Investigation of Heat Recovery in Different Refrigeration System Solutions in

Supermarkets

JIGME NIDUP

Master of Science Thesis Stockholm, Sweden 2009

Investigation of Heat Recovery in Different Refrigeration System Solutions in

Supermarkets

Jigme Nidup

Master of Science Thesis Energy Technology 2009:499 KTH School of Industrial Engineering and Management Division of Applied Thermodynamics and Refrigeration

SE-100 44 STOCKHOLM

Master of Science Thesis EGI 2009/ETT:499

Investigation of Heat Recovery in Different Refrigeration System Solutions in Supermarkets

Jigme Nidup

Approved

Examiner

Björn Palm Supervisor

Samer Sawalha Commissioner

Contact person Samer Sawalha

Abstract This is a study of heat recovery in different refrigeration system in supermarkets. It presents analysis of field measurement in two supermarkets; CO2 trans-critical system with heat pump and CO2 trans-critical system with heat recovery only in de-superheater. The results from the field measurements showed that during the heat recovery the minimum condensing temperature was 20oC. Second part present computer models for heat recovery in three refrigeration systems solutions; the CO2 trans-critical system with heat pump, CO2 trans-critical with heat recovery in de-superheater only and a CO2/R404A cascade system. As the preliminary study in the project the methods and procedures have been developed for investigation of heat recoveries in three different refrigeration systems. The parameters that may be possibly used to compare different system have been proposed which need further refinement before drawing conclusion.

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Acknowledgements I would like thank the following members who formed the team working in the related project: Per-Olof NIlson at IUC, Katrineholm, Miceal Antonsson at Green and Cool, Magnus Rehnby at Bengt Dahlgren AB and Per-Erik Jansson at ICA Sverige AB for rendering all the support. I heartily thank Sarah Johansson and Jörgen Rogstam for allowing me to use the methods developed in their ongoing project in “Evaluation of CO2 supermarket refrigeration systems”; David Freléchox for all the support and suggestions. This project would not have seen the light without patient support of my supervisor Samer Sawalha. Thank you! Jigme Nidup Stockholm, 2009

Table of contents ACKNOWLEDGEMENTS........................................................................................................................ I TABLE OF CONTENTS ..........................................................................................................................II NOMENCLATURE................................................................................................................................. III INDEX OF FIGURES ............................................................................................................................. IV INDEX OF TABLES............................................................................................................................... VI 1 NTR I O N ........................................................................................................................... 11.1 ENERGY USE IN SUPERMARKETS AND TYPICAL HEATING REQUIREMENTS.................................. 11.2 OBJECTIVES............................................................................................................................................ 3

DUCTIO

2 METHODOLOGY........................................................................................................................... 3 3 HEAT RECOVERY SYSTEM SOLUT IONS................................................................................ 43.1 HEAT RECOVERY SYSTEM DESIGNS ...................................................................................................... 4 3.1.1 Heat Recovery System Design 1....................................................................................................5 3.1.2 Heat Recovery System Design 2....................................................................................................5 3.1.3 Heat Recovery System Design 3....................................................................................................6 3.1.4 Heat Recovery and Floating Condensing .................................................................................7 3.1.5 Heat Pump and Floating Condensing........................................................................................7

4  FIELD MEASUREMENT............................................................................................................... 94.1 CO2 TRANS‐CRITICAL SYSTEM 1 ........................................................................................................ 9 4.1.1 System design and operation.........................................................................................................9 4.1.2 Calculations of heat recovery .....................................................................................................10 4.1.3 Existing mode of operation and selection of operating conditions for  heat 

 with heat pump ............... recovery ...............................................................................................................114 Recoverable Heat with heat pump solution ........................................................................14

4.2 CO.1.4

2 TRANS‐CRITICAL SYSTEM 2 ......................................................................................................16 4.2.1 System design and operation......................................................................................................16 4.2.2 Calculation of Heat recovery ......................................................................................................17 4.2.3 Total heat recovery .........................................................................................................................19 4.2.4 Performance of Medium temperature Unit (KA3)............................................................21 4.2.5 Performance of booster unit (KAFA1)....................................................................................23

5 SIMULATION MODELS .............................................................................................................265.1 ASSUMPTIONS AND CORRELATIONS IN THE MODELS ...................................................................27 5.1.1 Cooling capacity as a function of outdoor temperature................................................27 5 .....................................................................................................285 Other refrigeration and heat exchanger parameters .....................................................29 EAT RECOVERY WITH HEAT PUMP ................

.1.2 Heating period..............................

.1.3   5.2 H ................................................................................305 Evaluation of different modes of operation EAT RECOVERY IN DE‐SUPERHEATER ONLY (TC2 .2.3 .........................................................................31

)..........................

5.3 H ..........................................405.4 HEAT RECOVERY IN R404A/CO2 CASCADE SYSTEM......................................................................465.5 COMPARISON OF DIFFERENT HEAT RECOVERY SYSTEM SOLUTIONS ..........................................50

6 DISCUSSION.................................................................................................................................52 7 CONCLUSION ...............................................................................................................................53 8 REFERENCES ...............................................................................................................................55

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Nomenclature CC Cascade system

CFC Chlorofluorocarbon

COP Coefficient of Performance (refrigeration)

E Compressor power consumption (electrical)

EES Engineering Equation Solver

FA Low temperature unit

HCFC Hydrochlorofluorcarbons

HFC Hydrofluorocarbons

HC Hydrocarbons

HVAC Heating, ventilation and air conditioning system

IHE Internal heat exchanger

KA medium temperature unit

PR Pressure ratio

TC1 Trans-critical system 1

TC2 Trans-critical system 2

III

Index of figures Figure 1.1: Typical energy use in a supermarket in Sweden (Arias J, 2005)...... 2 Figure 3.1: Heat recovery system design 1(Arias J, 2005)................................. 5 Figure 3.2: Heat recovery system design 2 (Arias J, 2005)................................ 6 Figure 3.3: Heat recovery system design 3 (Arias J, 2005)................................ 6 Figure 3.4: Heat recovery system design 3 (Arias J, 2005)................................ 7 Figure 3.5: Heat pump with floating condensing (Minea V, 2007) ...................... 8 Figure 4.1: Schematic diagram of trans-critical system 1 ................................... 9 Figure 4.2: Monthly average of cooling capacity, compressor power, and heat rejected based on the measurements of trans-critical system 1(TC1).............. 12 Figure 4.3: Existing mode of operation of medium temperature unit (KA1) and selection of operating condition for heat recovery in TC1................................. 13 Figure 4.4: Existing operating conditions of low temperature unit (FA1) and selection of operating conditions for heat recovery in TC1............................... 14 Figure 4.5: Estimated heat recovery potential with heat pump system............. 15 Figure 4.6: Schematic diagram of trans-critical system 2 (TC2) ....................... 16 Figure 4.7: schematic diagram of de-superheater showing the assumed parameter in the heat recovery calculation in TC2. .......................................... 19 Figure 4.8: Total cooling capacity, heat recovered and borehole capacity in TC2......................................................................................................................... 21 Figure 4.9: Average monthly performance of medium temperature unit (KA3) in TC2 .................................................................................................................. 22 Figure 4.10: Operation of medium temperature unit over a period of 24 hours in TC2 .................................................................................................................. 23 Figure 4.11: Average monthly performance of (booster system) (KAFA1) in TC2......................................................................................................................... 24 Figure 4.12: Operation of booster unit (KAFA1) over a period of 24 hours. ..... 25 Figure 5.1: The profile of cooling capacity as function of outdoor temperature 28 Figure 5.2: Schematic of heat recovery with heat pump system solution (TC1).......................................................................................................................... 30 Figure 5.3: COP in floating condensing and heat pump modes of operation for TC1. ................................................................................................................. 31 Figure 5.4: Operation in the heat pump mode with heat pump supply temperature of 45 o C compared to the floating condensing mode for TC1 ...... 32 Figure 5.5: Operation of refrigeration system in air cooled condenser and brine (secondary fluid) cooled condenser for TC1............................................ 33 Figure 5.7: System solution with heat pump supply temperature of 35oC in TC1.......................................................................................................................... 34 Figure 5.8: Operating at conditions to supply temperature of coolant from condenser at 45oC in TC1. ............................................................................... 35 Figure 5.9: Operating at conditions to supply temperature of coolant from condenser at 35oC in TC1. ............................................................................... 36 Figure 5.10: Comparison of air cooled mode and heat recovery mode with supply temperature of 35oC in TC1 .................................................................. 37 Figure 5.11: Ratio of heat recovery/rejection to cooling capacity in different modes of operation for TC1.............................................................................. 38 Figure 5.12: Total power consumption in different mode of operation for TC1. 39

IV

Figure 5.13: Schematic of refrigeration system with heat recovery in de-superheater region (TC2). ................................................................................ 40 Figure 5.14: Coefficient of performance of system in different modes of operation in TC2............................................................................................... 41 Figure 5.15: heat recovery at supply temperature to heat recovery system of 35oC in TC2...................................................................................................... 42 Figure 5.16: Heat recovery at supply temperature to heat recovery system of 45oC in TC2...................................................................................................... 43 Figure 5.17: Compressor power for different capacity and temperature of heat recovery in TC2 ................................................................................................ 44 Figure 5.18: Performance of TC2 with heat recovery of 70% of the total heat rejected at de-superheater region at supply temperature of 35oC .................... 45 Figure 5.19: Schematic of heat recovery system solution in R404A/CO2 cascade system (CC) ....................................................................................... 46 Figure 5.20: Heat recovery with supply temperature of 35oC in cascade system (CC).................................................................................................................. 47 Figure 5.21: Heat recovery with supply temperature of 45oC in cascade system (CC).................................................................................................................. 48 Figure 5.22: Comparative performance of heat recovery with 35oC and air-cooled condenser in floating condensing mode in cascade system (CC)......... 49 Figure 5.23: heat recovery ratio for different heat recovery temperature levels and floating condensing mode in CC................................................................ 49 Figure 5.24: Power consumption in different modes of operation in cascade system (CC) ..................................................................................................... 50 Figure 5.25: COP of TC1, TC2 and CC in floating condensing mode. ............. 51 Figure 5.26: Heat recovery ratio for different heat recovery system solutions for at supply temperature of 35oC..................................................................... 52

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Index of tables Table 1.1: Temperature requirements in heating applications (BBR, 2008; Wulfinghoff D R, 1999) ....................................................................................... 2 Table 4.1: Manufacturer’s data of compressors in TC1.................................... 10 Table 4.2: Monthly average values of heat recovery with the supply temperature of brine at 35oC in TC2..................................................................................... 20 Table 5.1: Average and maximum compressor power (Adapted from Jaime Arias, 2005) ...................................................................................................... 27 Table 5.2: Heating season and average monthly temperature. (Jonsson. H, et al, 2005, SMHI, 2009 and weather base, 2009) ............................................... 29 Table 5.3: Parameters assumed in the models ................................................ 29

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1 Introduction Mankind started with the use of natural ice to cool the food and then learned to produce the ice mechanically thus making cooling independent of nature. Later advances in the field of cooling made it possible to directly cool the food with out the use of ice as intermediate medium. Since these mechanical refrigerators involved the use of working fluid, development was governed by the choice of refrigerants. Until 1922 ammonia, carbon dioxide and sulfur dioxide were used as refrigerants. Though these refrigerants were naturally occurring safety concerns limited the widespread application. The introduction of new refrigerants in 1930 was the beginning of use of Chlorofluorocarbon (CFC) and hydro chlorofluorocarbon (HCFC) (Granryd E et al., 2003). Though these refrigerants were safe, CFCs were found to deplete the ozone layer and have high global warming potential. Then the concern shifted to the long term effect of the refrigeration to the environment. In Sweden, the use of CFCs was banned in 2000 and refilling of HCFCs was banned in 2002. Though HFC refrigerants accepted as the replacement of CFCs and HCFCs, owing to its very high global warming potential, it is accepted only as temporary solution. Regulations on synthetic refrigerants brought in many changes in refrigeration system. Systems became tighter, contained less volume of charge. In the process of redesigning the refrigeration system, different solutions have evolved in the supermarket refrigeration systems. With the focus on reduction of charge, the conventional direct expansion systems have been replaced with indirect systems. Supermarket refrigeration systems today may be partially indirect to completely indirect systems. Side by side the energy utilization in general and supermarkets in particular have been closely scrutinized. Many of the supermarket systems in Sweden use heat recovery system to recover heat from the refrigeration system to reduce the net energy purchased. Re-emerging use of natural refrigerants like propane, ammonia and carbon dioxide makes it interesting possibility to recover heat at higher temperature. Different system solutions have been implemented in different supermarkets in Sweden. There is a need for a comparative study of these system solutions.

1.1 Energy use in supermarkets and typical heating requirements In Sweden, supermarkets accounts for about 3% of the total electricity consumed in the country. It also revealed that reduction in energy consumption by 50% could increase the profit by 15%. (Arias J, 2005). Therefore detailed investigation of energy utilization in a supermarket is worthwhile. Figure 1-1 shows the energy use in a typical supermarket in Sweden, as can be seen in the figure, considerable portion of energy is consumed in refrigeration, lighting

and HVAC. While refrigeration system rejects considerable amount of heat, separate heating is required in the HVAC system. The recovery of heat from the refrigeration system presents potential to reduce or replace heating need for HVAC and service water.

Energy usage in a supermarket

Refrigeration, 47%

Lighting, 27%

HVAC, 13%

Kitchen, 3%

Outdoors, 5%

Others, 5%

Figure 1.1: Typical energy use in a supermarket in Sweden (Arias J, 2005)

Common heating applications in a supermarket consist of floor heating, heater for HVAC system, service water heater and in modern application, defrosting of evaporator coils. Some heating applications and temperature requirements are listed in table 1-1. Table 1.1: Temperature requirements in heating applications (BBR, 2008;

Wulfinghoff D R, 1999) Applications Temperature level

(oC) Floor heating 16-27 HVAC dimensioning temp 33 Service hot water 54 Hydronic system 71-49 Alternately with recovery system hydronic system can use temperature level

54-32

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1.2 Objectives Main objectives of the thesis are as follows:

• Modelling of different system solutions for heat recovery in supermarket refrigeration system.

• Collect and analyze field measurement data. • Evaluation of the performance of heat recovery systems in supermarkets

using the field measurements and the computer simulation models.

2 Methodology The literature review was carried out to study the different refrigeration and heat recovery system solutions in the literature. Refrigeration systems in two supermarkets with CO2 trans-critical systems have been analysed for heat recovery. The data from these supermarkets have been collected through the online interface IWMAC (IWMAC 2009). The electric power consumption of the compressor, pressure and temperature at key points are recorded for every 5 minutes interval. The important performance indicators are thereby calculated. Microsoft excel have been used for data analysis and NIST reference properties have been used to calculate the properties of refrigerants through the REFPROP 7.0 (NIST, 2002). Analyses of data from the field measurement have been used as one of the basis to make assumptions and develop the different computer simulation models. Some of the parameters have been either taken from literature or derived using the figures in the literature.

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3 Heat Recovery System Solutions Theoretically, in refrigeration system heat can be recovered from compressor oil, de-superheater and condenser. Potential of heat recovery depends on the quality and quantity of heat required and experience varies on the percentage of high temperature that can be recovered. In an industrial ammonia vapour compression system only 11.5% of the total system heat rejection is available in superheat region and the remaining is rejected at lower quality (Reindl T D, et al, 2007). A study of supermarket in the United States showed that heat rejection in superheat could range from 14-20% of the total heat available (Walker D H, 2001). Visile Minea suggested that up to 30% heat can be recovered at the superheat region (Minea V, 2007). Therefore the percentage of heat can be recovered depends on the refrigerants use the system solution.. Heat recovery system solutions in supermarket are used mainly to heat the space air. The practical experience indicated that though seemingly high quantity of heat is rejected by supermarket refrigeration systems, only 40-70% of the necessary heat can be recovered (Arias J, 2005). Arias suggested that refrigeration system not operating continuously to be the likely reason for less recovery of heat. In a typical Swedish supermarket HVAC system and refrigeration system are installed and operated by different companies. Therefore HVAC and refrigeration systems are isolated from each other with an heat exchanger has been cited as another reason. To get the required temperature of the water at the using end of the real estate of HVAC system, ideal temperature of coolant leaving condenser of the refrigeration system is 38oC (Arias J, 2005). This increases the condensing temperature, which increases the power consumption of the compressor. Other concerns are the effect on the cooling capacity, matching the timing of cooling and the heat recovery, matching temperature of heat recovery and that of application, loss in heat exchangers, and risk of leakage and contamination in the heat exchangers.

The fall in the outdoor temperature decreases the indoor relative humidity. This reduces the cooling load and therefore the heat rejected by the refrigeration system. The effect of outdoor temperature could reduce the refrigeration load by 50-70% of refrigeration load at standard indoor conditions of 22oC and 65% relative humidity (Arias J, 2005).

3.1 Heat recovery system designs Different heat recovery systems have been implemented in the supermarkets in Sweden. The following designs deal with the heat recovery at the condenser rather than de-superheater and reject the recovered heat to the HVAC directly or through a heat exchanger. The system solutions discussed are adapted from the doctoral thesis by Jaime Arias (Arias J, 2005).

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3.1.1 Heat Recovery System Design 1

Fig 3-1 shows the layout of heat recovery design 1. The coolant extracting heat from condenser of the refrigeration system rejects heat to the HVAC system through a heat exchanger before entering the dry cooler. The additional heat exchanger between the air system and refrigeration coolant loop reduces the temperature efficiency. The approach temperature of 32OC and return of 28 OC is obtained on the HVAC side of the heat exchanger. To maintain these temperature level, requirements of water temperature on the refrigeration (heat rejecting side) is around 36 OC approach to the dry cooler and 32 OC return.

To ensure operating at the required temperature of the HVAC heating system, an auxiliary heating system is used to heat the water. This requires complicated control system, when the heat exchanger in the air system has reached the required temperature; heat from the auxiliary system is bypassed and rejected through the dry cooler.

The introduction of additional heat exchanger between the HVAC system and chillers is to enable independent operation of chillers in events of break down of the HVAC system. In Sweden, the installation and operation of space heating system and refrigeration are often executed by different parties.

Figure 3.1: Heat recovery system design 1(Arias J, 2005)

3.1.2 Heat Recovery System Design 2

Figure 3.2 shows the arrangement of heat recovery design 2. In this design the auxiliary heating in air system is separated from the heat recovery heat exchanger. The drawback of heat loss from auxiliary heating to ambient is

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eliminated in this design, the reduction in approach temperature introduced by the additional heat exchanger is still present.

Figure 3.2: Heat recovery system design 2 (Arias J, 2005)

3.1.3 Heat Recovery System Design 3

Figure 3-3 shows the third design. This design eliminates the inadequacies in previous designs related to the heat recovery exchanger. Since the coolant from condenser of the refrigeration system directly rejects heat to the HVAC heat exchanger, the system is operated with approach temperature of 36oC and return temperature of 32oC. Auxiliary heating is placed after the heat exchanger in the HVAC system.

Figure 3.3: Heat recovery system design 3 (Arias J, 2005)

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3.1.4 Heat Recovery and Floating Condensing

Inherent disadvantage of heat recovery system is the need to refrigeration system at the high condensing pressure even in winter to maintain the required temperature of the heat utilizing system. Had the refrigeration system not been connected to heat recovery system, in cold days, condensing pressure could be lowered taking advantage of low ambient temperature. This could reduce the power consumption of the compressors in the refrigeration system. But since the temperature level of the rejected heat in this case is not of the level that can be used by the heat utilizing systems such as HVAC, the heat from the condenser will have to be rejected to the ambient. System in Figure 3.4 incorporates both the advantages. Such a system enables some chillers to run at high condensing temperatures to meet the heating demand and rest of the chillers can operate at lower condensing temperature to take advantage of low ambient temperature. In this system design one of the chillers is operated at high condensing temperature so as to meet the heating demand and the second chiller is operated at lower condensing pressure corresponding to the outdoor temperature.

Figure 3.4: Heat recovery system design 3 (Arias J, 2005)

3.1.5 Heat Pump and Floating Condensing

Another concept of heat recovery from refrigeration system is the use of heat pump to extract heat from the condenser coolant at low temperature and transfer it to the HVAC system at high temperature. This system enables the use of rejected heat at the same time allows the chillers to operate at floating condensing mode, therefore minimize the penalty on the compressor power. Figure 3.5 shows the schematic diagram of heat pump and floating condensing design. Experience with HCFC-22 at a supermarket in Canada showed that

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even with the floating condensing, temperature of brine entering the heat pump was about 25oC and COP of heat pump was as high as 4.6 (Minea V, 2006). One drawback of this design cited in the study was impossibility of operating the heat pump in cooling mode during the summer. On hot summer days the brines temperatures from the condenser of the chillers could rise up to 40oC which did not permit the operation of the heat pumps in reverse mode.

Figure 3.5: Heat pump with floating condensing (Minea V, 2007)

Another system design for supermarkets is to recover heat only in the de-superheater region. A field measurement of such a system is discussed in section 4.2.

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4 Field Measurement

4.1 CO2 trans-critical system 1 The measurement is carried out on a supermarket located in the far north of Sweden. The refrigeration system is CO2 trans-critical system and heat recovery is with a heat pump. The system solution is designated as trans-critical system 1 (TC1)

4.1.1 System design and operation

Oil cooler

Inter-CoolerOil cooler

+ -

Heat pump COP= 3.6

Q_max =300kW T_hp in= 13o C

- Pressure

-TemperatureKA1/KA2

FA1/FA2 Figure 4.1: Schematic diagram of trans-critical system 1

The refrigeration system is a parallel system with two circuits for low temperature level and two circuits for medium temperature level. The Low temperature units designated as FA1 and FA2 have two two-stage compressors each. Medium temperature units KA1 and KA2 have four single-stage compressors each. The manufacturer data of compressors are given in table 4.1.

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Table 4.1: Manufacturer’s data of compressors in TC1

Units model Swept Volume Low temperature TCDH 372B-D Medium temperature TCS373-D 12.6m3/h @ 2900rpm

All four circuits reject heat to the common coolant loop which is cooled by dry cooler. The compressor oil cooler also reject heat to the same coolant loop. A heat exchanger is connected to the coolant loop before the dry cooler to recover heat. This heat exchanger is connected to 300kW heat pump, which is meant to deliver heat to the HVAC system of the supermarket. At present the heat pump is not in operation, so all the heat is reject through the dry cooler.

4.1.2 Calculations of heat recovery

The supermarket is installed with measurements of temperature and pressure on the refrigerant circuits. Temperatures are also measured on the coolant loop. The measurement points are shown in figure 4.1. Power consumption of the compressors in the four units is measured separately. Power consumption, temperatures and pressures are measured for five minutes intervals and recorded by online monitoring system which can be accessed through web interface www.iwmac.se. The estimation of heat recovery is based on the parameters on the refrigeration side. The method developed in an ongoing project on evaluation of supermarket refrigeration systems have been adopted for calculation of various parameters (Johansson S, 2009). Using equation 4.1, mass flow of the refrigerant is calculated

[ skgv

Vmk

ssCO /*

2

.

2

η=& ] (4.1)

Where is swept volume and is taken from the manufacture’s data; is the specific volume and is obtained as the function of pressure and temperature of refrigerant at the inlet of the compressor. Volumetric efficiency (

.

sV kv2

sη ) is calculated as a function of suction and discharge pressures from the manufacture’s data. The function is different for the single and the two-stage compressor. The volumetric efficiency for the single stage compressor is obtained using equation 4.2.

[ ]%42.1025843.64079.0 arg2

arg +⎟⎟⎠

⎞⎜⎜⎝

⎛−⎟⎟

⎠

⎞⎜⎜⎝

⎛−=

suction

edisch

suction

edischs P

PP

Pη (4.2)

And the volumetric efficiency of two stage compressor is obtained using 4.3.

[ ]%424.931706.10251.0 arg2

arg +⎟⎟⎠

⎞⎜⎜⎝

⎛−⎟⎟

⎠

⎞⎜⎜⎝

⎛−=

suction

edisch

suction

edischs P

PP

Pη (4.3)

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The heat transfer across the evaporator, the condenser and the intercooler is calculated as the product of mass flow rate of refrigerant and the enthalpy difference across the heat exchanger using equation 4.4. (4.4) [kWhmQ CO ∆= *

2&& ]

Where, is the heat in (kW) and Q& h∆ is the difference in enthalpy across the heat exchanger in (kJ/kg). Heat loss in the oil cooler is calculated as the difference between the measured electrical power of the compressor and the calculated shaft power using equation 4.5. Heat loss from the compressor body to the ambient is assumed as 7% of the measured electrical power input. [ ]kWEEQ compshaftcompoilcooler )( &&& −= (4.5)

oilcoolerQ& is the heat loss in the oil cooler, and is the measured electrical power of the compressor in kW. is the net power transmitted to the compressor shaft and is calculated using the equation 4.6.

compE&

compshaftE&

(4.6) [kWhmE COcompshaft ∆= *

2&& ]

h∆ is difference of enthalpy between compressor outlet and inlet.

Total heat rejected from the refrigeration system is the sum of the condensers’, the inter coolers’ and the oil coolers’ capacities.

4.1.3 Existing mode of operation and selection of operating conditions for heat recovery with heat pump At present the heat recovery system is not in operation. Figure 4.2 shows the plot of total cooling capacity, average condensing temperature (T_condensing), common supply temperature of coolant brine (T_com,b,o) and total heat rejected in the system. The top line is the curve of total heat rejected by the system. This includes heat rejected from oil cooler and the inter cooler of the two-stage compressor of the low temperature units. Oil cooler capacity was found to be rather constant over the year at about 8kW and intercooler capacity was found to be about 13kW.

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0

50

100

150

200

250

Jan_

08

Feb_

08

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_08

April_0

8

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8

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_09

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Q,E

(KW

)

-10

-5

0

5

10

15

20

25

T (C

)

Cooling capacity Compressor power T_ambient T_com,b,o T_condensing Figure 4.2: Monthly average of cooling capacity, compressor power, and heat rejected based on the measurements of trans-critical system 1(TC1)

The common brine temperature of the heat rejection loop for the months of January and February in 2008 was not recorded. The brine temperatures are average for one month. In the existing mode of operation the brine temperature is above 18oC only in July and August. The design inlet temperature of brine to the heat pump is 13oC. This requires that the minimum supply temperature of brine from the refrigeration system to be 18oC, this is assuming 5K temperature difference across the heat exchanger. In order to operate the heat pump during the heating season, the refrigeration system would have to operate at discharge pressure corresponding to that of august (57.7 bar). This point of selection is highlight in figure 4.2. The selection criteria of operating conditions for individual units can be understood from figures 4.3 and 4.4.

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Q(k

W),T

(C)

0.0

0.5

1.0

1.5

2.0

2.5

3.0

3.5

4.0

4.5

5.0

CO

P

Cooling capacity Total heat rejected T_cond T_com,b,o COP Figure 4.3: Existing mode of operation of medium temperature unit (KA1)

and selection of operating condition for heat recovery in TC1 From figure 4.3, the operating conditions for medium temperature unit to run the system in heat recovery mode is selected as pointed out by the ellipse in the figure. Corresponding to common brine outlet temperature of 18oC, condensing pressure is selected at 58 bars and COP of 3.2 in the heat recovery mode.

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40

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08

Feb_0

8

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_08

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8

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8

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Q(k

W),T

(C)

0.0

0.2

0.4

0.6

0.8

1.0

1.2

1.4

1.6

1.8

2.0

CO

P

Cooling capacity Total heat rejected T_com,b,o T_cond COP Figure 4.4: Existing operating conditions of low temperature unit (FA1)

and selection of operating conditions for heat recovery in TC1

From the above figure, operating conditions for heat recovery mode is selected corresponding to common brine outlet temperature of 18oC. Therefore low temperature units need to be operated at condensing temperature of at least 20oC at COP of 1.7. These operating conditions have been assumed to calculate the heat that can be recovered if operated in heat recovery mode.

4.1.4 Recoverable Heat with heat pump solution

From the above selections, the refrigeration system is set to operate at condensing temperature of 20oC on the heat recovery mode. The COP of medium and low temperature units are set to 3.2 and 1.7 respectively during the heat recovery mode. Using these values, compressor power, oil cooler capacity, and the total heat rejected is calculated. Total heat rejected is used as the heat source for the heat pump. Using the design COP of the heat pump, which is obtained from the heat pump manufacturing data (CIAT, 2009) total heat that is provided by the heat pump, is calculated. This potential is compared with the maximum capacity of the heat pump, thus heat potential is limited to maximum capacity of the heat pump. Power consumption of the heat pump is estimated by dividing the total heat supplied by the heat pump by its COP. Total heat recovery as calculated is presented in figure 4.5. To compare the heating potential and the maximum capacity of the heat pump, heat recovery was considered even for the warm months of June, July and August. Presented in the plot from bottom to top of the curve are power consumption of the compressors of the refrigeration system while operating without any heat

14

recovery (E_ref_ref only mode), the power consumption of the compressors of refrigeration system when operating on heat recovery mode (E_ref_HR mode), total power consumption in heat recovery mode (E_tot_HR mode), Cooling capacity of the refrigeration system, COP of heating and heating capacity of the heat pump. The total power on heat pump mode (E_tot_HR mode) is sum of the power consumed by the compressors of the heat pump and the refrigeration system.

0

50

100

150

200

250

300

Jan_

08

Feb_0

8

March_

08

April_0

8

May_0

8

June

_08

July_

08

Aug_0

8

Sept_0

8

Oct_08

Nov_0

8

Dec_0

8

Jan_

09

Feb_0

9

March_

09

Month

Q,E

(kW

)

0

1

2

3

4

5

CO

P

E_ref_ref only mode E_ref_HR mode E_tot_HR modecooling capacity heat pump capacity COP heating

Maximum capacity of heat pump= 301kW

Figure 4.5: Estimated heat recovery potential with heat pump system.

The flat curve for heat pump capacity from April to October is because of reaching the maximum capacity of the heat pump. The upper limit of COP of heating is limited by COP of the heat pump. It can also be observed that COP of heating is lower during the heating season than on the warmer months. This can be attributed to the increase in power consumption of compressors in the refrigeration systems during the heating season. During the heating season, the condensing pressure is maintained at higher pressure to maintain the brine temperature leaving the condenser at a level necessary for operating for the heat pump.

15

4.2 CO2 trans-critical system 2 This supermarket is located near the city of Goteborg, which is at the western coast of Sweden.

4.2.1 System design and operation

Oil cooler

Ground heat source

+ +

Floor heating

-Pressure

-Temperature

Desuperheater Desuperheater

KAFA1/KAFA2

KA3

Figure 4.6: Schematic diagram of trans-critical system 2 (TC2)

This is a CO2 trans-critical system with two booster systems for the low and medium temperature levels (KAFA1&KAFA2) and a medium temperature circuit (KA3). Compressor in the medium temperature circuit, and high stage of Booster system, is TCS373-D model from Dorin with swept volume of 12.6m3/h at 2900rpm. The booster compressor in the booster system is SCS362-D model from Dorin with swept volume of 10.7m3/h at 2900rpm. On the heat rejection side, all three units are connected with a de-superheater before the air cooled condenser. The de-superheaters are connected to a common brine loop which is used to recover the heat. The recovered heat is used for floor heating and HVAC system of the building. The compressors’ oil in this system is cooled by a separate air cooler. The heat recovery system operates so as to maintain the supply temperature of the brine to the heating system at about 40oC. In the field measurement the average temperature were about 35oC. So the system fails to maintain the temperature at 40oC, instead we get 35oC in average. The capacity of individual unit is maintained by opening or closing the electronic valve connected to the

16

refrigerant line after the condenser. This raises or decreases the condensing pressure thereby controlling the capacity at fixed temperature. The supply temperature of the brine is also controlled by controlling the flow rate of brine to the de-superheater. Supply of brine to each de-superheater is controlled with the variable speed pump. The refrigerant line is also externally sub-cooled from the borehole before the supplying to the cabinets.

4.2.2 Calculation of Heat recovery

The heat recovery calculations are based on the refrigerant parameters and calculated using the same method developed in the project on evaluation of refrigeration systems in supermarkets (Johansson S, 2009). Using equation 4.1, the mass flow of the refrigerant in medium temperature level circuit is calculated. Volumetric efficiency ( sη ) is calculated as a function of suction and discharge temperature from the manufacture’s data. The volumetric efficiency for the medium temperature compressor and high stage compressor in the booster system is obtained using equation 4.2. The heat transfer across the evaporator and de-superheater is calculated as the product of mass flow rate of refrigerant and the enthalpy difference across the heat exchanger using equation 4.4. Heat loss in the oil cooler is calculated as difference of measured electrical power of compressor and the calculated shaft power of the compressor using equation 4.5. For the booster systems, the mass flow of refrigerant in the medium and low temperature level need to be evaluated separately to calculate the respective cooling capacities. [ ]skgmmm lowCOmediumCOtotalCO /

222&&& += (4. 7)

totalCOm

2& is the total mass flow of refrigerant; is the mass flow of

refrigerant in the medium temperature level, and is the mass flow in the low temperature level.

mediumCOm2

&

lowCOm2

&

The volumetric efficiency of the booster compressor is calculated using equation 4.8.

[ ]%12.951854.41139.0 arg2

arg +⎟⎟⎠

⎞⎜⎜⎝

⎛−⎟⎟

⎠

⎞⎜⎜⎝

⎛−=

low

edisch

suction

edischs P

PP

Pη (4.8)

Since both the low and medium temperature levels depend on the high stage compressor for cooling, it is necessary to divide up the compressor power in order to evaluate performance of the booster system. Total cooling capacity is divided as the following:

[kWEQQQ shaftboosterlowmediumhigh ,&&&& ++= ] (4.9)

17

mediumQ& is the cooling capacity of medium temperature; is the cooling load of

the low temperature level, and is the net power transferred to the refrigerant at the low stage and is calculated as

lowQ&

shaftboosterE ,&

(4.10) [kWhmE lowCOftboostersha ∆= *

2&& ]

h∆ is the enthalpy difference across the inlet and outlet of the booster

compressor. Then performances are calculated using following equations:

⎟⎟⎠

⎞⎜⎜⎝

⎛=

high

highhigh E

QCOP

&

& (4.11)

highCOP is the COP calculated across the high stage compressor . includes the cooling load from both low temperature and the medium temperature sides. Therefore compressor power ( ) has to be divided into low temperature side and that of the medium temperature side.

highQ&

highE&

[kWCOPQE

high

mediummediumhigh ⎟

⎟⎠

⎞⎜⎜⎝

⎛=

&&

, ] (4.12)

mediumhighE ,

& is the portion of power consumed in order to provide the cooling load of the medium temperature level.

boosterhigh

lowmediumboostertot EE

QQCOP&&

&&

++

=_ (4.13)

boostertotCOP _ , is total COP of the booster circuit.

boosterlowhigh

lowbooster EE

QCOP&&

&

+=

,

(4.14)

boosterCOP is calculated using the power consumed to provide cooling capacity at

the low temperature level of the booster system. is the power of booster compressor. is the portion of power of the high stage compressor which is consumed to provide cooling for the low temperature level. It is calculated using equation 4.15

boosterE&

lowhighE ,&

(4.15) [ ]kWEEE mediumhighhighlowhigh ,,

&&& −=

18

To calculate the heat recovery capacity on the refrigerant side the measurement of temperature and pressure before and after the de-superheater is necessary. While measurements of pressure and temperature for key points on the refrigerant line were available from September 2008, the measurement of temperature at the exit of de-superheater was available only from march, 2009. The temperature measurement of brine in and out of the de-superheater was available from October 2008. The average difference between the temperature of hot gas out of de-superheater and that of the temperature of brine at the inlet of the de-superheater for month of March is used to back calculate the hot gas temperature for the past months. The average temperature difference between hot gas and the brine inlet for the heat exchanger of the medium temperature unit KA3 was 4K, but for low temperature unit KAFA1 was 1K and for KAFA2 was 0K. Since the approach temperature of 1 and 0 is not reasonable, which could be accounted to measurement error, both the values have been discarded and approach temperature of 4K was used for all the de-superheaters. Therefore the temperature difference of 4K is added to the common brine inlet temperature to get the temperature of hot gas out at the de-superheater for the missing measurements.

De-superheater

Heat recovery brine loop

Refrigerant loop

Approach temperature (assumed) = T1-T2 =4K

T2 (measured)

T1 (Calculated)

Figure 4.7: schematic diagram of de-superheater showing the assumed

parameter in the heat recovery calculation in TC2.

4.2.3 Total heat recovery

Table 4.2 presents monthly average values of power consumption, cooling capacity, COP of refrigeration, total heat rejected, heat recovered and sub-cooling with borehole. The heat recovered in de-superheater is the heat recovered and the percentage of heat recovery is calculated as the ratio of heat recovered in de-superheater to that of the total heat rejected by the refrigeration system. The percentage of heat recovered through de-superheating varies from 24 % to 35 %.

19

The sub-cooling with the borehole is expressed as the percentage of cooling capacity to appreciate improvement due to the sub cooling with borehole. The total heat rejected in the system is equal to sum of cooling capacity and compressor power minus oil cooler capacity, borehole capacity and heat loss from the compressor body. The heat loss from compressor body is assumed as 7% of the compressor power. The heat balance for the calculations gives an error of about 4%. The error is quite acceptable and could be accounted to errors in measurement.

Table 4.2: Monthly average values of heat recovery with the supply

temperature of brine at 35oC in TC2

Description Oct-08 Nov-08 Dec-08 Jan-09 Feb-09 Mar-09 Apr-09

Average outdoor temperature (o C) 10 6 4 3 2 5 11

Compressor power (kW) 57 59 57 54 55 56 55

Cooling capacity(kW) 183 177 174 163 165 168 176Total COP 3.34 3.10 3.04 3.00 3.00 3.02 3.27Total heat rejected (kW) 211 201 179 166 164 170 187Q_desuperheater (heat recovered) (kW) 50 55 56 55 57 56 46Q_borehole (sub cooler)(kW) 9 12 29 30 33 32 24

Q_oilcooler (kW) 11 11 12 11 12 11 10% of Heat recovery = (Q_desuperheater/total heat rejected) 24% 27% 31% 33% 35% 33% 25%Q_borehole as percentage of cooling capacity 5% 7% 17% 19% 20% 19% 13%

The relationship between cooling capacity, heat recovery and condensing temperature can be observed in figure 4.8. Since heat recovery is only in the de-superheater, heat recovery is the de-superheater heat referred as Q_desuperheater.

20

0

10

20

30

40

50

60

Oct-08 Nov-08 Dec-08 Jan-09 Feb-09 Mar-09 Apr-09

Month

Q, (

kW),

T(C

)

1.0

1.5

2.0

2.5

3.0

3.5

CO

P

T_ambient Q_desuperheater Q_borehole T_cond COP

Figure 4.8: Total cooling capacity, heat recovered and borehole capacity

in TC2

The oil cooler capacity doesn’t change much over time. The evaporation temperatures are found to be rather constant over time for low temperature unit at -35oC and medium temperature at -10oC. It is seen that the heat recovery capacity is increased during colder months of December to March compared to warmer months of October, November and April. Therefore condensing temperature during the colder months is higher by 3oC and COP of the system drops to 3 compared to 3.3 during the warmer months. The capacity is increased by raising the condensing pressure, which is achieved by closing the electronic valve in the refrigerant line after the condenser. The temperature of the brine is maintained by regulating the flow rate of brine into the de-superheater. It is more convenient to understand these controls on individual units over shorter period and has been discussed in the following section.

4.2.4 Performance of Medium temperature Unit (KA3)

Figure 4.9 presents the monthly average of cooling capacity, COP, condensing temperature, heat recovered and the borehole capacity

21

0

10

20

30

40

50

60

70

80

Oct-08 Nov-08 Dec-08 Jan-09 Feb-09 Mar-09 Apr-09

Month

Q,E

(kW

),T(C

)

2.0

2.5

3.0

3.5

4.0

4.5

5.0

CO

P

T_cond Cooling capacity Q_desuperheater T_ambient Q_borehole(subcooling) COP

Figure 4.9: Average monthly performance of medium temperature unit (KA3) in TC2

The heat recovery capacity increases, while the cooling capacity decreases. The increase in heat recovery capacity and increase in condensing temperature are directly proportional. This reduces the COP of the refrigeration system. The borehole sub cooling capacity is low during the period from October to January, while condensing temperature has been increasing during these months. Sub cooling from the ground source reduces the compressor power consumed and therefore improves COP. The improvement in COP due to sub cooling can be observed in the months of February and March. Heat recovery and control of medium temperature unit for 24 hours is presented in figure 4.10. The top curve is the percentage opening of the electronic valve in the refrigerant line after the condenser. The middle curve is the discharge pressure. The bottom curve is the heat recovered in the de-superheater. It is seen that heat recovery starts about 7 AM and stops at around 10 PM. The discharge pressure is increased after 7 AM to meet the heating demand. The pressure is increased by partially closing the electronic valve. The valve is almost 100% during the night time when there is no heating demand.

22

0

10

20

30

40

50

60

70

80

90

100

00:00 02:25 04:30 06:55 08:40 10:20 12:10 13:55 16:00 17:50 20:35 23:00

Tiime (march 24,2009)

Q(k

W),

P(ba

r), C

apac

ity(%

)

Electronic valve opening 2 per. Mov. Avg. (Discharge Presure) 2 per. Mov. Avg. (Q_desuperheater)

Figure 4.10: Operation of medium temperature unit over a period of 24

hours in TC2

4.2.5 Performance of booster unit (KAFA1) The monthly average of cooling capacity, COP and heat recovery capacity are presented in figure 4.11

23

0

10

20

30

40

50

60

Oct-08 Nov-08 Dec-08 Jan-09 Feb-09 Mar-09 Apr-09

Month

Q,E

(kW

),T(C

)

1.0

1.5

2.0

2.5

3.0

3.5

CO

P

T_cond Q_desuperheater Q_borehole(subcooling) cooling capcacity COP Figure 4.11: Average monthly performance of (booster system) (KAFA1) in

TC2 It is seen that that the heat recovery capacity is limited in booster system. It is also necessary to maintain higher condensing temperature of about 23oC in the colder months to recover the same or less amount of heat to that of October and April. The cooling capacity drops quite sharply in colder month. The decrease in cooling capacity is due to the decrease in cooling demand in the medium temperature unit. The cooling capacity of low temperature unit is found to be rather constant. The control of the booster system over a period of day and heat recovery capacity is presented in figure 4.12. It can be seen that the discharge pressure is controlled within smaller range of variation than in the case of KA3. This could be accounted for the fact that the heat recovery capacity in the booster system is higher than in case of the medium temperature level system.

24

0

10

20

30

40

50

60

70

80

90

100

00:00 01:40 03:20 05:05 06:45 08:25 10:05 11:45 13:25 15:05 16:45 18:25 20:05 21:45 23:25

Time (March 24, 2009)

Q(k

W),P

(bar

), C

apac

ity (%

)

Electronic valve opening 2 per. Mov. Avg. (Discharge pressure) 2 per. Mov. Avg. (Q_desuperheater)

Figure 4.12: Operation of booster unit (KAFA1) over a period of 24 hours.

25

5 Simulation models Comparative study of three different heat recovery system solutions have been done using simulation model in engineering equation solver (EES). The system solutions are: • CO2 trans-critical system with heat recovery with heat pump. The model is

similar to the one in section 4.1 and is designated as TC1. • CO2 trans-critical system with heat recovery only in de-superheater. It is

similar to the experimental supermarket discussed in section 4.2, except that external sub-cooling is omitted in the model. It is designated as TC2

• CO2/R404A cascade system. R404A is used as refrigerant on the high

temperature side and CO2 is used as refrigerant in the low temperature side. The intermediate fluid (brine) is circulated between the evaporator of R404A and CO2 condenser. The medium temperature cabinets are also cooled with the brine. It is designated as CC.

The parameters of the heat exchangers used in the models have been calculated from the experimental values. However, in some cases averages has to be taken in order to incorporate fair comparison. An important term that needs to be explained is “floating condensing”. Though, the lowest condensing temperature is limited to 10oC, systems normally operating within this condition are referred to as floating condensing. COP of heating has been introduced to compare heat recovery in different systems and is expressed as;

floatingtoteryretot

heatingHeat EE

QCOP

,cov,&&

&

−= (5.1)

Where, is net heat supplied for heating, is the total power

consumption when operating in heat recovery mode and is the total power consumption as if the system is running in floating condensing and without heat recovery; thus, the system is operating at the lowest condensing pressure possible and consumes the lowest energy to provide the required cooling.

heatingQ& eryretotE cov,&

floatingtotE ,&

A term that has been introduced is the heat recovery ratio (HRR), which is the ratio of the heat recovered to the cooling capacity. The COP of the refrigeration system is the total system COP and not of individual units.

26

5.1 Assumptions and Correlations in the models

5.1.1 Cooling capacity as a function of outdoor temperature

The cooling capacity is mainly dependent on outdoor temperature and relative humidity. Since indoor relative humidity is attached to the outdoor temperature, the cooling capacity increases with increasing outdoor temperature. Jaime Arias has compiled considerable plots for compressor power in respect to outdoor temperature (Arias, 2005). It was observed from the plots in above literature that during the cold season between October and May the compressor power is more or less constant. During this period outdoor temperature is less than 10oC. Thus average power has been taken for temperature below 10oC. The maximum compressor power consumption is reached in the month of July; therefore maximum power and temperature have been read from the plots presented by Jaime Arias (Arias, 2005) and tabulated in table 5.1. Table 5.1: Average and maximum compressor power (Adapted from Jaime

Arias, 2005)

Supermarket

Average compressor Power (kW) for out door temperature below 10oC

Maximum outdoor temperature (oC)

Maximum compressor power (kW) on the days with maximum outdoor temperature

Sala 46 21 56Hjo 25 24 32Hedeamora 15 28 27Täby 60 28 100Kista 35 23 65Farsta 40 23 56Average 36 25 56

A correlation has been developed for the compressor power as a function of outdoor temperature in the range between 10oC and 25oC.

[kWTE outdoorcompressor 6.23*32.1 +=& ]

]

(5.2) Using the correlation 5.2, the compressor power at outdoor temperature of 35oC has been extrapolated and the value of 70kW was obtained. Average power below 10oC is approximately 50 % of the maximum power at maximum temperature. Therefore compressor power above 10oC is obtained in terms of percentage of the maximum power and outdoor temperature as follows;

[kWTE outdoorcompressor 30*02.0 +=& (5.3) Applying the same relation to the cooling capacity, the following correlation is used to calculate cooling capacity at different outdoor temperature:

][30*02.02 kWTQ outdoor +=& : (5.4) CT ooutdoor 10>

27

][*5.0 max2 kWQQ && = ; (5.5) CT o

outdoor 10≤ Where, is the cooling capacity in kW, 2Q& outdoorT is the outdoor temperature in oC. and maxQ& maximum/design cooling capacity in kW. The profile of cooling capacity obtained from the above relation is presented in figure 5.1.

0

25

50

75

100

-10 0 10 35

Out door temperature (C)

% c

oolin

g ca

paci

ty

Figure 5.1: The profile of cooling capacity as function of outdoor

temperature In the above figure the cooling capacity is plotted as a function of outdoor temperature. The cooling capacity is expressed in percentage of the maximum cooling capacity/design capacity. The full capacity, 100% at the vertical axis in the plot, is reached at outdoor temperature of 35oC. Below this temperature the cooling capacity is decreasing due to the decreasing humidity ratio. The equations 5.4 and 5.5 describes the correlation between outdoor temperature and the cooling capacity and are used in the simulations models to obtain the cooling capacity as function of outdoor temperature.

5.1.2 Heating period

To have reasonably fair comparison of different systems, it was necessary to define the heating period in terms of outdoor temperature. Monthly average temperatures on the start and end of heating season for some cities in Sweden have been compared in table 5.2. The average monthly temperatures for cities of interest have been taken from an online web accessible at www.weatherebase.com. The monthly data obtained from the weather base have been compared to past record from SMHI for four cities (SMHI, 2009). From the information of heating season and average temperature of these cities, it is seen that heating season begins when the average outdoor temperature is below 10oC.

28

Table 5.2: Heating season and average monthly temperature. (Jonsson.

H, et al, 2005, SMHI, 2009 and weather base, 2009)

Location Start of heating

Monthly average temperature (oC)

End of heating

Monthly average temperature (oC)

weather base SMHI weather

base SMHI

Kiruna 15-Aug 11 15-Jun 8 Umeå 01-Sep 8 01-Jun 12 Östersund 01-Sep 8 06-Jun 10 Karlstad 15-Sep 11 10 16-May 10 10 Stockholm 18-Sep 11 11 14-May 10 10 Goteborg 24-Sep 12 14 07-May 11 13 Average 10 11 10 11

In a supermarket, heating is required throughout the year. However heating demand reduced drastically when daily average outdoor temperature reaches 10oC. Therefore 10oC has been fixed as outdoor temperature below which heat recovery is required in the simulation models.

5.1.3 Other refrigeration and heat exchanger parameters To enable fair comparison of the three systems, field measurements from three supermarkets have been used as reference to fix the parameter of refrigeration and heat exchangers. The assumed values are presented in table 5.3.

Table 5.3: Parameters assumed in the models Description of parameter Value Evaporation temperature of low temperature cabinet -35oC Evaporation temperature of medium temperature cabinet -10oC Approach temperature difference of dry cooler 5K Approach temperature difference of condensers 5K LMTD for condenser of Cascade system 6K Internal heat exchanger efficiency R404A unit in CC system 50% Internal heat exchanger efficiency CO2 unit in CC system 20% Capacity of pump in percentage of cooling capacity of high stage(R404A) unit in cascade system

4%

Heat losses in the tubing of intermediate brine loop 10kW Oil cooler capacity of CO2 compressors 15% Heat losses from CO2 compressors 7% The performances of the compressors have been modeled by curve fitting the manufacturers’ data

29

5.2 Heat recovery with heat pump

Oil cooler

Inter-CoolerOil cooler

+ -

Heat pump COP= 3.6

Q_max =300kW T_hp in= 13o C

- Pressure

-TemperatureKA1/KA2

FA1/FA2

T_con,b,o =18oC

45oC

-10oC-35oC

dTapp=5K

dTapp=5KdTapp=5K

dTsc=2K

dTsc=2K

dTsc=2K

Qoilcooler=15%

Qoilcooler=15%

dTsh=10KdTsh,ex=10K

dTsh,ex=15K

dTsh=10K

Figure 5.2: Schematic of heat recovery with heat pump system solution

(TC1). This is a parallel refrigeration system with CO2 trans-critical system in both low and medium temperature levels. It has a single-stage compression system on the medium temperature unit and two-stage in the low temperature unit. The condensers in both units reject heat to the secondary coolant which is connected to the heat pump system on the heat recovery system. The return from the heat pump is connected to the roof top dry cooler. The heat pump on the heat recovery side is designed to operate with COP of 3.6 within the temperature limit at the heat pump evaporator of 13oC/7oC. Operating under this condition the supply temperature from the heat pump is 45oC. Alternatively the heat pump could be operated to supply temperature of 35oC to the heated space with COP of 4.7. In order to obtain the temperature of brine at inlet of the heat pump evaporator (T_hp, b,in in figure 5.2) of 13oC, the supply temperature of coolant from condenser (T_con,b,o) has to be 18oC. This is taking account of temperature difference of 5K between the fluids in the heat exchanger. The system has been simulated in the floating condensing mode. COP and pressure at which the corresponding supply temperature of brine from condenser is 18 is selected and the system is controlled to run at this operating condition for ambient temperature up to 10oC. Beyond this the refrigeration system is again set back to floating condition. Figure 5.3, shows the plot of COP in floating condensing

30

mode (COP_FC) and after the condition have been changed to operate the heat pump during the heating season (COP_HP_S45). In the heat pump mode the supply temperature of the heat pump is 45oC. The lowest condensing temperature was set to 10oC; therefore, in floating condensing mode the curve of COP is constant for ambient temperatures below 0oC. In the heat pump mode the COP is constant at about 2.6 for ambient temperature below 10oC and for ambient temperature above 10oC; the refrigeration system operates in floating condensing mode.

0

5

10

15

20

25

30

35

40

45

50

-10 -8 -7 -5 -3 -1 0 2 4 6 7 9 11 12 14 16 18 19 21 23 24 26 28 30 31 33 35 37 38 40T_ambient (C)

T(C

)

0

0.5

1

1.5

2

2.5

3

3.5

4

CO

P

T_hp,b,in T_con,b,o COP_FC COP_HP

Figure 5.3: COP in floating condensing and heat pump modes of operation for TC1.

At about 18oC ambient, the transition from sub- to trans-critical region takes place.

5.2.3 Evaluation of different modes of operation The system has been simulated to operate as heat pump with supply temperature of 45oC. Presented in the figure 5.4 are system’s COP operating in floating condition with coolant loop (COP_FC) and heat recovery mode (COP_HP). Also presented is the total compressor power of the refrigeration system in floating condensing (E_ref_FC) and recovery mode (E_ref_HP). Total heat delivered by the heat pump and the heating COP are plotted for the heating period. Power consumption of the compressors are only considered, therefore does not account the auxiliary power such as fan power or circulation pumps.

31

0

50

100

150

200

250

300

-10 -8 -7 -5 -3 -1 0 2 4 6 7 9 11 12 14 16 18 19 21 23 24 26 28 30 31 33 35 37 38 40

T_ambient (C)

E,Q

(kW

)

0

0.5

1

1.5

2

2.5

3

3.5

4

CO

P

E_ref_FC E_ref_HP E_tot_HP Q_HP(HR) COP_FC COP_HP COP heating Figure 5.4: Operation in the heat pump mode with heat pump supply

temperature of 45 o C compared to the floating condensing mode for TC1 As can be observed from figure 5.4, power consumption could increase by average of 35% as penalty of operating the refrigeration system at the higher pressure to operate the heat pump system. Heat capacity up to 187kW is available while operating under 50% of cooling capacity of the refrigerating units. The condensing temperature of the refrigeration system while operating in heat pump mode is about 19oC. This condensing temperature corresponds to operation of the refrigeration system at ambient temperature above 10oC. Therefore beyond outdoor temperature of 10oC, there is no increase in compressor power of the refrigeration units due to heat recovery. Performance in heat recovery mode is compared to the cases when the system operates in floating condensing with air cooled condenser and with condensing with coolant loop. System’s cooling COP with air cooled condenser (COP_FC_AC) and coolant (secondary fluid) cooled condenser (COP_FCB) is presented in figure 5.5. There is a shift in the curves of COP in two cases which is due to the difference in condensing temperatures. Due to additional temperature difference introduced by the coolant, for same ambient conditions, the system with secondary fluid runs at condensing temperature higher than air cooled condenser.

32

0

5

10

15

20

25

30

35

40

45

50

-10 -8 -7 -5 -3 -1 0 2 4 6 7 9 11 12 14 16 18 19 21 23 24 26 28 30 31 33 35 37 38 40

T_ambient (C)

Tem

pera

ture

(C)

0

0.5

1

1.5

2

2.5

3

3.5

4

CO

P

T_con_FCB T_con_FC_AC COP_FCB COP_FC_AC Figure 5.5: Operation of refrigeration system in air cooled condenser and

brine (secondary fluid) cooled condenser for TC1. To study the performance of heat pump with supply temperature of 35oC, the performance in the heat pump mode with supply temperature of 35oC is shown in figure 5.7. Heat pump operates at COP of 4.7, which is higher than the COP of het pump while operating with supply temperature of 45oC. This reduces the total power consumption. The average increase in the power consumption of refrigeration system as a result of operating in this mode is about 34% compared to that consumed when operating in floating condensing mode. Average COP of heating is 3.9.

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5

CO

P

E_ref_FC E_ref_HP Q_HP(HR) E_tot_HP COP_FC COP_HP COP_heating

Figure 5.7: System solution with heat pump supply temperature of 35oC in TC1.

At outdoor temperature higher than 10o C the refrigeration system is already operating at higher pressure so system can provide the required coolant temperature of 18oC, therefore there is no need for increasing the condensing pressure. Increase in total power at this point is only from heat pump. Another alternative mode of operation that is considered is to avoid the use of heat pump in the heat recovery system where the refrigeration system will operate as a heat pump. It is simulated by increasing the condensing pressure in the refrigeration system to insure a supply temperature of coolant (inlet to the heating system) out of the condenser of 45oC. The performance parameters of such system are plotted in the following figure 5.8.

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(kW

)

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1.5

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CO

P

Q_con_FC E_ref_FC Q_con_FH(HR) E_tot_FH COP_FC COP_FH COP heating

Figure 5.8: Operating at conditions to supply temperature of coolant from condenser at 45oC in TC1.

It can be seen that average COP of heating is about 2.5 and compressor power increases by more than 300% in extreme case and 270% in average. Also at this operating condition COP of refrigeration system is less than 1 and discharge pressure under this operating condition is about 114 bars. This would mean that compressors of medium temperature units operating at pressure ratio more about 5. Figure 5.9, shows the performance of system operating at fixed head pressure to supply a supply temperature from the condenser at 35oC.

35

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(kW

)

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CO

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E_ref_FC E_ref_FH Q_con_FH (HR) Q_con_FC COP_FC COP_FH COP_heating Figure 5.9: Operating at conditions to supply temperature of coolant from

condenser at 35oC in TC1. With average heating COP about 3.3 and COP of the refrigeration system about 1.2 during the heating season, it provides heat recovery capacity of 168 kW. Though capacity of heating is quite large and COP of heating is quite satisfactory, COP of the refrigeration system is reduced considerably. Since the basic case of reference is the floating condensing with air cooled condenser instead of the coolant loop and dry cooler case, it is interesting to compare the performance of the refrigeration system with heat pump to the refrigeration system with air cooled condenser as presented in figure 5.10. The compressor power consumption of the refrigeration system in air cooled floating condensing mode (E_ref_FC_AC) is compared to the power consumption when operating at fixed head pressure to recovery heat at 35oC (E_ref_FH). The compressor power consumption of the refrigeration system is higher in heat recovery mode than in air-cooled condenser mode even beyond the heating season. It is because of the temperature difference introduced by the secondary fluid used in the heat recovery mode.

36

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CO

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E_ref_FC_AC E_ref_FH Q_con_FH COP_FC_AC COP_FH COP heating

Figure 5.10: Comparison of air cooled mode and heat recovery mode with

supply temperature of 35oC in TC1 Since cooling demand tend to decrease when the heating is required in cold season, a ratio of heat that is available to the cooling capacity could be used as a parameter to compare different systems. The ratio of capacity of heat recovered to the cooling capacity for different modes of operation are presented in figure 5.11. The heat recovery ratio in heat pump mode with supply temperature of 45oC is represented as HP_S45 and heat pump mode with supply temperature of 35oC as HP_S35. The heat recovery ratio for fixed head of refrigeration system to supply heat at 45oC and 35oC are represent as FH_T_b,o 45 and FH_T_b,o 35 respectively. Beyond the heating period the ratio represent the ratio of heat rejected to cooling capacity.

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Rat

io o

f Hea

t rec

over

y to

coo

ling

capa

city

HP_S45 FH_T_b,o 45 FH_T_b,o 35 HP_S35 Figure 5.11: Ratio of heat recovery/rejection to cooling capacity in

different modes of operation for TC1 Considering the ratio of heat recovered to the cooling capacity, heat recovery with the fixed head pressure to supply temperature of brine out of condenser at 45oC is seen to be highest at almost double the cooling capacity. This is mainly due to the fact that this mode of operation has the highest power consumption. However, this ratios and power consumption in different systems must be compared for the different modes. Figure 5.12 shows the total compressor power in different mode of operation. In the same plot the compressor power in floating condensing mode without heat recovery and for the case if air cooled condenser would be used is included in the plot.

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Pow

er (k

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E_FC E_tot_HP_S45 E_FH_T_b,o 45 E_FH_T,b,o 35 E_tot_HP_S35 E_FC_AC Figure 5.12: Total power consumption in different mode of operation for

TC1. Comparing figure 5.11 and 5.12, Heat pump mode with supply temperature of 45oC has heating ratio of 1.8 and at the same time power consumption is just over 80% of the total power consumed by system operating with fixed head pressure with supply temperature of brine at 45oC. Comparatively for both the temperature levels of 35oC and 45oC, heat pump system has better performance than the fixed head pressure system.

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5.3 Heat recovery in de-superheater only (TC2)

Oil cooler

+ +

Floor heating

-Pressure

-Temperature

KAFA1/KAFA2

KA3

dTapp=5KdTapp=5K dTapp=5K dTapp=5KdTsc=0.5KdTsc=0.5K

Tevap -35oCdTsh=10K

dTsh=10K

dTsh=10K

Tevap -35

Tevap -10oC

dTsh,ex=15K

dTsh,ex=10K

Qoilcooler=15%

Qoilcooler=15%

Qoilcooler=15%

30

35

Figure 5.13: Schematic of refrigeration system with heat recovery in de-

superheater region (TC2). Such a system solution is already in operation in a supermarket in Sweden and analysis of field data is discussed in section 4.2. However, in the simulation models the sub-cooling from borehole is not considered in the simulation model discussed in this section. It is interesting to study this system solution for many aspects. The system solution is different from the earlier design on both refrigeration and heat recovery side. The refrigeration side consists of two parallel circuits. One of the parallel circuits consists of a medium temperature level. The second parallel circuit is a booster system consisting of medium temperature level circuit and a low temperature level circuit. In this design, the heat recovery takes place only in the de-superheater region and condensation takes place in the air cooled condenser. The system simulated in two modes of operation, to supply the temperature of recovered heat at 35oC and 45oC. In both mode of operation, the heat recovery capacity is then fixed at 30% of the total heat rejected at the condenser. This value is fixed based on the field measurements of a similar heat recovery system, discussed in section 4.2. Concerning the operation during heat recovery with supply temperature of 35oC, the system is simulated to operate with variable capacity of heat recovery; so the system was not controlled according to requirements of the heat recovery. Then the refrigeration system is set to operate at elevated operating pressure to recover 30% of the rejected heat at a temperature of 35oC. Similarly operating conditions are selected for heat recovery at 45oC. The COPs of the system operating in floating condensing, and fixed percentage of

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heat recovery at supply temperature of 35oC and 45oC are shown in Figure 5.14.

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T_ambient (C)

CO

P

COP_FC COP_FH_35 COP_FH_45 Figure 5.14: Coefficient of performance of system in different modes of

operation in TC2 In the following figures (5.15 and 5.16) some values beyond the heat recovery period are not shown in the plot, these values have not been included because they are not much of interest for the current analysis. Figure 5.15 shows the heat recovery and power consumption in floating condition and operation at fixed head pressure during heating period.

41

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CO

P

E_FC E_FH Q_HR_FH Q_HR_FC COP heating

Figure 5.15: heat recovery at supply temperature to heat recovery system of 35oC in TC2

The power consumption in the refrigeration system due to the heat recovery of 30% of the total heat rejected in the de-superheat region increases by 54% and subsequent increase amount of heat recovery is over 100 %. Maximum COP of heating is just over 1.4. When the refrigeration system is operated at an elevated pressure to maintain heat recovery supply temperature of 45oC, the power consumption increases by 85 % and increase in quantity of heat recovery is 180 %. This can be observed in figure 5.16.

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1.1

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CO

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Q_FC E_FC E_FH Q_FH COP heating Figure 5.16: Heat recovery at supply temperature to heat recovery system

of 45oC in TC2 Since in both the cases the heat recovery capacity is fixed at 30% of the total heat rejected, the ratio of heat recovery capacity to cooling capacity at 35oC and 45oC is same and equal to 0.4. Thus, the compressor power consumption is an important parameter for the comparison. Compressor power consumption in different mode of operation is presented in figure 5.17.

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T_ambient (C)

Pow

er (k

W)

E_FC E_FH_35 E_FH_45 Figure 5.17: Compressor power for different capacity and temperature of

heat recovery in TC2 It can be seen that the compressor power consumption is considerably higher to operate the system in heat recovery mode for supply temperature of 45oC. It is because the refrigeration system needs to be operated at higher condensing pressure to recover same capacity of heat at 45oC compared to the 35oC case. So far the percentage of heat recovery is only 30% of the total heat rejected. The performance of the system with different amount of heat recovery was considered and performance to recover 70% of the heat at 35oC is presented in figure 5.18.

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Q_HR_FC Q_HR_FH E_FH E_FC COP heating Figure 5.18: Performance of TC2 with heat recovery of 70% of the total heat rejected at de-superheater region at supply temperature of 35oC

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5.4 Heat recovery in R404A/CO2 cascade system

Figure 5.19: Schematic of heat recovery system solution in R404A/CO2

cascade system (CC) This is a cascade system solution with R404A in the high stage and CO2 in the low temperature. Brine is circulated in the medium temperature cabinets and the CO2 condenser. Heat recovered in two temperature levels on the R404A heat rejection side in two separate loops, de-superheater and the condenser region.

First the system is operated in the floating condensing. To recover heat at supply temperature of 35oC, the refrigeration system will have to operate at the condensing temperature of about 36oC which results in COP of 1.7. Figure 5.20 shows the operation of refrigeration system in floating condensing mode and fixed head pressure to provide brine supply temperature of 35oC. Top two lines in the plot are COP of the refrigeration system on floating condition (COP_FCB) and with the conditions set for heat recovery (COP_FH). The two COP lines join above an outdoor temperature of 10oC and follow the shape of floating condensing mode. The pair of curves below this plot shows heat rejected on

46

floating condensing (Q con_FCB) and on heat recovery mode (Q con_FH(HR)). Next pair of curves is the represents the compressor power consumption in floating condensing mode (E_FCB) and in heat recovery mode (E_FH). The last curve is the heat from de-superheater (Q_desup_FH).

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Q con_FCB E_FCB E_FH Q con_FH(HR) Q_desup_FH COP_FCB COP_FH Figure 5.20: Heat recovery with supply temperature of 35oC in cascade

system (CC)

Power consumption of the refrigeration system increase by 63% when, operating under heat recovery mode to deliver supply temperature of 35oC. Out of 169 kW of total heat recovery capacity 42 kW is high temperature heat from the de-superheater at about 76oC. It is seen from the plot that considerable capacity of heat from de-superheater is available even while operating on floating condensing mode. Therefore it can be used for other applications in the supermarket throughout the year. Average COP of heating in this system is about 7. To recover heat at condenser supply temperature of 45oC, the refrigeration system is made to operate at corresponding COP of 1.4 during the heat recovery period. Figure 5.21 compares the performance with that of floating condensing mode.

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Q con_FCB E_FCB E_FH Q con_FH(HR) Q_desup_FH COP_FCB COP_FH Figure 5.21: Heat recovery with supply temperature of 45oC in cascade

system (CC) In case of heat recovery with supply temperature of 45oC average, the increase in compressor power consumption is more than 94%, when the outdoor temperature is below 1oC, the power consumption is about 200% of the power that refrigeration system would consume under floating condition. Average COP of heating is 5. The system performance of heat recovery with supply temperature of 35oC has been compared to the refrigeration system operating under air cooled floating condensing mode in figure 5.22. Presented in the plot is capacity of heat recovery during the heating season (Q con_FH (HR)), COP of refrigeration in air cooled floating condensing mode (COP_FCAC) and heat recovery mode (COP_FH). The last pair of curves is the compressor power consumption in heat recovery mode (E_FH) and air cooled condensing mode (E_FCAC).

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E_FCAC E_FH Q con_FH(HR) COP_FCAC COP_FH Figure 5.22: Comparative performance of heat recovery with 35oC and air-cooled condenser in floating condensing mode in cascade system (CC)

One way to study the system performance is to compare the heat recovery ratio at different temperature levels of heat recovery. Figure 5.23 presents the heat recovery ratio at 45oC (FH_T, b, o, 45), 35oC (FH_T, b, o, 35) and on floating condensing (FCB) mode.

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io o

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ling

capa

city

(HR

R)

FCB FH T_b,o 35 FH T_b,o 45 Figure 5.23: heat recovery ratio for different heat recovery temperature

levels and floating condensing mode in CC.

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It can be seen that heat recovery ratio is higher for heat recovery at 45oC than for 35oC and floating condensing mode. However net increase in compressor power consumption is also considerably high as can be compared in figure 5.24. The top tow curves represents the power consumption for heat recovery at 45oC (E_FH_T,b,o 45) and 35oC (E_FH_T,b,o 35). Two curves in the bottom are power consumption when operating the refrigeration system in floating condensing (E_FCB) and with air cooled condenser (E_FCAC) mode.

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E_FCB E_FH_T,b,o 35 E_FCAC E_FH_T,b,o 45

Figure 5.24: Power consumption in different modes of operation in cascade system (CC)

5.5 Comparison of different heat recovery system solutions To compare different system solutions it is proper to begin with the discussion on performance of different refrigeration system solution operating in floating condensing mode. Figure 5.25 presents the COP of refrigeration of different system solution in floating condensing mode. It can be see that CO2

trans-critical system 1 and CO2 trans-critical system 2 has better performance at lower outdoor temperatures. Trans-critical system 2 has the highest COP and cascade system has the lowest COP for outdoor temperature below 10oC. For outdoor temperatures higher than 23oC, cascade system has the highest COP and trans-critical system 1 has the lowest.

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TC1 TC2 CC

Heat recovery

Figure 5.25: COP of TC1, TC2 and CC in floating condensing mode.

Some parameters that could be used to compare different heat recovery system solutions are COP of refrigeration system, COP of heating and heat recovery ratio. The comparison of these parameters for TC1, TC2 and CC for heat recovery temperature of 35oC and for the heating season is presented in figure 5.26. From left in the figure are performance of TC1 in heat pump mode and without heat pump; the TC2 system to recover 30% and 70% of heat respectively. The last group of columns is the performance of cascade system (CC). It can be seen from the figure that TC1 with heat pump system has the highest COP and TC1 without the use of heat pump has the lowest COP. The cascade system has the highest COP of heating and HRR ratio while TC2 with heat recovery percentage of 30% of the total heat rejected has the lowest COP of heating and HRR.

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0.00

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TC1 with heat pump TC1 without heat pump TC2 with 30% heatrecovery

TC2 with 70% heatrecovery

Cascade (CC)

Heat recovery system solutions

CO

P,H

RR

, CO

P he

atin

gCOP COPheating HRR

Figure 5.26: Heat recovery ratio for different heat recovery system

solutions for at supply temperature of 35oC. Though the comparisons of the system solutions are self explanatory, this study is insufficient to draw general conclusion on the performance of the systems.

6 Discussion With the drive to reduce net energy consumption, heat recovery is gaining popularity in supermarket refrigeration systems. In Sweden most of the new supermarkets are installed with heat recovery systems. Field measurements of trans-critical system 1 and trans-critical system 2 did give some insight on the performance of heat recovery. However more detailed measurement for longer period is necessary for trans-critical system 2. In case of trans-critical system 1, actual field measurement on the heat recovery system would have provided more information on the performance of the system. Never the less; the analysis of field measurement in trans-critical system 1 and trans-critical system 2 has highlighted enough details to further the discussion of the subject. For trans-critical system 1, if the refrigeration system was operated without heat recovery, condensing temperature was as low as 13oC, where as to operate in heat recovery mode with heat pump the minimum condensing temperature has to be 20oC. This puts burden on the refrigeration system by reducing the COP of refrigeration system. But on the other hand, large capacity of heat is available. The power consumption as calculated is for the maximum capacity therefore the high values of power consumption by the heat pump may be misleading. In reality the heat pump could be running in part load thus lower

52

consumption. The measurement of the heat recovery side in the future studies would enable better observation of the heat recovery performance. From the field measurement of trans-critical system 2, it was found that up to 30% of the total heat rejected can be recovered at a temperature of 35oC. The rest of heat is rejected to the ambient through the air cooled condenser. It is however seen though amount of heat recovered is less compared to trans-critical system 1 the condensing temperature has to be raised above 20oC, which increases the pressure ratio thus the COP of refrigeration drops. There is also additional consumption of power by the condenser fans to reject the excess heat. The performance of this system is better because there is the external sub-cooling from the ground source which improves its COP. The field measurement for cascade system has not been conducted in this study. As discussed in section 5.5, using the models the comparative different system solutions could be compared for system performance such as COP of refrigeration system. COP of heating and heat recovery ratio (HRR). It is however necessary to conduct detailed field measurements and detailed analysis prior to drawing conclusion on the system performances. Further for completed evaluations of the systems it is necessary to study the annual energy consumption of the systems taking consideration of cooling and heating loads over the year in different climates.

7 Conclusion With the energy price continuing to rise, reducing the net energy purchase is a goal of any enterprise. Many new supermarkets refrigeration systems in sweden are equiped with the heat recovery system. While it has presented the opportunities for energy conservation there remains the challenge of matching the temperature and the amount of heat recovered to the demand in the supermarket. In this thesis the field measurement have been performed in two super markets with CO2

trans-critical system. The heat recovery systems were different in the two cases. In CO2 trancritical system 1, where heat recovery system with a heat pump it is found that the minimum condensing temperature of the the refrigeration has to be 20oC compared to possible lower codensing temperature of 10oC. The maximum COP of heating is limited to design COP of the the heat pump. In CO2 trans-critical system 2, the condensing temperature during the heat recovery mode is about 23oC. The increased condensing temperature doesn’t effect the COP because of the enhancement in the cooling capacity due to external sub cooling from the ground source. Heat rejection from oil cooler in both the cases were found to be more or less constant over the year. The heat recovery capacity from the inter cooler was found to be constant too. One of the problems faced in this project was to get the reading at the point of interest in both the refrigeration side and on the heat recovery side. The comprehensive field measurement that records reading at both available heat

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on refrigeration side and heat using end such as HVAC and service water heater is necessary to give better indication of system performance. Future works should incorporate the field measurement of cascade system, long term measurement of trans-critical system 1 and trans-critical system 2. The study must also incorporate the measurement of annual energy consumption taking consideration of cooling and heating needs in different climates. Theoritical analysis was carried out using the computer simulation models for three different system solutions: CO2 trans-critical system 1, CO2 trans-critical system 2 and R404A/ CO2 cascade system. In reality the systems solutions vary in some form or other. This makes it difficult to compare different systems based on a certain performance criteria such as COP refrigeration system, COP of heating, power consumption heating and cooling capacity etc. Therefore computer simulation modeling is important to develop fair comparison of different systems.

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8 References Arias J, 2005 “Energy Usage in Supermarkets-Modelling and Field measurements”, Doctoral thesis, Department of Energy Technology, KTH, Stockholm, Sweden Arias J, Lundqvist P, 2006 “Heat recovery and floating condensing in supermarkets” Journal of Energy and Buildings, Vol.38, pp. 73-81 Brownell K. A, 1998 “Fixed and Floating Head Comparison for Madison Ice Arena” Report, Energy Centre of Wisconsin, WI, USA Dorin (2009), “The widest CO2 compressor range, carbon dioxide for all your needs”, http://www.dorin.com/documents/Download/19/CO2_-0809a.pdf accessed on march 2009 Granryd E, Ekroth I, Lundqvist P, Melinder Å, Palm B, Rohlin P, 2003 “Refrigeration Engineering” Department of Energy Technology, KTH, Stockholm, Sweden. Iwmac (2009), “Centralised operation and surveillance, by use of WEB technology”, available at: http://www.iwmac.no/english/ Minea V, 2006 “Improved Supermarket Refrigeration and Heat Recovery System” ASHRAE Transaction: Vol.112, Part 2, pp. 592-596 Minea V, 2007 “Energy Efficiency of a Supermarket Refrigeration/Heat Recovery System with Secondary Fluids” International Congress of Refrigeration, 2007, Beijing, China Johansson S, 2009 “Evaluation CO2 supermarket refrigeration systems”, MSc Thesis, Department of Energy Technology, KTH, Stockholm, Sweden Reindl D. T, Jekel T.B, 2007 “Heat Recovery in Industrial Refrigeration” ASHRAE Journal August 2007 Wulfinghoff D. R, 1999 “Energy Efficiency Manual” Energy Institute Press, Maryland, USA

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