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 compre ssor racks and air-cooled condensers. At the second store, a low-refrigerant-char ge 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, ven tilating 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 system s showed significant reductions in atmospheric CO 2 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 co ndensers. 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 cabin ets 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 refri gerant 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 seco ndary global warming caused by direct refrigerant leakage and el ectric 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 th eir refrigerant charges. For distributed refrigeration, low noise-level scroll compressors must be employed in or der 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 incr ease 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 as sociated 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 re turn gas temperature. Scroll compressors can be operated at very low head pressures, because they have no suc tion 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 CO 2 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 appr oach, the heat of rejection from the refrig eration 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 th e condensing temperature of the refrigera tion 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 pump s 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 co operation 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. Th e test supermarket located in Marlborough, MA. was equipped with multiplex compressor racks with air-cooled condensers. The store consisted of 52,000 ft 2 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 ft 2 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 ener gy 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 grou ps. Rack 1 is the low temperature refrigeration system that employs R-404A as the re frigerant. 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 conde nser was sized with a design temperature difference of

10 o F while the medium temperature condense rs were sized to operate with a 15 o F temperature diff erence. During low ambient temperature operation, fan cycling was used to maintain condensing te mperature at approximately

70 o 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 temper ature refrigeration, while Cabinets E, G, and I are medium temper ature refrigeration. Cabinets F, H, and J each have two suction groups, oneIIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 for low temperature refrigeration and the other for medium temp erature. 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 re frigeration system was accomplished by fl uid-cooled condensers located at each compressor cabinet. Two fluid loops using a propylene glycol/w ater 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 ch oice of dry coolers was made by Price Chopper, because the store is located a
considerable distance from their he adquarters and the added maintenance of evaporative units could not be addressed by their in-house maintenan ce people. Price Chopper felt that im properly 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 associat ed with the operation of the fluid loops for heat rejection could not be offset by the dry coolers, since the c oolers 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 larg er 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 hi gher than expected because the loops were filled with a 50/50 mixture of propylene glyc ol and water for freeze protection. The c oncentration of glycol was higher than needed for this particular location, bu t the concentration was chosen to insure reliable operation of the fluid loops during winter.

Three water-source heat pumps were installed at the di stributed system test stor e. 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 h eat 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 coo ling 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 pr ovided approximately 20 tons. In sp ace 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 re turn 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 sy stem for store space heating. The dual-path approach for space co oling was found not to be appropriate for this particular s ite for space cooling. The dehumidificati on 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 conven tional single-path arrangement for all 3 heat pumps would have been more appropriate.

Space heating operation of the water-source heat pumps al so affected the minimum ope rating 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 50 o F, which meant that the fluid loops had to be operated at a minimum temperature of 40 o F. During the first winter of operation it was found that the heat pumps required a fluid temperature of 50 o F in order to meet the largest space heating lo ad at lowest ambient temperature. At th e 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, theIIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 fluid loops were maintained at a minimum temperature of 50 o F during winter operation, which resulted in a minimum condensing temperature for the distributed refrigeration system on the order of 60 o F.

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

Heat Pumps WSHP-1 and WSHP-2 Space Cooling Entering Air Temp ( o F ) Entering Water Temp( o F ) - 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 ( o F ) - 70 Entering Water Temp ( o F ) - 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 ( o F ) Entering Water Temp( o F ) - 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 ( o F ) - 70 Entering Water Temp ( o F ) - 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 wa lk-in coolers at both sites employed case controllers for control of evaporator refrigerant flow and discharge air temper ature. 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 su ction 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 inst rumentation associated with the multiplex refrigeration system. Each rack and each condenser was equipped w ith a watt transducer for po wer measurement; the watt transducer readings were used for all energy consumption comparisons. Data were coll ected 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 discharg e pressure readings). The rated capacities were then adjusted to take into account the actual return gas and liq uid refrigerant temperatures. The refrigeration load for the rack suction group was then determined fromIIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 

# ∑

= i i sg 

### Cap RF Q

where

sg 

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

i 

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

i 

### Cap is the refrigeration capacity of the comp ressor 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 fo r 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 u pon the rated capacity of co mpressors 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 interv als. The refrigeration loads were
calculated for each 5-minute interval a nd later averaging of these load estimates dete rmined 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 temper ature. Space heating or cooling was in dicated by the coil temperature, with a temperature above 150 o F indicating heating and a temperature below 50 o F indicating cooling. Compressor cycling could also be detected by a change in coil temperature ac companied by a change in total heat pump power. HeatIIF - 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 pr obe 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 Wo rcester 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 repr esentative. The NOAA ambien t readings were used for analysis of refrigeration performance data for both site s. 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 temperat ure control. Discharge air temperature readings for all
cases and walk-in coolers were collected re gularly as part of the field test data.

All instrumentation at both sites was connected to the st ore 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 operatio n and comparison showed that the water-cooled condensers of the distributed system were under-sized, particularly for the low temperature compressor cabinet s. 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 af ter 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 te mperature 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 (Cabinet s 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 mechani cal subcooling was estimated and removed, since this cooling was not used directly by the display cases or walk-in coolers. The compre ssor 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 gr oups of the 3 split cabinets. The power and energy data for heat rejection of either system were combined for bo th 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 theIIF - IIR – Commission D1/B1 – Urbana, IL, USA – 2002/07 months of September and Oc tober 2001 were not included, because prob lems 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 th e distributed system included both fan and pump energy, and the minimum rejection temperature for the fluid loops was 50 o F, versus 70 o F 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 sp ace heating. In general, the energy associ ated with heat rejec tion 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

M a y J u n e J u l y

A

u g ust N ov D ec J a

n F ebEnergy 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. Th e 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 refrigeratio n. The average state point values were calculated on the basis of refrigeration load by



# ∑ ∑

=

i i i

avg 

### Q Q SP SP

where

SP avg = the average state point value

SP i = the state point value for a particular suction group Q i = the refrigeration load associ ated with the suction group

The refrigeration load measurements indicated that th e low temperature refrigera tion 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 compar ing 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 a nd medium temperature refriger ation, respectively. For winter operation the multiplex system EER values were hi gher 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 ( o F ) -15.8 -19.2 -17.1 -20.2
SDT ( o F ) 90.0 82.8 61.0 71.0
Return Temp ( o F ) 11.5 2.2 16.3 11.0
Liquid Temp ( o F ) 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 ( o F ) 16.2 22.9 13.4 20.2
SDT ( o F ) 92.5 83.3 72.1 67.0
Return Temp ( o F ) 46.4 48.9 44.3 49.3
Liquid Temp ( o F ) 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 16.75IIF - 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. Actu al refrigerant leakage rates for the systems are not known. The es timated 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 amountIIF - 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 ambi ent conditions during this time period were very mild so that the amount of space heat needed fo r the store was much less than normal ly 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 38 1,000 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-so urce heat pumps were able to disp lace approximately 13,716 therms of natural gas. The value of this displ acement is dependent upon th e 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/th erm. The estimated energy co st 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 pr ojections from the earlier analy tical studies [2,3]. This increased energy use can be attributed to several factors:

♦ Use of dry fluid coolers for heat rej ection - Lower energy consumption should be expected if evaporative heat reject was used for the fluid loops. This is particular ly true for summer operatio n. 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. Ev aporative 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 compressor s - Comparison of manufacturer’s da ta for the scroll and reciprocating compressors suggests that the EER of the scroll compressors is less than that of the r eciprocating 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 temp erature 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 ex cess refrigeration load. The return ga s 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 sy stem 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 syst ems 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 di stributed system has less environmental impact with
lower total CO 2 production than the multiplex system. This result was obtained even though the indirect CO 2 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 we re located either in backrooms around the perimeter of the store, or above several of the walk-in coolers. This fact suggests that ot her 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 ef fective alternative for a distributed re frigeration 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 refrigerationIIF - 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, spon sored 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 Nu mber 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 -20 o F 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 ft 2 69,375 Ice Maker 36,000 Rack 1 Low Temperature - Sat. Suction Temp -30 o F Coffin Ice Cream 104 ft 74,600 Coffin Shrimp 16 ft 5,840

Walk-in Ice Cream 180 ft 2 17,100 Rack 2 Medium Temperature - Sat. Suction Temp 20 o F 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 ft 2 10,500 Floral Display 8 ft 10,336 Rack 2 Medium Temperature - Sat. Suction Temp 15 o F 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 20 o F Refrigerant - R22 Walk-in Coolers 2,195 ft 2 197,475 Rack 3 Medium Temperature - Sat. Suction Temp 35 o F Meat Prep Room 1,010 ft 2 141,400 Produce Prep Room 520 ft 2 39,000

Mechanical Subcooling 361,672IIF - 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 -25 o F Walk-in Freezer 1,013 ft 2 76,000 Cabinet B Sat. Suction Temp -25 o F

Walk-in Freezer 506 ft 2 38,000 Cabinet B Sat. Suction Temp -15 o F Reach-in Frozen Food 23 Doors 34,500 Cabinet C Sat. Suction Temp -15 o F Reach-in Frozen Food 52 Doors 67,500 Multi-Deck Frozen Food 8 ft 11,880 Cabinet D Sat. Suction Temp -25 o F Multi-Deck Frozen Meat 56 ft 57,120 Multi-Deck Frozen Fish 28 ft 23,740

Walk-in Frozen Fish 144 ft 2 12,000 Cabinet E Sat. Suction Temp 20 o F Multi-Deck Dairy 120 ft 160,800 Multi-Deck Cheese 14 ft 18,760 Walk-in Dairy 857 ft 2 62,000 Cabinet F Sat. Suction Temp 17 o F Service Deli 84 ft 25,920 Multi-Deck Cheese 20 ft 28,800 Isle Deli 16 ft 38,388

Walk-in Deli 153 ft 2 15,300

Walk-in Raw 48 ft 2 4,800 Cabinet F Sat. Suction Temp -20 o F

Walk-in Deli Freezer 81 ft 2 6,075 Cabinet G Sat. Suction Temp 17 o F 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 ft 2 11,000 Cabinet H Sat. Suction Temp 35 o F Meat Prep Room 6,144 ft 2 153,600 Cabinet H Sat. Suction Temp 5 o F Ice Maker 36,000 Cabinet I Sat. Suction Temp 20 o F Multi-Deck Produce 116 ft 122,680 Table Fish 15 7,500

Walk-in Produce 984 ft 2 66,520 Cabinet J Sat. Suction Temp 15 o F Walk-in Meat 1,056 ft 2 79,200 Walk-in Bakery 120 ft 2 19,800 Cabinet J Sat. Suction Temp -20 o F

Walk-in Freezer Meat 100 ft 2 7,650 Walk-in Freezer Bakery 144 ft 2 10,800 Multi-Deck Bakery 14 ft 15,780 All Compressor Cabinets employ R-404A as the refrigerant