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Applied Thermal Engineering 30 (2010) 1946e1958

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Applied Thermal Engineering

journal homepage: www.elsevier .com/locate/apthermeng

Energy performance of supermarket refrigeration and air conditioningintegrated systems

Luca Cecchinato, Marco Corradi, Silvia Minetto*

Dipartimento di Fisica Tecnica, Università di Padova, via Venezia 1, I-35131 Padova, Italy

a r t i c l e i n f o

Article history:Received 18 December 2009Accepted 16 April 2010Available online 27 April 2010

Keywords:Display cabinetEfficiencyHeat recoveryHeat pumpSupermarketWater chiller

* Corresponding author. Tel.: þ39 049 827 6879; faE-mail address: silvia.minetto@unipd.it (S. Minett

1359-4311/$ e see front matter � 2010 Elsevier Ltd.doi:10.1016/j.applthermaleng.2010.04.019

a b s t r a c t

The electricity consumption for air conditioning and refrigerated cases in large supermarkets representsa substantial share of the total electricity consumption. The energy efficiency of supermarkets can beimproved by optimising components design, recovering thermal and refrigerating energy, adoptinginnovative technology solutions, integrating the HVAC system with medium temperature and low-temperature refrigeration plants and, finally, reducing thermal loads on refrigerated cases. This study isaimed at investigating the performance of different lay-out and technological solutions and at finding thepotential for improving energy efficiency over the traditional systems in different climates. In theanalysis chillers and heat pumps working with R410A, medium temperature systems working withR404A and low-temperature systems working both with R404A and R744 were considered. The inves-tigated solutions enable an annual energy saving higher than 15% with respect to the baseline solutionfor the considered climates.

� 2010 Elsevier Ltd. All rights reserved.

1. Introduction

Energy saving and F-gas emission reduction have been topissues for the last two decades. Energy production and use haveundergone a rationalisation process in almost all application fields.In particular, since in many countries the structure of energyconsumption by different consumer groups is characterised bya high share of the residential and tertiary sector, which is veryoften responsible for about 40% of the total final energy demand,a great research effort is taking place to minimise their direct andindirect contribution to greenhouse effect. European CommunityDirective 2002/91/CE [1] proposes important guidelines for theimprovement of building energy efficiency; in Italy the Directivewas applied by the national regulation DL 192/05, issued in 2005[2]. Data supplied by the main Italian energy distribution Company[3] show that 25% of the Italian energy consumption is related tocommercial services and, amongst them, supermarkets are themain consumers. It is therefore of the greatest importance topropose new technologies that can improve energy efficiency insupermarkets and that can represent an acceptable trade-offbetween installation and running costs.

x: þ39 049 827 6896.o).

All rights reserved.

The analysis of each contribution to the total energy bill ofsupermarkets is the first step towards energy efficiency improve-ment. According to the Canadian Energy Efficiency Office [4], foodpreservation contributes for about one third to the total energyconsumption of supermarkets, while lighting and air conditioningaccount for 25% and 20% respectively. Efficient design of refriger-ation plants is mandatory for energy saving in supermarkets.Proposed technical solutions include the adoption of evaporativecondensing [5], the application of heat recovery and floatingcondensation [6] and mechanical subcooling of low-temperaturesystems [7e9].

Cogeneration and trigeneration have been recently proposed asfeasible solutions for supermarket applications to produce elec-tricity and to make cooling capacity available by means of anabsorbtion unit. Maidment and Tozer [10] review a number ofcombined cooling heat and power (CCHP) options involving the useof different cooling and engine technologies. The pay-back time fora CCHP system applied to a supermarket has been estimated to bearound 9e10 years by Arteconi et al. [11].

When air conditioning and space heating are demanded toliquid chillers and heat pumps, there is a real possibility of inte-grating vapour compression cycles operating at different temper-ature levels and of recovering waste heat deriving fromrefrigeration plants [6,12]. As in supermarket there are normallyvapour compression cycles operating at three different tempera-ture levels, corresponding to air-conditioning units, medium and

Nomenclature

A heat transfer surface [m2]Cc cycling degradation factor [e]COP heat pump energy efficiency [e]cp specific heat at constant pressure [J kg�1 K�1]d heat exchanger hydraulic diameter [m]EER chiller energy efficiency [e]F corrective factor [e]h enthalpy [J kg�1]hd heat dissipation factor [e]HER cabinets heat extraction rate [W]k transmittance correction coefficient [e]_m mass flow rate [kg s�1]P power consumption [W]PLF part load factor [e]PLR part load ratio [e]p pressure [Pa]Q thermal power [W]R thermal power ratio [e]Re Reynolds number: Re ¼ ð _m$dÞ=ðS$mÞ [e]RH relative humidity ratio [%]S cross sectional area [m2]T temperature [�C]UA thermal transmittance [W K�1]_V swept volume per time unit [m3 s�1]x specific humidity ratio [e]

Greek lettersa heat transfer coefficient [W m�2 K�1]

D difference [e]hel electric efficiency [e]his isentropic efficiency [e]hv volumetric efficiency [e]q logarithmic mean temperature difference [K]r density [kg m�3]

SubscriptsA ambientAHUc air handling unit cooling loadAHUh air handling unit heating loadaux auxiliaryc condenser/condensationC cabinetD designdb dry-bulbe evaporator/evaporationk compressorin inletload actual load conditionLT low-temperature applicationMT medium temperature applicationN nominalout outletr refrigerantSC subcoolingset set pointsf secondary fluidSH superheatwb wet-bulb

Fig. 1. Baseline system (System 0).

L. Cecchinato et al. / Applied Thermal Engineering 30 (2010) 1946e1958 1947

low-temperature refrigeration plants respectively, it is possible toconcatenate the systems according to a cascade lay-out, where theunit operating at the lower temperature level rejects the conden-sation or subcooling heat to the system at the higher temperaturelevel [13,14]. The integration may be operated directly, i.e. couplingevaporators, condensers and liquid subcoolers of different units,either using a secondary fluid.

Gianni and Oliva [15] suggested the adoption of a water loopheat pump system integrated with a heat recovery system for airconditioning in shopping centres. The authors presented a tech-nical and economical comparison of this system with a traditionalone, based on experimentation, reporting a 28% energy saving overthe total air-conditioning costs. Yang and Zhang [16] analysed theenergy saving potential of integrated supermarket HVAC andrefrigeration systems using three subcoolers. The subcoolers areplaced between the refrigeration medium and low-temperaturesystems and between the air-conditioning system and the mediumand low temperature plants. The energy saving associated to theproposed solution is strictly dependent on the air-conditioning andrefrigeration load ratio. The authors pointed out an energy savingequal to 15.6 and 6.5% with a ratio of 1 and 10 respectively. Eventhough Yang and Zhang [16] carried out a model based analysis onthe energy saving potential of integrated supermarket HVAC andrefrigeration systems usingmultiple subcoolers, they compared thedifferent solutions considering only two operating conditions,namely 35 and 18.3 �C external air temperature.

In this paper a seasonal dynamic building energy simulation iscarried out to compare different plant solutions integrating air-conditioning liquid chillers andheatpumpswithrefrigerationunits inthe climates of Treviso (Northern Italy), Stockholm and Singapore. Inorder to evaluate the different systems energy consumption, a math-ematicalmodel providing an accurate estimation of refrigeration and

air-conditioning units cooling or heating capacity and powerconsumption out of nominal rating conditionswas developed. R410Awas chosen as the working fluid for air-conditioning systems;refrigeration units for medium temperature applications wereassumed to use R404A, while for freezers units both R404A and R744were considered.

2. Plant options

The different plant options are presented in the following. Thebaseline plant, System 0 (Fig. 1), does not include any integrationbetween air-conditioning and refrigeration units. Two independentdirect expansion refrigeration machines for medium temperature(MT) and low temperature (LT) applications are used; the refrig-erant is R404A. An air condensed heat pump provides cold/hot

Fig. 2. System 1.Fig. 4. System 2. Wintertime operations.

L. Cecchinato et al. / Applied Thermal Engineering 30 (2010) 1946e19581948

water to the air handling unit (AHU) of the building variable airvolume (VAV) air-conditioning system.

System 1 (Fig. 2) provides the simplest level of integrationbetween refrigeration units. Saturated liquid out of the condenserof the LT refrigeration unit is subcooled by means of the evapora-tion of a certain amount of refrigerant in the MT refrigeration unit.

System 2 introduces the integration between the air-condi-tioning system and the refrigeration units; during summertime, theliquid chiller, beyond providing the air-conditioning unit with coldwater, subcools liquid out of both the MTand LT condensers (Fig. 3).During winter, MT and LT condensation waste heat is completelyrecovered for supplying the AHU hot coil with water at 45 �C.Excess heat is rejected using a finned coil condenser. A heat pump,in parallel with the heat recovery system, provides the additionalheat, whenever necessary (Fig. 4). System 2 includes also the sub-cooling of the LT refrigeration unit by the MT refrigeration unit, asdescribed for System 1.

System 3 consists in a typical cascade system, where the LTrefrigeration unit rejects the condenser heat to theMT refrigerationunit. TheMT unit is operated on R404A, while the LT unit uses R744.In summertime, the chiller provides theMT unit with coldwater forrefrigerant subcooling, while feeding the AHU cold coil (Fig. 5). Inwintertime, total heat recovery is applied to the MT refrigerationunit for direct heating service; the heat pump supplies the systemwith the additional heat (Fig. 6).

System 4 during summer overlaps system 2. In wintertime, thewaste heat from the condensers of theMTand LT refrigeration unitsis recovered in a water loop working as the heat source of the

Fig. 3. System 2. Summertime operations.

water/water heat pump (HP 1). Additional heat, if need be, issupplied by the air/water heat pump (HP 2) (Fig. 7).

System 5 during summer overlaps system 3, except for beingR744 the working fluid in the LT refrigeration unit. In wintertime,the waste heat from the MT refrigeration unit condenser is recov-ered in a water loop working as the heat source of the water/waterheat pump (HP 1). Additional heat, if need be, is supplied by the air/water heat pump (HP 2) (Fig. 8).

3. Building simulation

A single-floor building with a total area of 3620 m2 was simu-lated. Its height is 5 m. The longest dimension is southenorthoriented. The total area is divided into three parts: the supermarketarea is 3019 m2; the cafeteria area is 150 m2 and the shops area is450 m2. The cafeteria has 25.3 m2 windows south oriented and35.2 m2 east oriented. The shops area has 170.0 m2 windows southoriented and 59.3 m2 west oriented. Simulations were based onTreviso, Stockholm and Singapore test reference years; Treviso isa town located in the North-East of Italy (North 45� 400, East 12�

150). Monthly maximum and minimum mean average tempera-tures for the three cities are reported in Table 1.

The supermarket was considered to be located on a flat ground,without obstacles to the direct solar radiation, such as trees or build-ings, and the ground reflection coefficient was assumed equal to 0.2.

Table 2 summarises the supermarket MT and LT display casesand cold rooms main dimensions, their nominal heat extraction

Fig. 5. System 3. Summertime operations.

Fig. 6. System 3. Wintertime operations.Fig. 8. System 5. Wintertime operations.

L. Cecchinato et al. / Applied Thermal Engineering 30 (2010) 1946e1958 1949

rate (HERN) and the area where they are located. Nominal power isrelated to standard ISO 23953 Class 3 [17], i.e. 25 �C and 60% relativehumidity. The reference code identifies the cabinets according toISO 23953 [17].

The contribution to internal thermal loads related to lighting,equipment and people are detailed in Table 3. For the bakery area(51.6 m2), which is included in the supermarket, the internal loaddue to equipment was assumed to be equal to 50 W m�2, as rec-ommended by Stefanutti [18].

The resulting nominal display cases and cold rooms contributionto internal thermal loads is respectively �34.1 W m�2 and�26.6Wm�2, respectively, during opening (MoneSat 6 a.m.e8 p.m.)and closing hours (MoneSat from 8 p.m. to 6 a.m., Sunday).

The air-conditioning plant was assumed to be active 14 h perday (from 6.00 a.m. to 8.00 p.m.), six days per week. In summer-time, the air-conditioning set point is 26 �C, while in thewintertimethe set point for ambient heating is 20 �C. During the openinghours, the maximum relative humidity is set to 60%: above thislimit, dehumidification is activated. Air change rate is assumed tobe 10�2 m3 s�1 per person in the cafeteria, 1.2 � 10�3 m3 s�1 per m2

of shops area and 8 � 10�3 m3 s�1 per person in the supermarket.Air infiltration was considered negligible in the building air-

conditioned spaces because of the presence of door air curtains. TheAHU is equipped with a 60% effective heat recovery system and aneconomizer which is active from 11 �C to 25 �C external airtemperature. An economizer is a damper opening that draws up to100% outside air when the outside air is cooler than the tempera-ture inside the building, thereby providing free cooling. The designAHU flow rate is 8.47 m3 s�1. The supermarket was simulated with

Fig. 7. System 4. Wintertime operations.

Energy Plus [19] to calculate at hourly intervals the thermo-hygrometric variables and the thermal energy exchanged.

A particular effort was spent taken to modelise the interactionbetween the display cases and the surrounding ambient. Thermalloads of display cases and cold rooms depend on the supermarketair thermoigrometrical conditions, TA and RHA. In particular, in thesurroundings of vertical open cabinets, the cold air leakage createsthe so-called cold aisle effect, where temperature TA,C and relativehumidity RHA,C can be very different from the average supermarketair thermoigrometrical conditions. Orphelin et al. [20] show thedependence of cold aisle temperature from the average ambienttemperature and humidity. Equations system (1) was obtainedfromOrphelin et al. [20] data. The influence of the other cabinets onthe surrounding ambient temperature and relative humidity wasneglected.

�TA;C ¼ TA;CðTA;RHAÞ ¼ 6:5$10�3$T1:473

A $RH0:6067A $

RHA;C ¼ RHA;C�TA;C ; xA

�$

(1)

To correctly evaluate the display cases thermal load, correctivefactors were introduced to emend the nominal HER. Such coeffi-cients are dependent, in the case of open cases, on temperature andhumidity of the air entering the cabinet while they are onlya function of average ambient temperature in the case of closedcabinets. The following general equation (2) was derived fromtechnical datasheet of a refrigeration equipmentmanufacturer [21].It expresses the corrective factor FHER for the cabinets nominal HER(declared at 25 �C air temperature and 60% relative humidity), asa function of ambient temperature TA,C, and relative humidity RHA,C

near cabinets.

FHER ¼ FHER�TA;C ;RHA;C

�¼ a1$RH$TA;C þ a2$TA;C þ a3$RHA;C þ a4$ (2)

Table 4 summarises the coefficients of equation (2) for theconsidered cabinets. In the case of closed cabinets (VF4), ambientrelative humidity does not influence HER.

Table 1Monthly maximum and minimum mean average temperature, LT design load andMT and AHU load ratios for the different climates.

Max Min Load ratios

Tdb[�C]

Twb

[�C]Tdb[�C]

QLT

[kW]RMT-LT

[e]RAHUc-LT[e]

RAHUh-LT[e]

Treviso 22.6 17.5 3.3 26.5 2.2 2.6 7.8Stockholm 17.0 11.6 �3.6 25.5 2.1 2.0 9.3Singapore 28.5 25.0 26.3 27.5 2.3 5.8 3.2

Table 2MT and LT display cases. Nominal HER (HERN).

Display Cabinets Reference Lenght HERN Area

MT cabinet [e] [m] [kW] [e]

MT vertical, multi-deck VC2 40.0 41.1 SupermarketMT open, wall site HC3 22.5 10.2 SupermarketMT open, island HC4 7.5 3.0 SupermarketLT open, wall site HF3 7.5 6.0 SupermarketLT vertical, glass door VF4 20.0 20.0 Supermarket

Cold rooms Reference Volume HERN Area[e] [m3] [kW] [e]

MT e 224.0 33.0 SupermarketLT e 30.5 4.0 Supermarket

MT total [kW] 89.3

LT total [kW] 30.0

Table 4Coefficients used in equation (1).

Cabinet Reference a1 a2 a3 a4

VC2 5.08 � 10�4 2.46 � 10�2 �2.88 � 10�3 �1.77 � 10�1

HC3, HC4 1.73 � 10�4 2.95 � 10�2 �3.17 � 10�4 3.38 � 10�2

VF4 0 2.40 � 10�2 0 4.00 � 10�1

HF3 8.21 � 10�5 1.84 � 10�2 6.41 � 10�5 4.15 � 10�1

L. Cecchinato et al. / Applied Thermal Engineering 30 (2010) 1946e19581950

The required HER of each cabinet can be calculated by multi-plying the nominal HER by the corrective factor FHER, as expressedin equation (3):

HER ¼ HERN$FHER$ (3)

During the closing hours, night blinds reduce the HER of verticalmulti-deck cabinets by 45%. The HER of each cabinet, expressed asa function of the local thermal hygrometric variables, contributes tothe energy conservation equation of each building zone in thesimulation. According to Orphelin et al. [20], only 80% of displaycabinets HER contributes to the building energy equation, being20% of the HER related to cabinet internal gains (lighting, fans,heaters, etc). The average temperature and humidity of specificzones depend on many variables (conduction through the walls, airsupply, radiative loads, internal loads and display cases HER). Thecold aisle phenomenon was not considered in the calculation ofthermal load from the walls.

Simulations exhibit a maximum heating demand of 207.2, 238.7and 86.9 kW in Treviso, Stockholm and Singapore respectively. Thecooling peak is 69.3, 50.0 and 158.9 kW in the three cities respec-tively. In Table 1 the maximum LT system load resulting frombuilding simulation is reported together with the peak load ratiosof the MT and of the AHU to the LT system (RMT-LT, RAHUc-LT, RAHUh-LT). It appears that Stockholm loads are similar to Treviso ones butthe RAHUh-LT ratio is higher because of the colder climate. Singaporeis characterised by a much lower RAHUh-LT ratio and a more thandouble RAHUc-LT ratio. This is due to the hotter and more humidclimatic conditions.

4. Air-conditioning and refrigeration unit design

The peak value of heating and cooling loads, as resulting fromthe simulation which is described in the previous section, wasconsidered for the design of the air-conditioning and refrigerationunits. The evaporation temperatures of the MT and LT refrigerationunits were considered to be respectively �10 �C and �35 �C.

Superheat was set to 5 �C for all evaporators, while no sub-cooling was assumed at the condensers exit. During heating

Table 3Contributions to internal loads [18].

Area Lighting [W m�2] Equipment [W m�2] People [persons m�2]

Cafeteria 30 3 0.20Shops 15 3 0.10Supermarket 15 3 0.17

operations, water set point was 45 �C, while it was 7 �C whencooling is required. Tables 5 and 6 summarise the design conditionsfor refrigeration and air-conditioning units. In the case of cascadesystems (Systems 3 and 5), evaporation temperature and superheatof the MT refrigeration units are respectively Tsf,in and DTsf, as listedin Table 5.

Tc is condensation temperature, Tc,sf,in is the temperature of thesecondary fluid entering the condenser and DTsf is its temperaturelift. Te,sf,in is the temperature of the secondary fluid to the evapo-rator. _V indicates the compressors total swept volume. StockholmSystems 2e5 HP2 (air/water heat pump) was dimensioned at�3 �Cair temperature, supplementary heating is supplied througha condensing boiler of constant 1.1 energy efficiency.

For System 0 in Treviso, Bitzer 4CC-6.2Y, Bitzer 6G-30.2Y andCopeland ZP295KCE-TWD compressors were considered for theMT, LT and HP1-2 units respectively. Treviso System 3 adopts Bitzer2EHC-3K in R744 LT plant. For all other systems and climates, inorder to obtain a fair comparison, the swept volume was changedaccording to Tables 5 and 6, considering the same compressorsisentropic and volumetric efficiencies curves.

The design conditions for heat exchangers water/refrigerant SR1(Figs. 3 and 5), MT refrigerant/LT refrigerant SR2 (Figs. 2e4 and 7)are presented in Table 7 where Tr,in is the refrigerant inlettemperature, Tsf,in the secondary fluid inlet temperature, DTapproach is the minimum temperature approach between thefluids, TLT,in and TLT,out are respectively the inlet and outlettemperatures of the LT refrigerant and TMT,in and TMT,out the inletand outlet temperatures of the MT refrigerant.

The possibility of varying the chiller/heat pump water deliverytemperature according to the actual cooling/heating load wasintroduced. It has been widely demonstrated in the technicalliterature [22] that the modulation of the water temperature canlead to relevant energy saving. During summertime the watertemperature can vary in the range 7e16 �C, if the required coolingcapacity is inside the interval 20e100% of the design value, while itstays at 16 �C under 20%. In the case of Systems 2e5, the totalcooling load includes also the subcooling demand of the refriger-ation units. Contrary to Yang and Zhang [16] approach, the chilleralways operates the MT and/or LT plants subcooling even withoutAHU cooling demand.

During winter, the water delivery temperature is modulatedbetween 45 �C and 30 �C in the 20e100% interval of the designheating capacity, and never drops below 30 �C.

5. Air-conditioning and refrigeration units simulation

In order to evaluate the energy consumption of refrigeration andair-conditioning units a mathematical model providing an accurateestimation of refrigeration and air-conditioning units cooling orheating capacity and power consumption out of nominal ratingconditions was developed. The proposed procedure corrects thedesign transmittance taking into account the heat transfers coef-ficients variations associated to refrigerant and secondary fluidmass flow rates variations.

Table 5Design conditions for MT and LT refrigeration units.

System Season MT unit LT unit

Tc Tc,sf,in DTsf _Vb Tc Tc,sf,in DTsf _Vb

[�C] [K] [3600 m3 s�1] [�C] [K] [3600 m3 s�1]

0 Winter e 130 e 256e 100 e 184e 140 e 265

Summer 45/37a 35/27a 5 45/37a 35/27a 5

1 Winter e 166 e 130e 128 e 109e 178 e 135

Summer 45/37a 35/27a 5 45/37a 35/27a 5

2 Winter 50 40 5 191 50 40 5 144Summer 45/37a 35/27a 5 178 45/37a 35/27a 5 139

206 149

3 Winter 50 40 5 216 �7 �10 5 14Summer 45/37a 35/27a 5 203 14

231 15

4 Winter 20 10 5 90 20 10 5 130Summer 45/37a 35/27a 5 84 45/37a 35/27a 5 125

97 135

5 Winter 20 10 5 137 �7 �10 5 14Summer 45/37a 35/27a 5 127 14

148 15

a The second value refers to Stockholm design conditions.b For each system the three swept volumes refers to Treviso, Stockholm and Singapore respectively.

L. Cecchinato et al. / Applied Thermal Engineering 30 (2010) 1946e1958 1951

For a set refrigerating unit, the numerical procedure can beoutlined with the following steps:

A. Determination of design heat exchangers transmittancesA.1 input data gathering;A.2 calculation of design condensation and evaporation

temperatures;A.3 determination of design heat exchangers transmittances.

B. Determination of system performance under different load andsecondary fluid temperature operating conditions

Table 6Design conditions for liquid chiller/heat pump.

System Chiller/HP 2 (air/water)

Tc Tc,sf,in Te Te,sf,in _Vb

[�C] [3600 m3

0 50 40 �10 �3 180207413

1 50 40 �10 �3 180207413

2 50 35/40a 2/�10a 12/�3a 5453

100

3 50 35/40a 2/�10a 12/�3a 6564

120

4 50 35/40a 2/�10a 12/�3a 9088

166

5 50 35/40a 2/�10a 12/�3a 5857

106

a The second value refers to Stockholm design conditions.b For each system the three swept volumes refers to Treviso, Stockholm and Singapor

B.1 calculation of cooling/heating capacity and powerconsumption based on corrected transmittances as a func-tion of the actual refrigerant and secondary fluid flow ratesand of the logic of the unit control system;

B.2 part load performance estimation.

“B” steps should be iterated for each different working conditionto be considered. The numerical procedure steps are detailed in thefollowing.

Chiller/HP 1 (water/water)

Tc Tc,sf,in Te Te,sf,in _Vb

s�1] [�C] [3600 m3 s�1]

e

e

e

e

50 40 5 15 867392

50 40 5 15 837088

e respectively.

Enthalpy [103 J kg-1]

3

4

2

1

2d 2s

Pres

sure

[105 P

a]

Fig. 9. Single stage vapour compression cycle in peh diagram.

Table 7Design conditions for heat exchangers SR1 and SR2.

System SR1 SR2

Tr,in Tsf,in DTsf DTapproach TLT,in TLT,out TMT,in TMT,out

[�C] [K] [�C]

1 e 45/37a �7 �10 �52 45/37a 7 5 3 50 �7 �10 �53 45/37a 7 5 3 e

4 45/37a 7 5 3 20 �7 �10 �55 45/37a 7 5 3 e

a The second value refers to Stockholm design conditions.

L. Cecchinato et al. / Applied Thermal Engineering 30 (2010) 1946e19581952

A.1. Input data gathering

The following data are required for the model application toa given refrigeration unit:

� the refrigerant fluid and the secondary fluids at evaporator andcondenser;

� for a refrigerating unit, the rated cooling capacity, Qe,D, andcompressor power consumption, Pk,D (or design efficiency, EER)in design conditions;

� for a heating unit, the rated heating capacity, Qc,D, andcompressor power consumption, Pk,D (or design efficiency,COP) in design conditions;

� design refrigerant evaporator outlet superheat and condenseroutlet subcooling values, DTSH,D and DTSC,D;

� compressor coefficients with reference to the EN 12900 [23]polynomial equations;

� secondary fluids design inlet temperature, temperature varia-tions between inlet and outlet and pressure drops, Tsf,D, DTsf,D(or in alternative _msf ;D)

� heat exchanger pumps or fans design power consumption,Paux;D;

� heat exchangers design ratio of refrigerant side to secondaryfluid side heat transfer area;

� heat exchangers design ratio of refrigerant to secondary fluidheat transfer coefficient with reference to the considered heattransfer surface area;

� refrigerant and secondary fluid circuits pattern;� compressors sequencing and inverter management for multi-compressor and inverter driven unit respectively, with refer-ence to cooling capacity control.

Refrigerant pressure drops in the heat exchangers data are notrequired since they are not easily available in manufacturers datasheets. Heat exchangers heat transfer surface areas and heattransfer coefficients, together with condenser air inlet-outlettemperature difference (or mass flow rate), are seldom present indata sheets; nevertheless they can be quite easily obtained from themanufacturer or from open literature correlations or heatexchangers design practice.

A.2. Calculation of design condensation and evaporationtemperatures

Considering the single stage vapour compression cycle shown inFig. 9 peh diagram, the thermodynamic properties of points 1, 2s, 3and 4 can now be determined at design conditions. The processundergone by the refrigerant inside the compressor is described byan adiabatic compression (1e2) and an isobaric process (2e2d),which accounts for heat transfer phenomena between thecompressor and the environment. The thermodynamic propertiesassociated to points 2 and 2d depend on the compressor charac-teristics. In fact, 2s is the discharge point corresponding to an

isentropic compression process, while 2 and 2d refer respectively tothe end of an adiabatic compression process and to the compressordischarge, taking into account both the compression irreversibilityand compressor heat dissipation. With reference to EN 12900superheat and subcooling standard values [23], it is possible toexpress the cooling capacity, the power consumption and therefrigerantmass flow ratewith the following polynomial equations:

bQe ¼ a1 þ a2$Te þ a3$Tc þ a4$T2e þ a5$Te$Tc þ a6$T

2c þ/

þ a7$T3e þ a8$T

2e $Tc þ a9$Te$T

2c þ a10$T

3c ; (4)

bPk ¼ b1 þ b2$Te þ b3$Tc þ b4$T2e þ b5$Te$Tc þ b6$T

2c þ/

þ b7$T3e þ b8$T

2e $Tc þ b9$Te$T

2c þ b10$T

3c ; (5)

b_m ¼bQebh1 � bh4

$ (6)

Considering equations (4)e(6) and the EN 12900 [23], superheatand subcooling values, the compressor volumetric and isentropicefficiencies can be thus easily expressed as:

hv ¼b_mbr1$ _V ; (7)

his ¼b_m$

�bh2s � bh1

�bPk

: (8)

In the hypothesis that these efficiencies values don’t depend onthe refrigerant superheat at the compressor suction, the refrigerantmass flow rate, the cooling capacity and the power consumptioncan be calculated with reference to the Fig. 9 cycle:

_m ¼ hv$_V$r1; (9)

Qe ¼ _m$ðh1 � h4Þ; (10)

Pk ¼ _m$ðh2 � h1Þ; (11)

h2, being the refrigerant enthalpy at the end of the adiabaticcompression process, determined by:

h2 ¼ h1 þh2s � h1

his: (12)

The compressor discharge enthalpy, h2d, can be defined asa function of the compressor heat dissipation factor, hd:

h2d ¼ h2 �Pk$hd

_m: (13)

L. Cecchinato et al. / Applied Thermal Engineering 30 (2010) 1946e1958 1953

Through equations (12) and (13), the thermal power to bedissipated in the condenser can be calculated as:

Qc ¼ _m$ðh2d� h4Þ ¼ _m$

��h1 þ

h2s� h1his

� Pk$hd_m

�h4

: (14)

The vapour compression system energy efficiency is obtainedwith the following two expressions, for a refrigerating and a heat-ing unit respectively:

EER ¼ Qe

Pk þP

Paux;D; (15)

COP ¼ Qc

Pk þP

Paux;D$ (16)

Combining equations (4)e(16), the vapour compression cycledesign operating temperatures, Tc;D and Te;D, can be determinedsolving the following equations system:�Q�Te;D; Tc;D

�� QD ¼ 0Pk�Te;D; Tc;D

�� Pk;D ¼ 0 ; (17)

where Q represent the cooling or the heating capacity for a refrig-erating and a heating unit respectively.

A.3. Determination of design heat exchangers transmittances

The evaporator and condenser transmittances in design condi-tions can be subsequently determined with a logarithmic meantemperature difference approach. If the heat exchanger heattransfer surface is divided into M partitions of equal refrigerantenthalpy difference, its transmittance can be calculated with thefollowing expression in the co-current hypothesis:

UA ¼X

j¼1;M

_m$�hj � hj�1

�qj

; (18)

where qj is defined as the logarithmic mean temperature differencebetween the refrigerant and the secondary fluid with reference tothe j-th subdivision of the refrigerant enthalpy difference. It can becalculated solving the following algebraic system:8>>>>>><>>>>>>:

8><>:hj ¼ hjþ1�

hout�hinM

Tsf ;j ¼ Tsf ;jþ1�_m$ðhout�hinÞM$ _msf$cp;sf

; j¼ 0;M�1�hj ¼ houtTsf ;j ¼ Tsf ;in

; j¼M

$ (19)

For a given refrigerating unit, it is thus possible to calculate thedesign cycle operative temperature together with the heatexchangers transmittance, UAe;D and UAc;D, through equations(4)e(19).

For each heat exchanger secondary fluid, the volumetric flowrate is calculated on the basis of the following equation, neglectingmass transfer phenomena:

_Vsf ¼ Qrsf$cp;sf$DTsf

; (20)

where Q is the heat exchanger thermal power.

B.1. Calculation of cooling/heating capacity and power consumption

For secondary fluids temperature values different from designconditions, it is possible to determine the refrigeration machineworking conditions and energy efficiency based on design data inthe following hypotheses:

� constant design evaporator outlet superheat and condenseroutlet subcooling values;

� neglected refrigerant pressure drops influence;� constant design heat exchangers secondary fluids volumetricflow rates and pressure drops;

� refrigerant and secondary fluids heat transfer coefficient, a,proportional to the m-th power of Reynolds number: afRem

The vapour compression cycle operating temperatures, Tc andTe, determination problem can be reduced to the solution of thefollowing non-linear equations system:

�UAeðTe; TcÞ � UAe;D$ke ¼ 0UAcðTe; TcÞ � UAc;D$kc ¼ 0 ; (21)

where the heat exchanger transmittances, UA, are obtained fromequations (4)e(12), (14), (18) and (19) that represent the basicequations system of a refrigerating unit for given values of evapo-ration and condensation temperatures. Instead, the condenser andevaporator transmittance correction coefficients, k, are calculatedwith the following expression:

k ¼1þ

�asfar

D$

�AsfAr

D�

_m_mD

�mfs

sfþ�asfar

D$

�AsfAr

D$

�_m_mD

�mr

r$

(22)

where Ar and Asf are the heat transfer surface on the refrigerant andsecondary fluid side respectively and ar and asf are the corre-sponding heat transfer coefficients.

The system of equations is solved with a standard solver basedon the MINPACK subroutine HYBRD1, which uses a modification ofPowell’s hybrid algorithm. This algorithm is a variation of Newton’smethod, which implements a finite-difference approximation tothe Jacobian and takes precautions to avoid large step sizes orincreasing residuals [24]. The system power consumption, coolingand heating capacity are calculated from system (21) solutionsusing equations (11), (10) and (14). The system energy efficiency isobtained according to equations (15) and (16). Refrigerant proper-ties were evaluated with NIST Reference Fluid Thermodynamic andTransport Properties e REFPROP, Version 7.0 [25].

In presence of systems with cooling or heating capacity control,such as multi-compressor scroll chillers, multi-step screwcompressormachines or inverter driven units, equations (4) and (5)can be rewritten as a function of the number of active compressorsand of the compressor capacity partialisation rate.

If the refrigeration unit has multiple refrigerant or secondaryfluids circuits which can be activated as a function of the actual unitworking conditions and active compressors, equation (22),expressing the heat exchangers design transmittance correctionfactor, is corrected. The auxiliaries design power consumptionmustalso be corrected as a function of the secondary fluids activecircuits. Further details and equations development can be found inCecchinato et al. [26].

Usually air condensed units for civil and commercial refrigera-tion are equipped with condensation temperature control systemsto assure safe working conditions for the compressor and theexpansion valve. These systems regulate the condenser air volu-metric flow rate in order to keep or the condensation temperatureabove a minim value, Tc,min, or the pressure difference at the twoends of the throttling valve above a minim value, Dpc,min. In thiscase, the system (21) must be replaced with the following con-strained equations system:

Fig. 10. Annual energy consumption for the analysed systems.

L. Cecchinato et al. / Applied Thermal Engineering 30 (2010) 1946e19581954

8>>>><>>>>:

fe�Te; _Vsf ;c

�¼ UAe � UAe;D$ke ¼ 0

fc�Te; _Vsf ;c

�¼ UAc � UAc;D$kc ¼ 0

Tc ¼ Tc;min or Tc ¼ T�pe þ Dpc;min

� : (23)

Table 8Monthly loads and energy consumption for the analysed systems in Treviso climate for

January February March April May

Load [kWh]MT 27,639 25,431 29,228 27,152 28,723LT 16,272 14,823 16,665 15,899 16,592AHU e heating 21,030 16,135 11,894 21 0AHU e cooling 299 528 1942 2898 5946

Energy consumption [kWh]System 0 MT 6345 5890 7065 6989 8257

LT 8067 7384 8519 8466 9568HP(air/water) 6656 5076 3611 1 0Chiller(air/water) 15 29 195 544 1207HP(water/water) e e e e e

System 1 MT 7272 6738 8074 8066 9657LT 6066 5556 6380 6288 6884HP(air/water) 6656 5076 3611 1 0Chiller(air/water) 15 29 195 544 1207HP(water/water) e e e e e

System 2 MT 9748 8473 8829 6472 7518LT 6059 5419 6135 5942 6968HP(air/water) 794 563 294 0 0Chiller(air/water) 450 471 871 1456 2520HP(water/water) e e e e e

System 3 MT 13,423 11,761 12,473 9757 11,314LT 4162 3795 4276 4077 4265HP(air/water) 847 607 323 0 0Chiller(air/water) 464 487 883 1453 2489HP(water/water) e e e e e

System 4 MT 6290 5774 6774 6471 7518LT 5411 4917 5756 5942 6968HP(air/water) 848 607 326 0 0Chiller(air/water) 447 469 869 1456 2520HP(water/water) 3700 2877 2184 0 0

System 5 MT 9196 8456 9950 9757 11,314LT 4162 3795 4276 4077 4265HP(air/water) 929 675 374 0 0Chiller(air/water) 460 483 880 1453 2489HP(water/water) 3647 2832 2153 0 0

Usually commercial refrigeration plants are equipped witha suction pressure control system to assure stable evaporationtemperature to the cabinets. Parallel compressors are relaycontrolled in order to keep the evaporation pressure set point value,pe,set. For a given number of active compressors, the system (21)must be replaced with the following constrained equations system:

constant set point supply temperature.

June July August September October November December

29,518 32,350 31,784 28,374 30,484 28,051 26,97416,472 17,456 17,320 16,197 16,952 16,068 16,148

0 0 0 0 10,956 13,320 22,6468909 14,128 13,470 8058 3463 1014 304

9341 11,330 10,974 8896 8016 6719 620010,225 11,881 11,592 9969 9181 8175 8008

0 0 0 0 2891 3877 72831915 3295 3080 1746 314 54 12

e e e e e e e

11,041 13,497 13,085 10,539 9200 7672 71217095 7837 7727 6941 6767 6144 6018

0 0 0 0 2891 3877 72831915 3295 3080 1746 314 54 12

e e e e e e e

8268 9593 9367 7889 9444 8768 98537354 8108 8031 7176 6831 5916 6180

0 0 0 0 110 342 9493607 5765 5371 3357 1284 675 424

e e e e e e e

12,341 14,243 13,930 11,852 13,409 12,349 13,5274249 4520 4481 4172 4364 4123 4127

0 0 0 0 122 365 10173566 5735 5331 3318 1277 691 432

e e e e e e e

8268 9593 9367 7889 7536 6469 61997354 8108 8031 7176 6478 5492 5489

0 0 0 0 133 378 10123607 5765 5371 3357 1282 673 421

0 0 0 0 2034 2437 3931

12,341 14,243 13,930 11 852 11,059 9515 90654249 4520 4481 4172 4364 4123 4127

0 0 0 0 158 419 11133566 5735 5331 3318 1273 687 428

0 0 0 0 2010 2402 3872

L. Cecchinato et al. / Applied Thermal Engineering 30 (2010) 1946e1958 1955

�fcðTe; TcÞ ¼ UAc � UAc;D$kc ¼ 0pe ¼ pe;set

: (24)

B.2. Part load performance estimation

The effect of the part load working of the refrigeration unit iscalculated on the basis of its full load performance and strictlydepends on its capacity control system. For single compressor unitswith simple relay control law, the part load factor, PLF, can becalculated with the equations proposed in the European and Italianstandards [27,28] as a function of the part load ratio, PLR:

PLF ¼ PLRCc$PLR þ ð1� CcÞ; (25)

where Cc is a degradation factor associated to the unit cyclingoperation losses.

The part load ratio is calculated as the ratio between the actualload condition, Qload, and the unit full load cooling or heatingcapacity considering the of operative secondary fluids temperaturevalues. For a refrigerating or a heating system, the part load powerconsumption, cooling or heating capacity, Pk;PL, Qe;PL or Qc;PL, can bethus defined as a function of full load performance:

Pk;PL ¼ PLF$Pk; Qe;PL ¼ PLR$Qe or Qc;PL ¼ PLR$Qc: (26)

The systemenergy efficiency is subsequently obtained accordingto equations (15)e(16). In refrigerating units with multiplecompressors or with compressors with different capacity steps,a step capacity control is usually adopted. In this case the full loadcapacity of each step is determined. It is now possible to determinethe two steps which mostly approximate the actual load condition,

Table 9Monthly energy consumption for the analysed systems in Treviso climate for floating se

January February March April May

Energy consumption [kWh]System 0 MT 6345 5890 7065 6989 8257

LT 8067 7384 8519 8466 9568HP(air/water) 5833 4428 3119 1 0Chiller(air/water) 12 24 165 459 1025HP(water/water) e e e e e

System 1 MT 7272 6738 8074 8066 9657LT 6066 5556 6380 6288 6884HP(air/water) 5833 4428 3119 1 0Chiller(air/water) 12 24 165 459 1025HP(water/water) e e e e e

System 2 MT 9301 8226 8943 7360 8492LT 5944 5326 6063 5949 6978HP(air/water) 814 584 321 0 0Chiller(air/water) 125 143 389 798 1672HP(water/water) e e e e e

System 3 MT 12,811 11,382 12,486 10,667 12,299LT 4162 3795 4276 4077 4265HP(air/water) 869 630 353 0 0Chiller(air/water) 129 147 387 788 1644HP(water/water) e e e e e

System 4 MT 6876 6354 7539 7360 8492LT 5424 4926 5764 5948 6978HP(air/water) 833 597 326 0 0Chiller(air/water) 126 144 389 798 1672HP(water/water) 3118 2408 1812 0 0

System 5 MT 9799 9054 10 733 10 666 12,299LT 4162 3795 4276 4077 4265HP(air/water) 909 659 372 0 0Chiller(air/water) 129 147 388 788 1644HP(water/water) 3075 2373 1787 0 0

the first onewill have a lower capacity, the second one a higher one.The unit part load capacity, power consumption and efficiency aredetermined by linear interpolation between these two stepsperformances [27]. If the unit smallest control step cooling capacityis higher than the required one, the performance data are obtainedas detailed for single compressor units, referring the PLR value inequation (25) to the unit smallest control step cooling capacity.

Finally, the seasonal energy consumption of the proposedsystems has been based on the dynamic building energy loadresults, as described in Section 3; the average performance of theplant has been calculated with a steady-state approach for eachbuilding simulation time step (1 h), as proposed in the standard UNI11135 (UNI 2003) [29].

6. Simulation results

The total annual energy consumption of refrigeration and air-conditioning systems, related to the overall supermarket area, isrepresented in Fig. 10 for the considered plant solutions andclimates. Treviso monthly results are reported in Tables 8 and 9.With reference to System 0 in the case of constant water deliverytemperature (constant set point), System 4 presents the largestenergy saving (14.8, 20.5 and 12.9% in Treviso, Stockholm andSingapore respectively), thanks to the integration between theliquid chiller/heat pump and the refrigeration units. The water/water HP of System 4, which uses the recovered heat as heat source,satisfies 89.4, 86.8 and 100% of the heating energy demand in thethree climates. System 2, which directly recovers the refrigerationunits condensation heat, is slightly penalised by the high conden-sation temperature in wintertime. But, if the water deliverytemperature can vary (floating set point) according to the heating/

t point supply temperature.

June July August September October November December

9341 11,330 10,974 8896 8016 6719 620010 225 11 881 11,592 9969 9181 8175 8008

0 0 0 0 2460 3343 64381642 2910 2710 1500 265 46 10

e e e e e e e

11,041 13,497 13,085 10,539 9200 7672 71217095 7837 7727 6941 6767 6144 6018

0 0 0 0 2460 3343 64381642 2910 2710 1500 265 46 10

e e e e e e e

9229 10 525 10,300 8816 9688 8748 93717364 8117 8040 7185 6757 5832 6062

0 0 0 0 140 367 9752642 4657 4277 2435 666 232 121

e e e e e e e

13,311 15,190 14,877 12,787 13,548 12,208 12,8704249 4520 4481 4172 4364 4123 4127

0 0 0 0 156 391 10472606 4627 4240 2400 652 234 122

e e e e e e e

9229 10,525 10,300 8816 8413 7184 67627364 8117 8040 7185 6486 5500 5501

0 0 0 0 140 374 9992642 4657 4277 2435 667 233 121

0 0 0 0 1667 2019 3338

13,311 15,190 14,877 12,787 11,937 10,250 96374249 4520 4481 4172 4364 4123 4127

0 0 0 0 166 411 10952606 4627 4240 2400 652 234 123

0 0 0 0 1654 1995 3288

Fig. 11. Annual energy consumption for each unit of the analysed systems in Treviso climate.

L. Cecchinato et al. / Applied Thermal Engineering 30 (2010) 1946e19581956

cooling demand, System 2 operates at lower average wintercondensation temperature, thus offering the best result, with 15.6,22.5 and 14.9% energy saving in comparison to System 0 (constantset point) in Treviso, Stockholm and Singapore respectively. Theenergy efficiency improvement depends on the climatic conditionsand is higher for colder cities. This is related to the higher heatrecovery during the heating season.

In general, Systems 2 and 4 show very similar performances; theplant results much simpler in the case of System 2, with lowercomplexity in the piping integration system and with only one heatpump/chiller.

System 3, which has the same integration level between theHVAC unit and the MT refrigeration unit of System 2, while the LTrefrigeration unit rejects the condensation heat to the MT refrig-eration system in a typical cascade configuration, does not performas well as System 2, because of the lower cycle efficiency of thecascade system with respect to the LT and MT separate cycles. Infact, in Treviso design conditions the COP of the System 2 LT unit is18.5% higher than System 3 one considering the cascade arrange-ment. For the same reasons, System 5 shows poorer performancesthan System 4. The energy penalisation of these systems is higherin Treviso than in Stockholm and Singapore. In fact, the energypenalisation of cascade lay-out with respect to single stage one islower for high compression pressure ratio. This is the case of

Fig. 12. Annual energy consumption for each unit

Stockholm plant operating winter heat recovery for a long timeperiod and of Singapore one operating at high dry-bulb tempera-ture during almost all the year.

The energy saving of the proposed solutions is at least 6.2, 5.9and 7.4%, in the case of System 1 with constant water set point forTreviso, Stockholm and Singapore respectively. The subcooling ofthe LT unit by the MT plant is more effective in the hotterclimates.

Figs. 11e13 represent the annual energy consumption related tothe overall supermarket area, of each unit of the analysed systems(MT refrigeration, LT refrigeration, liquid chiller, heat pump air/water, heat pumpwater/water), in the case of constant temperatureof water delivery for Treviso, Stockholm and Singapore. Stockholmsupplementary condensing boiler primary energy consumptionwas converted to electricity with a 2.5 conversion factor; for eachsystem the boiler annual consumptionwas below 0.3 kWhm�2 andit was not reported in Fig. 12.

During wintertime, the integration between the refrigerationunits and the HVAC system, resulting from the recovery of thecondensation heat, leads to a very low energy consumption for theair/water heat pump. During summertime, Systems 2e5 showhigher energy consumption for liquid chillers than that of Systems0e1, because of the added heat load coming from the subcoolingneeds of the refrigeration units.

of the analysed systems in Stockholm climate.

Fig. 13. Annual energy consumption for each unit of the analysed systems in Singapore climate.

L. Cecchinato et al. / Applied Thermal Engineering 30 (2010) 1946e1958 1957

As for the MT refrigeration unit is concerned:

a) the energy consumption of System 2 is higher than that ofSystem 4 in Treviso and Stockholm, because of highercondensation temperature associated to winter heat recovery;

b) Systems 3 and 5 show the highest energy consumption becausethey manage, as an additional load, the condensation heat ofthe LT R744 units. System 1 is also penalised in Singaporebecause of the higher load associated to LT unit subcooling.

7. Conclusions

Amedium-size supermarket was considered, with respect to therefrigerating machinery for commercial refrigeration and air-conditioning/heating system; different levels of integrationbetween the HVAC and the refrigeration units was introduced.A mathematical model providing an accurate estimation of refrig-eration and air-conditioning units cooling or heating capacity andpower consumption out of nominal rating conditions has beendeveloped. The proposed procedure corrects the design trans-mittance taking into account the heat transfers coefficients varia-tions associated to refrigerant and secondary fluid mass flow ratesvariations.

Calculations show that all the proposed solutions offer a benefitin terms of energy efficiency over the reference traditional solution.The integration between the refrigeration units and the HVACsystem and their optimal management can produce energy savingup to 15.6, 22.5 and 14.9% in comparison with the traditionalsolution in Treviso, Stockholm and Singapore respectively. The bestperformance is achieved in the colder climates and its associated towintertime condensation total heat recovery from the refrigerationunits. The simplest solution, operating the LT unit subcooling bymeans of the MT plant, allows energy efficiency improvementsequal to 6.2%, 5.9 and 7.4% for Treviso, Stockholm and Singaporerespectively. This solution is suitable for hot climates. LT cascadelay-out is not to be adopted in temperate climates (Treviso).

As widely demonstrated in the technical literature, simulationsconfirm that floating water set point is effective for energy saving.

As a final remark, the integration approach should be proposedalso for large supermarkets and hypermarkets, in consideration oftheir significant energy consumption.

Acknowledgments

This research work is part of the “Supermercato in classeA (Class A Supermarket)” project, carried out by Distretto Venetodel Condizionamento e della Refrigerazione Industriale (District ofVenetian Region for AC and Industrial Refrigeration) and financedby Regione Veneto (Venetian Region) with legislative measureDDSE n. 258 on 21/12/2007. The following companies, managed bythe above mentioned district, were collaborating on the project :Bluebox Group Srl, Carel SpA, Climaveneta SpA, DGD Srl, Elet-tromeccanica SpA, Enofrigo Srl, Evco Srl, Ever-est Srl, G. R. Elet-tronica Srl, Guerrato SpA, La Felsinea Srl, Rhoss SpA, SCM Frigo Srl,Sirman SpA, Studio 54 Srl, Uniflair SpA, Zoppellaro Srl.

Finally we would like to thank Dr. Alessio Gastaldello and MrSergio Girotto for the management of the project.

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