im

ai, C

andel wandCOPr coup.betd ac

2011 Elsevier Ltd. All rights reserved.

in buiith mimpoeater(ASH

d enerer hea

method based on experimental results. Their research indicatedthe effect of ambient temperature (surrounding air temperature)on the system performance is more signicantly than the effectof initial water temperature during the heating process. Rankinet al. [16] presented a study about demand side management forcommercial building using an inline air-source heat pump waterheater methodology. Kim et al. [17] suggested that the tank size

and timing control patterns are two commonly control patternsadopted by the consumers. Thus, the operation strategy optimiza-tion should be considered according to different control patterns.Whereas the potential of such operation strategy optimization isfar from being developed, there is lack of information about theoperation strategy optimization of ASHPWH systems, especiallyin the residential buildings.

In this paper, structural and operation strategy optimizationof an experimental set-up of air-source heat pump water heater(ASHPWH) system were investigated. Based on a simulation

Corresponding author. Tel.: +86 21 3420 6776; fax: +86 21 3420 6309.

Applied Energy 88 (2011) 41284138

Contents lists availab

lseE-mail address: jywu@sjtu.edu.cn (J.Y. Wu).and more interests for its high efciency, energy-saving and envi-ronmental benign recently [3,4].

The thermal performance of the ASHPWH system is inuencedby refrigerants, system structure, ambient temperature, watertemperature, etc [512]. And many achievements have been ob-tained by the researchers. Mei et al. [13] investigated the differenceof R22 and R407C in the practical ASHPWH system. Afterward,they studied the effect of different style natural convection im-mersed condensers to the heat pump water heater [14]. Morrisonet al. [15] presented an ASHPWH seasonal performance assessment

water temperature for the system, satisfactory thermal and eco-nomic performance can be also obtained [19,20]. Rankin et al.[16] solved the peak demand of national electricity supply gridby controlling the operation of inline air-source heat pump waterheater and auxiliary heater simultaneously. Huang et al. [21] pre-sented a fast response heat pump water heater by controlling dif-ferent water temperatures of two tanks, and the key of itsoperation strategy is to setting different priority for two tanks. Asdiscussed above, the operation strategy is greatly subject to theconsumers control patterns, in practical application, thermostatic1. Introduction

The proportion of water-heatingtion is about one-third. Nowadays, wbuilding energy-saving, considerableon the energy efciency of water hair-source heat pump water heaterheating device with about one-thirconventional electric resistance wat0306-2619/$ - see front matter 2011 Elsevier Ltd. Adoi:10.1016/j.apenergy.2011.04.012lding energy consump-ore concern is taken tortance has been placedin buildings [1,2]. ThePWH) system, a watergy consumption of theter, has attracted more

should be optimized through simulation. In Japan, the scholars[18] had studied the inuence of ambient temperatures to thewhole system, and pointed out that operation strategy of ASHPWHsystem is greatly important and should be considered in their sub-sequent work.

However, due to the variation of the ambience and heating load,ASHPWH systems are only optimal under certain working condi-tions if the structure is xed. Besides, if suitable operation strategyis established, such as selecting correct start-up time or settingOptimization that, based on this set-up, setting water temperature should be set higher than 46 C in summer and 50 Cin other seasons.Experimental research and operation optwater heater

J.J. Guo, J.Y. Wu , R.Z. Wang, S. LiInstitute of Refrigeration and Cryogenics, Shanghai Jiao Tong University, 200240 Shangh

a r t i c l e i n f o

Article history:Received 18 October 2010Received in revised form 17 March 2011Accepted 2 April 2011Available online 27 April 2011

Keywords:Air-source heat pump water heaterExperimental research

a b s t r a c t

In order to optimize designset-up and simulation modinto account thermostaticindicated that the averageside area ratio of condense6.0 and 6.5 m2 for this set-electricity price differenceperature should be adjuste

Applied

journal homepage: www.ell rights reserved.ization of an air-source heat pump

hina

operation strategy of air-source heat pump water heater, an experimentalere constructed. Also, a methodology of optimizing operation, which takestiming control patterns, was presented and applied. Experimental resultsranged from 2.82 to 5.51 under typical conditions. The recommended out-il to evaporator is 0.140.31 when the evaporator outside area is betweenThe optimal start-up time was between 12:00 and 14:00 if there was noween day and night, or it was near 22:00. The optimal setting water tem-cording to the variation of seasonal ambient temperature. It was suggested

le at ScienceDirect

Energy

vier .com/ locate/apenergy

p pitch, mm

ergy_Q heat transfer rate, kWr latent heat, kJ/kgs revolutional slipt time, sT temperature, CU heat transfer coefcient, kW/(m2 K)v specic volume, m3/kgV volume, LNomenclature

A heat exchange area, m2

C specic heat, KJ/(kg K)COP coefcient of performanced diameter of copper tube, mf operation frequency of compressor, Hzf(t) domestic hot water demand prole, kg/sg(t) hot water ow rate from water tank, kg/sh specic enthalpy, KJ/kgj number order of time stepk1, k2 modifying factor of thermostatic expansion valve_m mass ow rate of refrigerant, kg/sM mass, kgn polytropic indexP pressure, Pa

J.J. Guo et al. / Applied Enmodel, the relationship between system thermal performanceand outside area ratio of condenser coil to evaporator was dis-cussed. Additionally, a methodology for the operation strategyoptimization of the system was introduced and applied, andthe optimal start-up time and setting water temperature wereobtained based on timing and thermostatic control patterns,respectively.

2. Description of system and experiment set-up

2.1. Working process of the ASHPWH system

The system is composed of outdoor and indoor units, as shownin Fig. 1. The outdoor unit, such as compressor, lterdrier, ther-mostatic expansion valve (TXV) and the evaporator, is mountedin the environmental chamber. The other parts of the ASHPWHsystem, for example water tank and condenser, is installed in theambient out of the environmental chamber. The condenser is im-mersed in the hot water tank. The air temperature in the environ-mental chamber can be adjusted from 5 C to 40 C.

The working process undergone by the refrigerant during acycle can be represented by an idealized heat pump cycle as shownin Fig. 2. Here, 12, 23, 34 and 41 represent compression, con-densation, throttling and evaporation processes, respectively. Afterthe throttling process, the refrigerant in the evaporator will absorbthe heat from the ambient air, then after the compression process,the heat will be released into the water tank. As the water tank

Vd compressor displacement volume rate, cm3/revW total input power, kWW(t) electrical power consumption of compressor and fan,

kW

Subscriptsa ambient; aircmp compressorcond condensation or condenserconsume consumed watereva evaporation or evaporatorSuperscriptsDt testing time step between the initial and end time0 testing initial timen ni inside; inletl refrigerant liquidload hot water loadlon longitudinal tubeo outsidep constant pressurer refrigerants sensible heat transferset setting water temperaturesc subcooledsh superheatedtap tap watertp two-phasetra transverse tubetube heat transfer tubev volumetric; vaporw waterwi initial water

88 (2011) 41284138 4129getting warmer, the working process cycle will be updated accord-ingly, such as 3040102030.

2.2. Experimental facility and procedures

An experimental set-up of ASHPWH system was developed inthe lab. Fig. 3 and Table 1 show the details of indoor and outdoorunits.

In this experiment, the evaporator, TXV and the compressorwere mounted in an environmental chamber, as shown in Fig. 3aand b. The rated input power of the rotary-type hermetic compres-sor was 0.93 kW. To avoid the overload, an overheated protectorand low-high pressure cut-off switches were connected to thecompressor. The evaporator was a n-tube heat exchanger. Thecondenser was made up of a smooth copper helical coiled tube(9.90 0.75), which was immersed in the water tank. The curva-ture radius of the coil was 0.16 m. The lterdrier and TXV wereinstalled downstream the condenser. The whole system was con-trolled by a micro-controller.

Before the experiment, the environmental chamber was ad-justed at certain value (it represents ambient temperature), andthe tap water was charged into the 150 L water tank. Then, thewater was heated after the ASHPWH system started. If the watertemperature in the tank reaches to the setting point, the systemis stopped by the micro-controller. In the process of hot water con-suming, the hot water was drawn from the top of the tank, and coldtap water was recharged from the bottom of the tank.

Greek letterd n thickness of evaporator, mmg efciencya heat transfer coefcient, kW/(m2 K)k thermal conductivity, kW/(m K)n dehumidifying coefcientq density, kg/m3

gn general n efciencyf delivery coefcientDs duration of one time step in numerical calculation

wcondenser

tap water

TemperaturePosition

PressurePosition

PT

P

T

TT

P

m of the ASHPWH system.

ergycompressor

TXV filter-drier

environmentalchamber

evaporator

T

T

T

TP

T P

Fig. 1. Schematic diagra

4130 J.J. Guo et al. / Applied EnThe main aim of the experimental work is to test the systemthermodynamic performance, such as the COP and the averageCOP.

COP Cp;wMwTDtw T0wR Dt

0 Wtdt1

As shown in Eq. (1), the COP is the ratio of heat transfer from thecondenser to water to the electric power consumption of compres-sor and fan. The tank water temperature (i.e. TDtw or T

0w) is repre-

sented by the average value of the three platinum resistancethermometers (i.e. Pt100) set at the bottom, central and top of thetank, as shown in Fig. 1. 0 and Dt represent the testing initial timeand time step between the initial and end time, in which there is3 C increment in water temperature. The heating capacity is thetime average heat transfer from the condenser to water within3 C increment in the water tank.

R Dt0 Wtdt represents the electri-

cal power consumption of compressor and fan within time stepmeasured by power meter unit, kJ. Eq. (1) represents COP at watertemperature T

Dtw T0w2 .

When the water in the tank is heated from certain initial tem-perature (i.e. Twi) to the setting point (i.e. Tset), the average COPduring the whole process is

COPave Cp;wMwTset TwiR t0 Wtdt

2

1

23

4

Heat to thewater tank

Ambient energy

logP

h

1'

2'3'

4'

Fig. 2. Heat pump cycle on a logp h diagram.ater tank hot water

T

88 (2011) 41284138where t represents the duration of the whole process.R t0 Wtdt rep-

resents the electrical power consumption of compressor and fanmeasured by power meter unit during the whole process.

2.3. Error analysis

The power meter unit (with grade 0.5 accuracy, uncertaintyabout 10W) was used to measure the electrical power

Fig. 3. Experimental set-up of ASHPWH system: (a) indoor unit; (b) outdoor unit.

ash,eva represent the heat transfer coefcient of the two-phase sec-

entcoi

ace a0.5

ergySince the compression of refrigerant vapor is assumed to be apolytropic process, the energy equation for the compressor canbe expressed as:

_mrh2 h1 _mr Peva v1 nn 1PcondPeva

n1n

1" #

3

where n is the polytropic index.The mass of refrigerant ( _mr) pumped by the compressor is cal-

culated by [23],

_ _f 1 sVdgvconsumption of the compressor and fan automatically. The threeplatinum resistance thermometers (i.e. Pt100, with grade A,uncertainty of 0.2 C) inserted along the axis of the tank, wereset at the bottom, central and top of the tank. All the data werecollected by a data acquisition system and stored in a computer.The interval of data collection can be set fteen seconds, and theerror of the data acquisition system can be ignored. Therefore,the error analysis of the experimental results on the basis ofthe uncertainties in the measurements was performed usingthe Kline and McClintock relationship [22], and the relative errorof testing instrument was 9.4% for COP, 0.8% for average COP,respectively.

3. Mathematical model and simulation process

In order to predict the operation performance of ASHPWH sys-tem and nd the inuence of various parameters on its operation, amathematical model has been formulated in this paper. Somephysical assumptions employed for the present model are:

(1) Compression of refrigerant vapor is assumed to follow apolytropic process.

(2) Pressure drop is negligible in evaporator, condenser as wellas pipes.

(3) Expansion of refrigerant liquid is considered to beisenthalpic.

3.1. Compressor model

Table 1Specication of the main components of the system.

Name Type Remarks

Compressor Rotary R22, Rated input power:0.93KW, displacemCondenser/water

tankPressureresistance

150 L tank, immersed 50 m smooth copper

Expansion valve TXV Interior equilibrium typeEvaporator/fan Finned tube/

axialAxial fan rated input power: 35 W; front fptra:25.1 mm, smooth copper tube (9.52

J.J. Guo et al. / Applied Enmr mcmp v1 4

where Vd is the displacement volume, s is the revolutional slip andgv is the volumetric efciency.

3.2. Evaporator model

The physical components and working environments of theevaporator in the ASHPWH system, though the working conditionin summer is formidable, is similar to the normal air-source uni-tary heat pump. So the evaporator model of the system can be pre-sented as follow.

The energy equation for the evaporator can be expressed astion and superheated section at the inside of evaporator, whichare expressed in [25].

3.3. Condenser model

The condenser energy balance equation can be expressed as:

_Qcond _mrh2h3Cr;sh;cond _mrT2Tcond _mrrcondUsc;cond Ai;condCr;sh;cond

_mrUsh;cond

lnTcondTwT2Tw

_mrrcondUtp;condTcondTw

TcondT3lnTcondTwT3Tw

8

where _Qcond is the condensing heat transfer rate, Utp,cond, Ush,cond, andUsc,cond denote the heat transfer coefcient of the two-phase section,superheated section and subcooled section in the condenser,respectively, which are expressed as:_Qeva _mrh1 h4

Utp;eva Ai;eva Cr;sh;eva_mr

Ush;eva ln Ta T1

Ta Teva

Ta Teva

Cr;sh;eva _mrT1 Teva 5

where Utp,eva and Ush,eva denote the heat transfer coefcient of thetwo-phase section and superheated section in the evaporator,respectively, which are expressed as:

Utp;eva 1atp;eva di;eva2ktube

lndo;evadi;eva

1ao;s;eva n gfin

Ai;evaAo;eva

!16

Ush;eva 1ash;eva di;eva2ktube

lndo;evadi;eva

1ao;s;eva n gfin

Ai;evaAo;eva

!17

where ao,s,eva is sensible heat transfer coefcient at the outside ofevaporator, which is developed by Wang and Du [24]. atp,eva and

volume: 16.5 cm3/revl (9.90 0.75 mm), polyurethane insulation (thickness 38 mm)

rea: 0.3445 m2, air-side area: 6.17 m2, d:0.13 mm, plon:25.1 mm, pn:1.34 mm,mm).

88 (2011) 41284138 4131Utp;cond 1atp;cond di;cond2ktube

lndo;conddi;cond

1aw

Ai;condAo;cond

19

Ush;cond 1ash;cond di;cond2ktube

lndo;conddi;cond

1aw

Ai;condAo;cond

110

Usc;cond 1asc;cond di;cond2ktube

lndo;conddi;cond

1aw

Ai;condAo;cond

111

where atp,cond, ash,cond and asc,cond represent the heat transfer coef-cient of the two-phase section, superheated section and subcooledsection at the inside of condenser, and aw is the heat transfercoefcient in the water side. The four heat transfer coefcientsare suggested as [26].

consumption, COP of the system and all the state points data (suchas, temperature, pressure, specic volume, specic enthalpy andthe mass ow rate), were calculated as output data. The propertiesof the chosen refrigerant R22 were from REFPROP (version 6.01)developed by NIST (National Institute of Standard and Technology).

4. Results and discussion

4.1. Experimental results and model verication

In order to learn the performance of the ASHPWH system, a ser-ies of experiments were carried out in the environmental chamberunder different ambient conditions based on the meteorologicaldata of Shanghai, as shown in Figs. 46 and Table 2.

In the experiments, the ambient temperature was set constant.After the ASHPWH system starts, the 150 L water in the tank was

con

d

ergy 88 (2011) 412841383.4. TXV model

Since the expansion of refrigerant is assumed to be isenthalpic,the TXV energy balance equation is:

h3 h4 12The mass of refrigerant ( _mr) through the TXV can be calculated by[27]

_mr _mTXV k1 k2 Teva qlPcond Peva

q13

where k1 and k2 were found to be equal to 3.578 105 m2 and2.442 107 m2/K for the valve considered in present study.

3.5. Water tank model

The water is insulated and the helical coiled tube type con-denser is immersed in the water tank, therefore the heat loss fromthe water tank to ambient is negligible. The variation of heat stor-age in the water tank is the combined result of condensing heat re-lease and hot water load, see Hawlader et al. [23] for a detailedaccount of the water tank model. The energy balance equationcan be obtained by

ddt

Cp;wMwTw _Qcond _Qload 14

Before the hot water is delivered to the end user, the hot water owfrom water tank and cold supply water (tap water) would be mixed.The energy conservation equation is expressed as

Cp;w gt Tw Cp;w f t gt Ttap Cp;w f t Tconsume 15where g(t) is hot water ow rate from water tank, which dependson the domestic hot water demand prole f(t). Where Tconsume isthe temperature of consumed hot water, and Ttap is the tap watertemperature.

3.6. Performance index

The COP of the system performance

COP Pn2

jn1_Qcondj DsPn2

jn1Wj Ds16

where Ds is the duration of one time step in numerical calculation.n1 and n2 represent the n1thn2th time step. _Qcondj and W(j) arecondensing heat transfer rate and the electrical power consumptionof compressor and fan at the jth time step. In order to verify the cal-culated results with the experimental data, 3 C water temperatureincrement is chosen from the n1th to n2th time step. The water tem-perature at the n1th and n2th time step are T

n1w and T

n2w , respectively.

Hence, the calculated COP can be regarded as the COP at water tem-perature T

n1w T

n2w

2 .The average COP during the whole heating process,

COPave Pn

j0_Qcondj DsPn

j0Wj Ds17

where 0 and n are the rst and last time steps of the whole heatingprocess, respectively.

3.7. Computing procedure

Based on the mathematical model developed for this ASHPWHsystem, a simulation program was developed to estimate the ther-

4132 J.J. Guo et al. / Applied Enmal performance of the system. The input data for simulation werethe structural parameters, meteorological data and the initialwater temperature in water tank. The heating capacity, electrical15 20 25 30 35 40 45 50 55

3

6

9P Pevaheated from initial temperature to 55 C, and then the system willbe stopped by the micro-controller.

Fig. 4 shows the variation of condensing and evaporating pres-sure when the water tank was heated from 15 C to 55 C in winter(i.e. Ta = 5 C). The simulated results agree well with the experi-mental results. For the condensing pressure, the largest deviationbetween the experimental and simulated results was less than0.5Bar. For the evaporating pressure, the largest deviation betweenthe experimental and simulated results was less than 0.3Bar.

Fig. 5 shows the comparison of heating capacity variation whenwater tank temperature was from 16.5 C to 53.5 C during opera-tion. For winter case (i.e. Ta= 5 C), the experimental heating capac-ity decreased gradually from 2.46 kW to 1.99 kW. While forsummer case (i.e. Ta = 35 C), it increased from 4.18 kW to4.57 kW rstly, and then decreased to 4.05 kW. For winter case,with the increase of water temperature and condensing tempera-ture, the enthalpy difference between the inlet and outlet of thecondenser was decreasing (Fig. 2), which led to lower heatingcapacity. For summer case, when condensing temperature waslower than the ambient temperature, the pressure difference be-tween the condenser and evaporator was too small, which led tolower heating capacity. With the increase of water temperature,the pressure difference was increasing, and then yielded moreheating capacity. When the condensing temperature increased tosome value, the decrease of the enthalpy difference between theinlet and outlet of the condenser dominated, which caused lowerheating capacity. Also, the gure shows that the simulation resultswere in good agreement with the experimental data with maxi-mum deviation of 9.8%. This attributed to actual pressure dropand heat loss of the system.

12

15

18

21

24 Experimental Simulated

, Pe

va (10

5 Pa)

PcondTw ( )Fig. 4. Pcond, Peva vs. Tw in winter (Ta = 5 C).

average COP in the whole year, respectively. The experimentalaverage COP was lower than the simulated average COP. Thiswas due to the pressure drop and heat loss during actual operation,which in turn caused more compressor work and lower efciency.The maximum deviation between the experimental and the simu-lated results was less than 10%.

The cases study showed good agreement between the experi-4.0

4.5

5.0

5.5 Experimental Simulated

acity

(kW

)

J.J. Guo et al. / Applied Energy 88 (2011) 41284138 413315 20 25 30 35 40 45 50 55

2.0

2.5

3.0

3.5

Tw ( )

Hea

ting

cap

Ta= 5

Ta= 35

Fig. 5. Heating capacity vs. Tw in summer and winter condition.

9

10 Experimental As described above, the heating capacity of the ASHPWH systemin the experiments varied with the water temperature and ambi-ent temperature, which led to the variation of COP. Fig. 6 showsthe experimental and simulated COP variation with the water tem-perature from 16.5 C to 53.5 C during operation. The COP de-creased with increasing water temperature for both summer andwinter cases. The maximum deviation between experimental andsimulated results at the same water temperature was less than10%.

Table 2 shows the comparison of the experimental and simu-lated results of system average COP during the whole heatingprocess, under typical working conditions. The experimentalresults show that the average COP during the whole heatingprocess was 5.51 in summer (Ta = 35 C, Twi = 10 C) and 2.82 inwinter (Ta = 5 C, Twi = 25 C), which was the peak and valley of

15 20 25 30 35 40 45 50 552

3

4

5

6

7

8

Tw ( )

COP

Simulated

Ta= 5

Ta= 35

Fig. 6. COP vs. Tw in summer and winter condition.

Table 2Comparison of experimental and simulated average COP with variable Ta and Twi.

Ta Twi = 10 C Error (%) Twi = 15 C Error (%)

Exp Sim Exp Sim

5 3.32 3.45 3.9 3.21 3.33 3.715 4.02 4.13 2.7 3.72 3.88 4.325 4.80 4.98 3.8 4.60 4.71 2.435 5.51 6.00 8.9 5.42 5.64 4.1mental data and the simulation results. Hence, the model was reli-able for further analysis.

4.2. Optimization analysis based on structural parameters and controlpatterns

In order to have a more understanding on the factors affectingthe performance of the system and to be able to optimize its per-formance, more theoretical work was done. Based on the simula-tion model, the effect of structure parameters and two differentcontrol patterns were analyzed in the following section.

4.2.1. Optimization analysis based on structural parametersAs discussed above, such a system is most suitable for residen-

tial buildings in mild weather and domestic hot water is needed allyear round. To ensure well performance of the system, it is neces-sary to learn the effect of variation of the evaporator and condensercoil area on the system performance. As shown in Fig. 7, the 150 Lwater in the tank was heated from 15 C to 55 C. It could be seenthat, initially, the average COP increased with the increase of theevaporator and condenser coil outside area obviously. However,if the evaporator outside area increases beyond a certain value,the slope of average COP trends to be at. This indicated that theeffect of evaporator and condenser coil outside area on averageCOP became less sensitive, based on this kind of heat exchanger.It also indicated that it was unnecessary to enlarge heat exchangearea excessively when a relatively high average COP has been ob-tained. Considering the actual nite assembling room and materialsaving, satisfactory performance can be obtained when the evapo-rator outside area is between 6.0 and 6.5 m2, and the recom-mended outside area ratio of condenser coil to evaporator isbetween 0.14 and 0.31.

4.2.2. Relationship between thermostatic and timing control patternsIn order to optimize the operation strategy, selecting correct

control pattern is a premise. In practical application, thermostaticand timing control patterns are two common control patterns.For the ASHPWH system, the thermostatic control pattern is verywidely used, in this pattern, water temperature of the ASHPWHsystem is set at Tset (setting water temperature). If the water tem-perature is lower than Tset, the ASHPWH system will turn on auto-matically, and the ASHPWH system turns off until watertemperature reaches to Tset [11]. The advantage of this pattern issimple and convenient. While, the disadvantage of this pattern isthat water tank is almost kept at Tset, as a result, the ASHPWH sys-tem always operates when condensing pressure is high, whichlowers the system average COP. For the timing control pattern,when the ASHPWH system turns on, water is heated from initial

Twi = 20 C Error (%) Twi = 25 C Error (%)

Exp Sim Exp Sim

3.00 3.10 3.3 2.82 2.91 3.23.62 3.75 3.6 3.41 3.51 3.0

4.31 4.48 4.0 4.01 4.22 5.25.01 5.31 6.0 4.80 5.04 5.0

hot water consumption volumes of consumers. Since water in

0 2 4 6 8 10 12 14 16 18 20 22 240369

12151821242730

Winter

Transitional season

Ta (

)

Time of day (hr)

Summer

0.044US$/kWh

Fig. 8. Ta in a typical day of Shanghai (1US$ = 6.853RMB).

0 2 4 6 8 10 12 14 16 18 20 22 24

0.4

0.6

0.8

1.0

1.2

1.4

1.6

1.8

2.0

2.2

2.0

2.5

3.0

3.5

4.0

4.5

5.0

5.5

6.0

COP a

ve

Ener

gy c

on

sum

ptio

n(kW

h), Fe

e(10-1

US$

)

Time of day (hr)

Energy consumption Fee COP

ave

Fig. 9. Energy consumption, fee and average COP in a typical day (summer).

2.0

2.2

5.5

6.0

(10-1 U

S$)

ergythe tank is always heated from initial water temperature to Tset,but not almost kept at Tset like in the thermostatic control pattern,thus, it is more energy-saving because higher system COP can beobtained at relative lower water temperature [28]. The timing con-trol pattern is suitable in some special sites (such as the consumerhas the xed domestic hot water consumption habit, the with-drawal frequency is low and interval between twice withdrawalsis long).

Therefore, the optimization operation strategy should be basedon different control patterns, and a methodology of optimizationanalysis based on these two control patterns will be depicted inthe subsequent contents.

4.2.3. Optimization analysis based on timing control patternIn this pattern, the start-up time determines the thermal and

economic performance of the ASHPWH system. From energy-saving concern, higher COP can be obtained with the higher ambi-water temperature to Tset and then turns off before domestic hotwater withdrawal. After the withdrawal, the water temperaturein the tank is lower than Tset until next start-up of ASHPWH sys-tem. As to the start-up times, it depends on the practical total

5.0 5.5 6.0 6.5 7.0 7.53.0

3.2

3.4

3.6

3.8

4.0

4.2Ta=15 ; Twi=15

Ao, cond= 0.93m

2

Ao, cond= 1.24m

2

Ao, cond= 1.56m

2

Ao, cond= 1.87m

2

Ao,eva

(m2)

COP a

ve

Fig. 7. Effect of evaporator and condenser coil area.

4134 J.J. Guo et al. / Applied Enent temperature during the heating process [15]. So it seems thatthe optimal start-up time should be when the ambient tempera-ture is higher. While, subject to the solar radiation, the diurnalambient temperature is not always kept constant during the heat-ing process, such as in Shanghai, China. Additionally, the policy ofpeak and valley electric power price was carried out in some cities,which impelled the electric consumption at night. However, theCOP of the ASHPWH system drops for the lower ambient temper-ature at night. Energy-saving and money-saving are not alwayscoincident, so that the two factors (uctuation of ambient temper-ature and policy of peak and valley electric power price) should besynthetic considered in timing operation pattern. In order to illus-trate the effect of two factors to the whole ASHPWH system, thecase of Shanghai was discussed as following.

Fig. 8 shows the ambient temperature of a typical day in Shang-hai, the Stages A and B represent peak and valley time of a wholeday with different electric power prices 0.088 US$/kWh and0.044 US$/kWh. Figs. 911 show the energy consumption, electricfees and average COP during the whole heating process when 150 Lwater of the ASHPWH system was heated from 15 C to 55 C fromcertain time in a typical day. Fig. 9 shows the case in summer. Itcould be seen that the valley energy consumption, electric feesand highest average COP during Stage A were 1.3 kWh, 0.113 US$333639

Stage BStageA 0.088US$/kWh

stage B

88 (2011) 41284138and 5.43 respectively when water tank was heated from 14:00for the highest ambient temperature, and the system got the high-est performance, while the valley electric fees of the whole day was0.065 US$ at 22:00 for the valley electric power price. For the samereason, the valley electric fees were 0.079 US$ and 0.094 US$respectively, as shown in Figs. 10 and 11, both at 22:00. It was

0 2 4 6 8 10 12 14 16 18 20 22 24

0.4

0.6

0.8

1.0

1.2

1.4

1.6

1.8

2.0

2.5

3.0

3.5

4.0

4.5

5.0

Energy consumption Fee

Ener

gy c

onsu

mpt

ion(k

Wh),

Fee

COP a

ve

Time of day (hr)

COPave

Fig. 10. Energy consumption, fee and average COP in a typical day (transitionalseason).

00 03 06 09 12 15 18 21 240

2

Time of day (hr)

00 03 06 09 12 15 18 21 240

2

4

6

8

10

12

14

Time of day (hr)

Flow

rate

(L/m

in)

2

4

6

8

10

12

14

Flow

rate

(L/m

in)

(b)

(c)

ergysuggested that the optimal start-up time was near 22:00 for thelowest electric power fees. However, if the electric power priceof the day and night was equal, the optimal start-up time shouldbe chosen between 12:00 and 14:00. It could be derived that theoperation strategy is affected by the price difference of peak andvalley electric power price. The smaller the price difference of peakand valley, the more economic the operation is in the daytime. Thecurves of fee in Figs. 911 also show that the electric power fee wasvery near when the start-up time was from 22:00 to the next 4:00in summer and transitional season, and from 22:00 to the next2:00 in winter. According to the current policy of peak and valleyelectric power price, the start-up time of ASHPWH system is betterto selected as close as possible to these time ranges. As to the start-up times and total energy consumption, it depends on the practicalhot water consumption volume in some special sites.

4.2.4. Optimization analysis based on thermostatic control patternIn the thermostatic control pattern, the water temperature in

the tank is controlled according to Tset (setting water temperature).Usually, Tset is set to be a xed value, such as the ASHPWH systemwith R22, it is always set to 55 C for hot water consumption habitconcern. Whereas, from the viewpoint of energy-saving, decreasingthe Tset will improve the COP of the whole system. Additionally, atoo low Tset will lead to the hot water shortage if the hot water con-sumption volume is large. Therefore, there must be a compromise

0 2 4 6 8 10 12 14 16 18 20 22 24

0.4

0.6

0.8

1.0

1.2

1.4

1.6

1.8

2.0

2.2

2.0

2.5

3.0

3.5

4.0

4.5

5.0

5.5

6.0

Energy consumption Fee

Time of day (hr)

Ener

gy c

onsu

mpt

ion(k

Wh),

Fee(1

0-1U

S$)

COP a

ve

COPave

Fig. 11. Energy consumption, fee and average COP in a typical day (winter).

J.J. Guo et al. / Applied Enbetween Tset and consumed volume of domestic hot water. How-ever, the consumed volume of domestic hot water as well as theconsumption pattern with time varies from household to house-hold [29], it is therefore difcult to dene general domestic hotwater demand proles [3032]. Considering the diurnal and sea-sonal variation of domestic hot water consumption, Roman Spuret al. [33] presented a typical scheme which could contain widescale of domestic hot water usages in household, as shown inFig. 12. The consumption volume of domestic hot water wasnon-homogeneous during the typical day, and the domestic hotwater demand prole was classied into three levels: heavy, med-ium and light [33]. Similarly, domestic hot water demand of a com-mon family is usually the highest in winter, followed bytransitional season and summer, which also indicates three levelsof hot water consumption. Thus, the three levels domestic hotwater demand proles presented by Spur were assumed the dailyhot water consumption of a family in winter, transitional seasonand summer in this study.

In order to learn the effect of Tset on the ASHPWH system perfor-mance, so as to establish suitable operation strategy, a case of afamily in Shanghai was investigated. The ambient temperaturewas based on the description in Fig. 8. The (a), (b) and (c) in4

6

8

10

12

14

Flow

rate

(L/m

in)

(a)

88 (2011) 41284138 4135Fig. 12 were assumed a typical day hot water demand prole f(t)in summer, transitional season and winter, respectively. The heat-ing performance of the ASHPWH system was shown in Figs. 1315,which represent the typical seasonal performance.

An evaluating indicator called delivery coefcient (i.e. f) wasintroduced, and it was dened as the ratio of actual heat the userconsumed hot water to the domestic hot water demand during aday time, as below

f R 864000

_QconddtCp;wVconsumeqwTconsume Ttap

18

where the Vconsume represents 100 L, 180 L and 320 L when it is insummer, transitional seasons and winter, respectively. The

00 03 06 09 12 15 18 21 240

Time of day (hr)Fig. 12. Hot water demand prole of a typical day: (a) light; (b) medium; (c) heavy[33]. (The width of each column in this gure indicates the duration of eachwithdrawal. The duration of each withdrawal corresponds to the ow rate 1, 6, 8and 14 L/min is 1, 1, 5 and 10 min, respectively.)

00 03 06 09 12 15 18 21 24

00 03 06 09 12 15 18 21 24

30

32

34

36

06 07 08

40

42

Time of day (hr)

Tw,c

on

sum

e (

T w,c

on

Time of day (hr)

32

34

36

38

40

42

44

46

06 07 0838

40

42

44

46

Tw,c

on

sum

e (

)

Time of day (hr) Time of day (hr) 201938

40

42

44

46

Tw,c

on

sum

e (

)

Time of day (hr)

Tw,c

onsu

me

()

Tset=46

Tset=48

Tset=50

36

38

40

42

44

46

4446

(

)

w,c

onsu

me

(

Tset=46

Tset=48

Tset=50

Tset=52

(b)

(c)

ergytemperature of tap water and consumed domestic hot water are15 C and 45 C, respectively, in the simulation.

Fig. 13 shows the inuence of Tset on the electric consumption ofthe system. With the decrease of Tset from 55 C to 46 C, the elec-tric consumption of the system during a typical day in summer,transitional season and winter decreased by 20.1%, 20.6% and22.8%, respectively. It seems that a lower Tset should be a betterchoice for improving the system performance. Nevertheless, sucha strategy may lead to the hot water supply trouble, as discussedbefore, the hot water shortage. Such an effect was depicted inFig. 14 detailedly.

Fig. 14 shows the user consumed water temperature under dif-ferent Tset during a typical day in different seasons. The region sizewith temperature below 45 C (the required temperature ofdomestic hot water) indicated the shortage extent of hot water.It can be seen that hot water shortage extent was quite differentwith different Tset and different seasons. The hot water shortage ex-tent can be characterized by two parameters: the shortage timeand the temperature valley value of user consumed water, whichsomewhat like the bottom and height of a triangle. With a sameTset, the shortage time was shortest in summer and longest in win-ter, and the temperature valley value was highest in summer andlowest in winter. The case in transitional season was in the middleof summer and winter. With the increase of Tset, the hot watershortage will be relieved.

46 48 50 52 54 56

0.8

1.2

1.6

2.0

2.4

2.8

3.2

3.6

4.0El

ectri

c co

nsum

ptio

n (kW

h)

Tset ( )

Electric consumption (Winter)Electric consumption (Transitional season)Electric consumption (Summer)

Fig. 13. Daily electric consumption vs. Tset.

4136 J.J. Guo et al. / Applied EnIn summer, the hot water load is relatively light, thus even avery low Tset can achieve the goal of hot water supply. As shownin Fig. 14a, even when the Tset was set as low as 46 C, whichwas very close to the required temperature 45 C, the shortagetime in the whole day was less than fteen minutes and the tem-perature valley value was higher than 40 C. In this case, the hotwater shortage extent was less than 1%, which can be seen inFig. 15 (delivery coefcient f > 99%). Such a light shortage can beneglected and is acceptable for the user. Therefore, the Tset couldbe set as low as 46 C in summer. In winter, the hot water load isrelatively high, thus a low Tset might be unable to afford the hotwater requirement. As shown in Fig. 14c, if the Tset was set as46 C (same as the case in summer), the hot water shortage willbe very large. In this case, the shortage time lasted about 1.5 hand the temperature valley value was as low as 33 C. Such a hotwater shortage was too heavy and unacceptable. Therefore, a high-er Tsetwas required. As shown in the gure, when the Tsetwas set as52 C, the hot water shortage will be eliminated.

According to above discussion, the Tset inuenced both the en-ergy performance and the hot water supplement performance ofthe system, and these two effects contradicted each other. The Tset38

40

42

44

46

44

46

)

Tset=46

Tset=48

Tset=50

sum

e (

)

(a)

88 (2011) 41284138can not be too high and too low, and there will be an optimum va-lue of Tset in every season for the ASHPWH system. Fig. 15 showsthe inuence of Tset on the daily average COP and delivery coef-cient during a whole day of the system. According to the denition,the delivery coefcient reects the hot water load satisfaction de-gree. If there is no hot water shortage during the whole day, thedelivery coefcient will be 100%, otherwise, the delivery coefcientwill be smaller than 100%. As shown in Fig. 15, the daily averageCOP increased with the decrease of Tset for the lower water temper-ature and condensing temperature. While the delivery coefcient

00 03 06 09 12 15 18 21 2430

32

34

06 07 08 0936384042

Tw,co

nsu

me

Time of day (hr)

Time of day (hr)

T

Fig. 14. User consumed water temperature during a typical day: (a) summer; (b)transitional season; (c) winter.

ergydeclined with the decrease of Tset. Considering that the deliverycoefcient should be over 99%, which ensures the shortage timefor the user to wait less than 20 min, the Tset should be set above46 C, 49 C and 51 C, in summer, transitional season and winter,respectively. These values were the optimum Tset of the system fordifferent seasons, and it was obvious that the optimum settingwater temperature should be adjusted according to the variationof seasonal ambient temperature. If the Tset was lower than thesevalues, the hot water supplement can not meet the hot waterrequirement. In these cases, it was not suitable to evaluate the sys-tem performance by the daily average COP. Under the optimumconditions, the daily average COP of the system were 4.3, 3.2 and2.6, for the summer, transitional season and winter, respectively.

5. Conclusions

An experimental set-up of air-source heat pump water heater(ASHPWH) was constructed and tested. Based on a simulationmodel, the optimization analysis of structural parameters was em-ployed. Also, a methodology of optimizing operation, which takesinto account thermostatic and timing control patterns, was pre-

46 48 50 52 54 562.4

2.7

3.0

3.3

3.6

3.9

4.2

4.5

4.8

5.1

84

88

92

96

100

104

108 COPave COP

ave

COPave

Tset ( )

COP a

ve

(%

)

(Winter) (Transitional season) (Summer)

Fig. 15. Daily average COP, delivery coefcient vs. Tset.

J.J. Guo et al. / Applied Ensented and applied. The detailed conclusions are as follow:

(1) The experimental results of this experimental set-up indi-cated that the average COP of the system under typicalmeteorological conditions of Shanghai ranged from 2.82(Ta = 5 C, Twi = 25 C) to 5.51 (Ta = 35 C, Twi = 10 C) whenwater was heated to 55 C.

(2) The effect of evaporator and condenser coil outside area onaverage COP becomes less sensitive if outside area of con-denser coil and evaporator increases beyond certain value.The recommended outside area ratio of condenser coil toevaporator is between 0.14 and 0.31 when the evaporatoroutside area is between 6.0 and 6.5 m2 for this set-up.

(3) As for the timing control pattern, the optimal start-up timewas affected by uctuation of ambient temperature andthe price difference of peak and valley electric power. Thesmaller the price difference of peak and valley, the more eco-nomic the operation is in the daytime. The analysis aboutthe typical day in Shanghai suggested that the optimalstart-up time was between 12:00 and 14:00 if there wasno electricity price difference between day and night, or itwas near 22:00 for the lowest electric power fees. As tostart-up times, it depends on the practical hot water con-sumption volume of consumers in some special sites.[21] Huang BJ, Wang JH, Wu JH, Yang PE. A fast response heat pump water heaterusing thermostat made from shape memory alloy. Appl Therm Eng2009;29:5663.

[22] Kline J, MoClintock FA. Describing uncertainties in single sample experiments.Mech Eng 1953;34:38.

[23] Hawlader M, Chow SK, Ullah MZ. The performance of a solar assisted heatpump water heating system. Appl Therm Eng 2001;21(10):104965.

[24] Wang CC, Du YJ. Airside performance of herringbone n-and-tube heatexchangers in wet conditions. Can J Chem Eng 1999;77(6):122530.

[25] Klimenko VV. A generalized correlation for two-phase forced ow heat(4) Based on the thermostatic control pattern, an evaluatingindicator delivery coefcient and the optimum setting watertemperature were introduced. The optimum setting watertemperature should be adjusted according to the variationof seasonal ambient temperature. It was suggested that,based on this system, setting water temperature should beset higher than 46 C in summer and 50 C in other seasons.

Acknowledgments

This research is supported by Hi-Tech Research and Develop-ment Program of China (Grant No. 2007AA05Z220). Authors areparticularly grateful to Dr. Qiang Ma for his technical supportand to Mr. Yuxiong Xu and Yunkang Sun for their help in theexperimental work.

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4138 J.J. Guo et al. / Applied Energy 88 (2011) 41284138

Experimental research and operation optimization of an air-source heat pump water heaterIntroductionDescription of system and experiment set-upWorking process of the ASHPWH systemExperimental facility and proceduresError analysis

Mathematical model and simulation processCompressor modelEvaporator modelCondenser modelTXV modelWater tank modelPerformance indexComputing procedure

Results and discussionExperimental results and model verificationOptimization analysis based on structural parameters and control patternsOptimization analysis based on structural parametersRelationship between thermostatic and timing control patternsOptimization analysis based on timing control patternOptimization analysis based on thermostatic control pattern

ConclusionsAcknowledgmentsReferences