk-in Freezers 925 ft2 69,375 

Ice Maker 36,000 

Rack 1 Low Temperature - Sat. Suction Temp -30oF 

Coffin Ice Cream 104 ft 74,600 

Coffin Shrimp 16 ft 5,840 

Walk-in Ice Cream 180 ft2 17,100 

Rack 2 Medium Temperature - Sat. Suction Temp 20oF Refrigerant - R22 

Reach-in Dairy 46 Doors 62,560 

Single-Deck Produce 56 ft 53,200 

Multi-Deck Produce 24 ft 32,160 

Service Meat & Deli 64 ft 23,692 

Multi-Deck Deli 48 ft 69,920 

Coffin Cheese & Deli 16 ft 24,800 

Tables Fish 20 ft 10,000 

Cooler Floral 140 ft2 10,500 

Floral Display 8 ft 10,336 

Rack 2 Medium Temperature - Sat. Suction Temp 15oF 

Bakery Display 15 ft. 12,550 

Multi-Deck Meat 68 ft 91,800 

Isle Cheese & Deli 56 ft 63,360 

Isle Ready Meals 52 ft 57,200 

Rack 3 Medium Temperature - Sat. Suction Temp 20oF Refrigerant - R22 

Walk-in Coolers 2,195 ft2 197,475 

Rack 3 Medium Temperature - Sat. Suction Temp 35oF 

Meat Prep Room 1,010 ft2 141,400 

Produce Prep Room 520 ft2 39,000 

 

Mechanical Subcooling 361,672 
IIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 

Table 2 - Description of the Distributed Refrigeration Test Store 

Circuit Case Length, No. of Doors, or 

Walk-in Floor Area 

Design Refrigeration Load (Btu/hr) 

Cabinet A Sat. Suction Temp -25oF 

Walk-in Freezer 1,013 ft2 76,000 

Cabinet B Sat. Suction Temp -25oF 

Walk-in Freezer 506 ft2 38,000 

Cabinet B Sat. Suction Temp -15oF 

Reach-in Frozen Food 23 Doors 34,500 

Cabinet C Sat. Suction Temp -15oF 

Reach-in Frozen Food 52 Doors 67,500 

Multi-Deck Frozen Food 8 ft 11,880 

Cabinet D Sat. Suction Temp -25oF 

Multi-Deck Frozen Meat 56 ft 57,120 

Multi-Deck Frozen Fish 28 ft 23,740 

Walk-in Frozen Fish 144 ft2 12,000 

Cabinet E Sat. Suction Temp 20oF 

Multi-Deck Dairy 120 ft 160,800 

Multi-Deck Cheese 14 ft 18,760 

Walk-in Dairy 857 ft2 62,000 

Cabinet F Sat. Suction Temp 17oF 

Service Deli 84 ft 25,920 

Multi-Deck Cheese 20 ft 28,800 

Isle Deli 16 ft 38,388 

Walk-in Deli 153 ft2 15,300 

Walk-in Raw 48 ft2 4,800 

Cabinet F Sat. Suction Temp -20oF 

Walk-in Deli Freezer 81 ft2 6,075 

Cabinet G Sat. Suction Temp 17oF 

Single-Deck Meat 24 ft 10,680 

Multi-Deck Meat 68 ft 91,800 

Floral Display 24 ft 31,008 

Multi-Deck Deli 28 ft 42,420 

Single-Deck Fish 16 ft 17,600 

Walk-in Fish 110 ft2 11,000 

Cabinet H Sat. Suction Temp 35oF 

Meat Prep Room 6,144 ft2 153,600 

Cabinet H Sat. Suction Temp 5oF 

Ice Maker 36,000 

Cabinet I Sat. Suction Temp 20oF 

Multi-Deck Produce 116 ft 122,680 

Table Fish 15 7,500 

Walk-in Produce 984 ft2 66,520 

Cabinet J Sat. Suction Temp 15oF 

Walk-in Meat 1,056 ft2 79,200 

Walk-in Bakery 120 ft2 19,800 

Cabinet J Sat. Suction Temp -20oF 

Walk-in Freezer Meat 100 ft2 7,650 

Walk-in Freezer Bakery 144 ft2 10,800 

Multi-Deck Bakery 14 ft 15,780 

All Compressor Cabinets employ R-404A as the refrigerant