Applied Energy 102 (2013) 1229–1245

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Applied Energy

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Characterization of a solar photovoltaic/loop-heat-pipe heat pump waterheating system

Xingxing Zhang a, Xudong Zhao a,⇑, Jihuan Xu b, Xiaotong Yu c

a Institute of Energy and Sustainable Development, De Montfort University, UKb Shanghai Pacific Energy Centre, Shanghai, Chinac Shanghai Solar Energy Research Centre, Shanghai, China

h i g h l i g h t s

" Describing concept and operating principle of the PV/LHP heat pump water heating system." Developing a numerical model to evaluate the performance of the system." Experimental testing of the prototype system." Characterizing the system performance using parallel comparison between the modelling and experimental results." Investigating the impact of the operating conditions to the system’s performance.

a r t i c l e i n f o

Article history:Received 28 March 2012Received in revised form 11 June 2012Accepted 17 June 2012Available online 24 July 2012

Keywords:PVLoop heat pipeHeat pumpEfficiencySolarEnergy Performance

0306-2619/$ - see front matter � 2012 Elsevier Ltd. Ahttp://dx.doi.org/10.1016/j.apenergy.2012.06.039

⇑ Corresponding author. Tel.: +44 116 257 7971; faE-mail address: xzhao@dmu.ac.uk (X. Zhao).

a b s t r a c t

This paper introduced the concept, potential application and benefits relating to a novel solar photovol-taic/loop-heat-pipe (PV/LHP) heat pump system for hot water generation. On this basis, the paperreported the process and results of characterizing the performance of such a system, which was under-taken through dedicated thermo-fluid and energy balance analyses, computer model development andoperation, and experimental verification and modification. The fundamental heat transfer, fluid flowand photovoltaic governing equations were applied to characterize the energy conversion and transferprocesses occurring in each part and whole system layout; while the energy balance approach was uti-lized to enable inter-connection and resolution of the grouped equations. As a result, a dedicated com-puter model was developed and used to calculate the operational parameters, optimise thegeometrical configurations and sizes, and recommend the appropriate operational condition relating tothe system. Further, an experimental rig was constructed and utilized to acquire the relevant measure-ment data that thus enabled the parallel comparison between the simulation and experiment. It is con-cluded that the testing and modelling results are in good agreement, indicating that the model has thereasonable accuracy in predicting the system’s performance. Under the given experimental conditions,the electrical, thermal and overall efficiency of the PV/LHP module were around 10%, 40% and 50% respec-tively; whilst the system’s overall performance coefficient (COPPV/T) was 8.7. Impact of the operationalparameters (i.e. solar radiation, air temperature, air velocity, heat-pump’s evaporation temperature, glaz-ing covers, and number of the absorbing heat pipes) to the performance of the system (in terms of effi-ciencies of the PV/LHP module and the system’s overall performance coefficient COPPV/T) was investigatedindividually. The results indicated that lower solar radiation, lower air temperature, higher air velocityand smaller cover number led to enhanced electrical efficiency but reduced thermal efficiency of themodule; whereas lower heat-pump’s evaporation temperature and larger number of heat absorbing pipesgave rise to both thermal and electrical efficiencies of the module. The research results would assist indeveloping a high efficient solar (space or hot water) heating system and thus contribute to realisationof the energy saving and associated carbon emission targets set for buildings globally.

� 2012 Elsevier Ltd. All rights reserved.

ll rights reserved.

x: +44 116 257 7981.

1. Introduction

The PV technologies are currently positioned at the forefront ofthe global measures leading to sustainability and low carbon

Nomenclature

Am effective module area (m2)Ac,r cross area of refrigerant flow (m2)Ahx,r cross area of refrigerant in heat exchanger (m2)Cp,lr specific heat capacity of liquid refrigerant (J/kg K)D diameter (m)F standard fin efficiencyFth thermal efficiency factorg gravity acceleration (m/s2)hc convective heat transfer coefficient (W/m K)hfg latent heat of vaporization (J/kg)hr refrigerant heat transfer coefficient (W/m K)hr,l heat transfer coefficient of liquid refrigerant (W/m K)hR radiative heat transfer coefficient (W/m K)H thermal enthalpies (J/kg)I solar radiation intensity (W/m2)K thermal conductivity (W/m2 K)L length (m)m fin variablemr mass flow rate of refrigerant (kg/s m2)mw mass flow rate of water (kg/s)n mesh numberN numberNu Nusselt numberp pressure (pa)q energy rate per unit area (W/m2)Q energy rate (W)Pr Prandtl numberr thermal resistance per square meter (m2 K/W)R thermal resistance (K/W)R0 universal gas constant (kJ/kmol K)Ra Rayleigh numbert temperature (�C)U overall heat coefficient (W/m K)V velocity (m/s)W width (m)

Greeka absorption ratioar,l refrigerant thermal diffusivitybp PV packing factorbpv cell efficiency temperature coefficientd thickness (m)e emissivityewi porosity of the wickg efficiency

h collector slop (deg)l dynamic viscosity (kg/m s)q density (kg/m3)s visual transmittancer Stefan–Boltzman constant

Subscriptsa airabs absorptionb backplanec coverc,e compressor electrical energyc,t condensation thermal energyc1 internal cover sheetc2 external cover sheete electricityei electrical insulatione,n net electricitye,t evaporation thermal energyEVA ethylene–vinyl-acetatef three-way fittingg,pv glass layer of PV limitationhp heat pipehp,e heat pipe evaporatorhp,in inner heat pipehp,o outer heat pipehp,w heat pipe wallhp,w-r heat pipe wall to refrigeranthx heat exchangerm meanl liquidlf liquid filmL lossp PVp-fin PV to fin sheetr refrigerantrc reference temperaturerm mean refrigerants solid; isentropicth thermaltl transporting lineo overallu usefulv vapourwi wick

1230 X. Zhang et al. / Applied Energy 102 (2013) 1229–1245

economy, and present great potential of delivering around 5% ofglobal electricity demand by 2030 and 11% by 2050 [1]. To acceler-ate the process, increasing the PV’s solar conversion efficiency andreducing its capital cost are regarded the top priority in PV technol-ogy innovation. It is recognised that the PV’s efficiency falls whenits cells temperature rises [2–4]. To control the temperature ofthe cells, several PV/Thermal (PV/T) measures were given consider-able trials; some of those were found to be able to effectivelyremove the accumulated heat from the back surface of the PVsand deliver the collected heat to buildings for energy supplypurpose, thus developing the effective methods in enhancing PV’ssolar conversion ratio and making economic use of the PV systems.

Concept of the PV/T was initiated by Kern and Russell [5] in1978. Sooner after this, Florschuetz [6] developed a mathematicaltheory dedicated for the flat-plate PV/T collector, which is actuallythe extended Hottel–Whillier [7] model. In recent years, numerous

researchers made efforts to develop PV/T technologies includingthose by air [8–13], water [14–18], refrigerant [19–23], and heatpipe [24–26]. To give a brief, the electrical and thermal efficienciesof the PV/T modules at the peak loading condition are (1) 8% and42% for air-based type [13]; (2) 9.1% and 41% for water-based type[18]; (3) 12% and 50% for refrigerant-based type [22]; and (4) 9.4%and 41.9% for heat-pipe-based type [26]. These figures have indi-cated the excellent effectiveness of the PV/T devices in increasingsolar energy yield. Meanwhile, use of these PV/T devices has alsodiscovered several inherent technical shortfalls. In terms of theair based PV/T type, the major problem lies in its relatively poorheat removal effectiveness owing to the relatively lower thermalmass of the air. The water-based PV/T type has certain improve-ment in removing the PV heat but the degree of improvement isvery limited due to the gradually increasing water temperaturewithin the loop; it has also presented the problem of potential

Fig. 1. Schematics of (a) experiment rig of the heat pump assisted PV/LHP solar water heating system, and (b) heat pump thermodynamic cycle in T–S chart.

X. Zhang et al. / Applied Energy 102 (2013) 1229–1245 1231

freezing during winter operation. Although the refrigerant-basedPV/T system could achieve significant increase in solar energy con-version, its practical feasibility faced several challenges e.g., diffi-culty in remaining the required pressures (positive and negative),high risk of the refrigerant leakage and the uneven refrigerant dis-tribution across multiple coils installed in a large photovoltaic area[27]. The heat-pipe-based PV/T type, under an adequate operatingtemperature/pressure, could obtain a similar thermal efficiency asto the refrigerant-based system. This type of system may thereforehave potential of overcoming the difficulties existing in above sys-tems and become the next generation PV/T device. The advantagesof the heat pipe technology lie in: (1) efficient thermal transfercapacity for a distance travel; (2) hermetically sealed vessel with-out risk of fluid leakage; (3) homogeneous built-in capillary forceleading to an even heat distribution; (4) availability for use ofthe anti-freezing liquid [28].

However, the heat-pipe-based PV/T system is still at the re-search stage and its practical use remains certain degree of uncer-tainty and challenge. For distant heat energy transportation underbuilding application, loop heat pipe is a feasible solution which

could transport solar heat from the outer façade/roof of the build-ing to its inside. The LHP is a two-phase heat-transfer measurewith the working fluid circulating in a loop, thus enabling remote,passive heat transfer at enhanced capacity. The LHP has beenwidely used in thermal control of satellites, spacecrafts, electronicsand cooling/heating systems [29–34], while the use for solar en-ergy collection and transportation is only the recent development.The biggest problem facing the loop heat pipe is the ‘dry-out’ po-tential of the water film on the upper side wall of the heat pipe ab-sorber, due to limited water uplift height caused by the insufficientwick capillary force [28]. To overcome this difficulty, a unique loopheat pipe structure with the top positioned three-ways tube wasinitiated. This structure, in combination with the PV layer, couldform a modular PV/thermal solar collector. Combined operationof the modular PV/thermal and the heat pump will expect to over-come the ‘dry-out’ problem remaining in conventional heat-pipe-based PV/T configurations, thus leading to highly efficient andlow cost heat and electricity generation using solar energy. In linewith this initiative, a characteristic study of such a PV/LHP heatpump system will be conducted theoretically and experimentally

Fig. 2. Schematic of loop heat pipe operation.

Fig. 3. Schematic of (a) the PV/LHP collector and (b) configuration of PV lamination.

1232 X. Zhang et al. / Applied Energy 102 (2013) 1229–1245

in this research; which will expect to identify the real behavior ofthe system, suggest the favorite operational conditions and opti-mize the geometrical sizes of the system configurations. The re-search results will help wide deployment of the PV/thermaltechnologies, and thus contribute to significant fossil fuel energysaving and cut of the associated carbon emission in building sector.

2. System descriptions

The proposed PV/LHP heat pump water heating system andassociated T–S chart based thermodynamic cycle are shown sche-

matically in Fig. 1. This system, as shown in Fig. 1a, comprises amodular PV/LHP solar collector, an electricity control/storage unit,the vapour/liquid transportation lines, a flat-plate heat exchangeracting as the condenser of the heat pipe loop and the evaporatorof the heat pump cycle, a hot water tank, a compressor, a coil-typecondenser embedded into the water tank and an expansion valve.The LHP absorber (Fig. 2), as the most important heat collecting de-vice, is an externally finned and internally wicked heat pipe withthe three-way tube on the top. This pipe, under the condition thatthe water streams are continuously evaporated from the heat pipewall due to absorption of solar radiation, could deliver evenly dis-tributed water film across the heat pipe wall from the upper side,

Fig. 4. Schematic of the solar energy conversion and transfer processes.

Fig. 5. Thermal network of heat loss for a typical double cover module.

X. Zhang et al. / Applied Energy 102 (2013) 1229–1245 1233

and keep constant wall wetting condition throughout its full sur-face, thus preventing the ‘dry-out’ potential of the water acrossthe wall. The three-way tube, meanwhile, could also deliver the va-pour upward to the flat-plate exchanger (LHP condenser) throughthe vapour transportation line. This will create a clear separationbetween the liquid and vapour flows in the heat pipe.

In the PV/LHP module, a unique LHP absorber is fitted under-neath the PV layer, as shown in Fig. 3, in order to extract heat fromthe PV’s back and thus reduce the PV cells temperature and in-crease PVs’ electrical efficiency. During operation, this part of heatwill be delivered to the flat-plate heat exchanger through vapourtransportation line, within which heat transfer between the heatpump refrigerant and heat pipe working fluid will occur. This inter-action between the heat pipe fluid and heat pump refrigerant willlead to condensation of the heat pipe working fluid. The condensedliquid will return to the LHP absorber via the liquid transportationline, thus completing the heat pipe fluid circulation.

In the heat pump cycle (compressor–condenser–expansion-valve–evaporator–compressor), the liquid refrigerant will bevaporized in the heat exchanger, which, under the pressurizationby the compressor, will be subsequently converted into higherpressure, supersaturated vapour, and thus transfer heat energyinto the tank water via the coil exchanger (condenser of the heatpump cycle), thus leading to temperature rise of the tank water.It should be addressed that the heat transfer process within the coilexchanger will also lead to condensation of the high pressure,supersaturated vapour, which, when passing through the expan-sion valve, will be downgraded to the low pressure liquid refriger-ant. This refrigerant will undergo the evaporation process withinthe flat-plate heat exchanger (evaporator of the heat pump cycle),thus completing the whole heat pump cycling. The thermodynamicprocess of the refrigerant within the heat pump cycle is shownschematically in Fig. 1b.

The distinct features of the PV/LHP heat pump system lie in (1)temperature of the LHP working fluid could be controlled to a low-er level through adjustment of the evaporation pressure of therefrigerant in the heat pump cycle; this will lead to the reducedPV cells temperature, increased PV electrical and thermal efficien-cies, and increased solar output per unit of absorbing surface; (2)refrigerant temperature/pressure will be increased to a required le-vel by using a compressor to enable heat to be transferred from therefrigerant to the hot water; (3) power needed for compressoroperation can be provided by the PV generated electricity if thesystem is appropriately designed, thus creating a near-to-zero-car-

bon heating operation. It can be predicted that more or less elec-tricity surplus or deficiency may occur, which could be matchedthrough a battery storage or grid.

This system may be installed on a building where the PV/LHPmodules could be mounted onto its façade or roof. For this applica-tion, the heat exchanger could be positioned at the upper side ofthe modules, while the heat pump is installed at inside of thebuilding. Alternatively, the system can be separately installed asan independent heat and power generation unit.

3. Mathematical model of the solar energy conversion andtransfer

Zondag et al. [35] indicated that the simple 1D steady-statemodel is effective in simulating the performance of a combi-paneland uses much less computing time than the 2D and 3D models.For this regard, the 1D matrix was considered in this study.

In a PV/LHP module based heat pump system, the solar energyconversion and transfer involves four processes (Fig. 4), namely: (i)

Fig. 6. Temperature drops and equivalent thermal resistances along the heat transfer path.

1234 X. Zhang et al. / Applied Energy 102 (2013) 1229–1245

absorbing the certain percentage of striking solar radiation whileremaining being dissipated into the surrounding air, (ii) convertingpart of the absorbed energy into electricity via the PV cells, (iii)transporting the other part of the absorbed energy into the passingrefrigerant via the loop heat pipe, and (iv) upgrading the refriger-ant-received heat into higher grade energy using a heat pump.These processes are inter-linked and finally can achieve a balanceunder the steady state operation. To simplify the energy model,the following hypotheses are made:

a. The system operates under the quasi-steady state condition.b. The ohm-electrical losses within the solar cells and PV mod-

ule are ignored.c. The transmittance of EVA (ethylene–vinyl-acetate) layers is

considered to be 100%.d. Heat losses across the module insulation layers (back and

edge) are negligible.e. Heat losses through the heat pipe transportation lines are

ignored.

3.1. Absorbing certain percentage of striking solar radiation whileremaining being dissipated into the surroundings

When the solar radiation passes across the top covers andstrikes on the surface of the PV/LHP module, small percentage ofthe radiation energy will be dissipated into the surroundings ow-ing to the directional/diffusive reflection and conductive/convec-tive heat transfer occurring; while the remaining radiationenergy will be converted into electricity and heat using the PV/LHP settings. Under a steady-state operating condition, the usefulenergy obtained by the module is equal to the energy reachingthe PV surface minus the associated direct or indirect heat losses,which refers to the heat dissipation from the module surface tothe surrounding air through conduction, convection and infraredradiation.

The solar energy received by the PV absorber is a function of thesolar radiation striking on the panel, the transmittance of glazingcover and the absorptance of the PV surfaces, and can be expressedas [36,37]

Qabs ¼ sNcc sg;pv ½apbp þ abð1� bpÞ�AmI ð1Þ

Owing to temperature difference between the PV surface andsurrounding air, certain amount of the PV receiving energy willbe transferred into the surrounding air through the top cover;while the back and edge heat transfers could be ignored as superinsulation at those directions is provided. Under the steady-statecondition, the heat loss from a double-glazed module will experi-ence (1) heat transfer from the PV absorber surface to the innerglazing cover; (2) heat transfer from the inner cover to outer cover;and (3) heat transfer from the outer glazing cover to the surround-ing air [36,37]. As shown in Fig. 5, the three heat transfers will belaid in series and achieve the balance. Therefore, the total heat losscould be expressed as [37]

QL ¼ ULAmðtp � taÞ ð2Þ

where the UL is the overall heat transfer coefficient and could bewritten [37]

UL ¼1

hc;p�c2 þ hR;p�c2þ 1

hc;c2�c1 þ hR;c2�c1þ 1

hc;c1�a þ hR;c1�a

� ��1

ð3Þ

3.1.1. Heat transfer from the PV absorber surface to the inner glazingcover

In this case, a relatively steady convective air layer is in exis-tence between the PV absorber surface and the inner glazing cover,its associated convective heat transfer coefficient, hc,p�c2, can be ex-pressed as [37]

hc;p�c2 ¼Ka;p

da;p1þ 1:446 1� 1708

Raa;p cos h

� �þ1� 1708 sinð1:8hÞ1:6

Raa;p cos h

" #(

þ Raa;p cos h5830

� �0:333

� 1

" #þ)ð4Þ

Raa;p ¼gðtp � tc2Þd3

a;p

m2a;pta;m

Pra;p ð5Þ

ta;m ¼ ðtp þ tc2Þ=2 ð6Þ

Fig. 7. An elemental length ‘dx’ showing heat flow pattern at the fin sheet.

X. Zhang et al. / Applied Energy 102 (2013) 1229–1245 1235

Converting the radiation transfer into the equivalent convectiveone, a radiation-relevant factor, hr,p�c2, is termed and expressed as [37]

hr;p�c2 ¼rðtp þ tc2Þðt2

p þ t2c2Þ

ð1=epÞ þ ð1=ec2Þ � 1ð7Þ

3.1.2. Heat transfer from the inner glazing cover to the outer coverBy carrying out a similar analysis, the heat transfer from the in-

ner glass at tc2 to the outer glass at tc1 could be expressed as:

hc;c2�c1 ¼Ka;c

da;c1þ 1:446 1� 1708

Raa;c cos h

� �þ�

� 1� 1708 sinð1:8hÞ1:6

Raa;c cos h

" #þ Raa;c cos h

5830

� �0:333

� 1

" #þ)ð8Þ

Raa;c ¼gðtc2 � tc1Þd3

a;c

m2a;ctc;m

Pra;c ð9Þ

tc;m ¼ ðtc1 þ tc2Þ=2 ð10Þ

hr;c2�c1 ¼rðtc1 þ tc2Þðt2

c1 þ t2c2Þ

ð1=ec1Þ þ ð1=ec2Þ � 1ð11Þ

3.1.3. Heat transfer from the outer surface of the cover to thesurrounding air

For a surface exposed to the outside wind, its convective coeffi-cient could be calculated using the Klein equation [38], as below

hc;c1�a ¼8:6V0:6

L0:4 ð12Þ

It should be addressed that the minimum convective coefficientfor wind-exposed surface is considered to be 5 W/m2 K [38]; if theabove calculation gives a lower value, this should be replaced bythe minimum value of 5.

Since the sky temperature has little impact to the calculation re-sult, it is usually represented by the air temperature, thus [37]hr;c1�a ¼ ec1rðtc1 þ taÞðt2

c1 þ t2aÞ ð13Þ

For the modules with single glazing or with no glazing top cov-er, the items addressed in 3.1.2 or 3.1.1 and 3.12 should be re-moved, while the heat transfer from inner glass to outer glasswill not be counted. In this case, only heat loss between the mod-ule surface and ambient air is considered.

3.2. Converting part of the absorbed energy into electricity

The PV cells’ electrical efficiency is adversely proportional to itssurface temperature and this dependency can be written as

gc ¼ grc½1� bPV ðtp � trcÞ� ð14Þ

The overall electricity output is therefore given as

Qe ¼ gcbpapsNcc sg;pv IAm ð15Þ

The module’s solar electrical efficiency could be calculated byusing the following equation

ge ¼Q e

IAmð16Þ

3.3. Transporting the remaining absorbed energy into evaporationheat of the refrigerant

Under the steady-state condition, the rate of useful heat deliv-ered by module equals to the rate of the absorbed energy minusthe overall heat loss and converted electricity, which could be ex-pressed as

Qth ¼ Q abs � Q L � Q e ð17Þ

This part of heat will eventually be converted into the heat re-ceived by the refrigerant, which is denoted by Qu. In that case, themodule’s thermal efficiency can be defined as

gth ¼Qth

AmI¼ Qu

AmIð18Þ

The module’s overall solar efficiency, go, would be the sum ofboth electrical and thermal efficiencies, defined as

go ¼ ge þ gth ð19Þ

At the heat pipe’s evaporation section (heat absorbing pipes),part of the solar energy is converted into the heat pipe receivedheat (Qth), which leads to evaporation of the heat pipe workingfluid. This vapour fluid, via the vapour transportation line, movesforward to the condensing heat exchanger where the evaporatedfluid is condensed and transfers the condensation heat into theadjacent refrigerant flow, thus leading to evaporation of the refrig-erant at the next channel of the heat exchanger. The condensedfluid in the heat pipe, via the liquid transportation line, then flowsback the heat pipe evaporation section to regain heat, thus forming

1236 X. Zhang et al. / Applied Energy 102 (2013) 1229–1245

the complete heat transportation cycle, as depicted in Fig. 6. Thisprocess involves a number of thermal resistances that result inchanging in the working fluid temperature, which is detailed as fol-low. It should be addressed that the resistances of the silicon seal-ant and liquid–vapour interface at wicked surface are ignoredowing to their significantly smaller values compared to others [29].

3.3.1. Thermal resistances between the PV cells and heat pipe finsThe heat transfer between the PV absorber and heat pipe fins is

the one-dimensional multi-layer heat conduction process and itsassociated thermal resistance could be written as [39]

rp-fin ¼ rp þ rEVA þ rei ¼dp

Kpþ dEVA

KEVAþ dei

Keið20Þ

3.3.2. Thermal resistance across the fin lengthThe heat received by the fin is conducted to the heat pipe wall

along its width direction, by leading the heat flow to travel acrossthe cross sectional area of the fin. This is considered a one-dimen-sional heat transfer process starting from the fin end (x = 0) and

Fig. 8. Flow chart of co

finishing at fin base [x = (W/Nhp � Dhp,o)/2]. Fig. 7 indicates the heatflow simulation process using the finite element approach whichtakes dx as the step length of the numerical calculation.

For a controlled finite element per unit width, the following en-ergy conversation equation can be applied [37]

qabs � ULðt � taÞ � qe

1þ rp-finUL

� �dxþ �Kede

dtdx

� �x

� �Kededtdx

� �xþdx

¼ 0 ð21Þ

Along the heat pipe evaporation section, no temperature gradi-ent would be in existence along its length direction, owing to theeven heat input. By applying the Hottel–Whillier model [7,36],the overall useful solar heat conducted from the fin to heat pipecould be expressed asQu ¼ LWFth½qabs � ULðtr;m � taÞ � qe� ð22Þ

Fth ¼1=UL

LWNhp

1

LUL ðW=Nhp�Dhp;oÞFþDhp;o

1þrp-finUL

h iþX6

i¼1

Ri

8<:

9=;

ð23Þ

mputation process.

Table 1Photovoltaic characteristics of the PV module under standard testing conditions.

At short-circuit current ISC = 5.54 A, VSC = 0 VAt open-circuit voltage IOC = 0 A, VOC = 22.32 VAt the maximum power point Imp = 4.89 A, Vmp = 18.23 V (Pmp = 89.1 W, go = 16.8%)

Table 2Design parameters of the LHP operation and heat exchanger.

Parameters Nomenclature Value Unit

External diameter of evaporator Dhp,o 0.022 mInternal diameter of evaporator Dhp,in 0.0196 mInternal diameter of vapour column (three-way fitting) Dvt 0.014 mThermal conductivity of evaporator wall Khp 394 W/m KOperating pressure in heat pipe Php 1.3 � 10�4 PaEvaporator length Lhp,e 1.2 mEvaporator-to-condenser height difference Hhx-hp 0.3 mLiquid filling level mf 75 mlTransportation line outer diameter Dltl,o/Dvtl,o 0.022 mTransportation line inner diameter Dltl,in/Dvtl,in 0.0196 mTransportation line length Lltl/Lvtl 1.0/0.9 mWire diameter (wick layer I) Dowi,ms 7.175 � 10�5 mLayer thickness (wick layer I) dowi,ms 3.75 � 10�4 mMesh number (wick layer I) nowi,ms 6299 m�1

Wire diameter (wick layer II) Diwi,ms 12.23 � 10�5 mLayer thickness (wick layer II) diwi,ms 3.75 � 10�4 mMesh number (wick layer II) niwi,ms 2362 m�1

Wick conductivity Ks,ms 394 W/m KHeat exchanger plate thickness dhx 0.00235 mHeat exchanger plate height Hhx 0.206 mHeat exchanger plate cluster width Whx 0.076 mHeat exchanger plate cluster length Lhx 0.055 mHeat exchanger plate conductivity Khx 16.28 W/m KHeat exchanger number of plate nhx 20 –

X. Zhang et al. / Applied Energy 102 (2013) 1229–1245 1237

The standard fin efficiency, F, is defined as

F ¼ tanh½mðW=Nhp � Dhp;oÞ=2�mðW=Nhp � Dhp;oÞ=2

ð24Þ

And, the variable, m, is expressed as [36]

m ¼ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi

UL

Kf df ð1þ rp-finULÞ

sð25Þ

In terms of the physical implication, Fth represents the ratio ofthe system’s actual useful heat gain to the overall converted solarheat at a certain working fluid temperature. The system’s thermalefficiency factor is a constant figure under the fixed physical andoperating condition. However, this factor’s value varies in the fol-low ways: it decreases with increasing the fin width (less absorb-ing pipe), increases with increasing material thicknesses andthermal conductivities, decreases with increasing the overall heatloss coefficient, and increases with decreasing the overall systemheat transfer resistance.

3.3.3. Thermal resistances from absorbing heat pipe wall to heat pipeworking fluid3.3.3.1. Heat pipe wall, R1. Heat transfer through the heat pipe wallis a typical steady-state conduction process, and its thermal resis-tance can be written as [39]

R1 ¼lnðDhp;o=Dhp;inÞ

2pLhp;eKhpð26Þ

3.3.3.2. Thermal resistance of wick structure, R2. Inner surface of theheat pipe wall is attached with the mesh wick which causes certainresistance in heat transfer; this part of resistance can be written as[39]

R2 ¼lnðDhp;in=Dv;eÞ

2pLhp;eKwið27Þ

Kwi ¼Kl½ðKl þ KsÞ � ð1� ewiÞðKl � KsÞ�½ðKl þ KsÞ þ ð1� ewiÞðKl � KsÞ�

ð28Þ

ewi ¼ 1� 1:05pnwiDwi

4ð29Þ

3.3.3.3. Thermal resistance of vapour flow, R3. The vapour flow pro-cess from the evaporation section to condensing heat exchangerexperiences certain pressure loss and consequently temperaturedrop. This creates a resistance in heat transfer which could be writ-ten as [29]

R3 ¼t2vR0DPvNhp

Q uhfgPvð30Þ

DPv ¼ DPv;e þ DPv;f þ DPv;tl þ DPv;hx ð31Þ

(i) Pressure drop in the evaporator section

DPv;e ¼ �Q u

8qvðDv;e=2Þ4hfgNhp

ð32Þ

(ii) Pressure drop in the three-way fitting

DPv;f ¼ �4lvLf Q u

pqvðDv ;f =2Þ4hfgNhp

ð33Þ

(iii) Pressure drop in the vapour transportation line

Fig. 9. Indoor testing rigs: (a) PV/LHP module with single heat absorbing pipe; (b) PV/LHP module with double heat absorbing pipe; (c) aluminium-alloy based PV laminationwith no heat absorbing pipe (d) detachable glazing cover.

1238 X. Zhang et al. / Applied Energy 102 (2013) 1229–1245

DPv;tl ¼ �4lvLtlQu

pqvðDv;tl=2Þ4hfgNhp

ð34Þ

(iv) Pressure drop in the condensation section

DPv;c ¼4p2

Q u

8qvðDv;hx=2Þ4hfgNhp

1ðnhx=2Þ � 1

ð35Þ

3.3.3.4. Thermal resistance of condensed liquid film, R4. The con-densed liquid film will be evenly distributed on the surface ofthe condensing heat exchanger (heat pipe fluid side) and its asso-ciated flow resistance is [39]

R4 ¼ln½Dhx;in=ðDhx;in � 2dlf Þ�

2pLlf Klf ðnhx=2� 1Þ ð36Þ

3.3.3.5. Thermal resistance of heat exchanging plat, R5. The equiva-lent thermal resistance of heat exchanging plate is written as [39]

R5 ¼lnðDhx;o=Dhx;inÞ

2pðHhx=2ÞKhxðnhx=2� 1Þ ð37Þ

3.3.3.6. Thermal resistance of the heat pump refrigerant, R6. Therefrigerant within the heat pump cycle passes across the channelsof the condensing heat exchanger (refrigerant side) where it is

evaporated into vapour. This process involves the turbulent andforced convection heat transfer, and its associated thermal resis-tance is [39]

R6 ¼1

hrAhx;rðnhx=2Þ ð38Þ

hr ¼ hr;l ð1� xrÞ0:8 þ3:8x0:76

r ð1� xrÞ0:04

Pr0:38r

" #ð39Þ

hr;l ¼NurKr;l

Dhx;inð40Þ

The Nusselt number of the refrigerant flow could be expressedby Dittus–Boelter equation, as below [39]

Nur ¼ 0:023Re0:8r;l Pr0:4

r;l ð41Þ

The Reynolds number of the refrigerant within the channel is[39]

Rer;l ¼mrð1� xrÞDhx;in

lr;lðnhx=2ÞNhpð42Þ

The Prandtl number of refrigerant (Prr,l) is calculated by the fol-lowing equation [39]

X. Zhang et al. / Applied Energy 102 (2013) 1229–1245 1239

Prr;l ¼lr;lCp;lr

Kr;lð43Þ

3.4. Transferring the refrigerant received heat into tank water using aheat pump

The heat received by the refrigerant causes its evaporationwithin the condensing heat exchanger. This refrigerant vapour isthen upgraded through a compressor to a high temperature refrig-erant vapour, which, at the condenser of the heat pump cycle, iscondensed and releases heat to the tank water, resulting tempera-ture rise of the water and condensation of the refrigerant.

The heat received by the refrigerant is given as [40]

Q u ¼ Q e;t ¼ mrAcrðH1 � H4Þ ð44Þ

The operation of the compressor is assumed at isentropic condi-tion and the heat output of the heat pump could be expressed as[40]

Q c;t ¼ mrAcrðH2s � H3Þ ¼ mwCp;wðtiþ1w � ti

wÞ ð45Þ

The power input of the heat pump at the ideal isentropic oper-ating condition could be expressed as [40]

Q c;e ¼ mrAcrðH2s � H1Þ ð46Þ

The net electricity generation of the system is:

Q e;n ¼ Q e � Q c;e ð47Þ

Thus, work-back ratio of system, defined as the ratio of net elec-tricity output to the overall PV electricity generation, could be ex-pressed as

ge;n ¼Q e;n

Qeð48Þ

As such a PV/LHP system yields not only heat but also electric-ity, an overall coefficient indicating the thermal-and-electrical per-formance (COPPV/T) of the system is needed. This coefficientconverts the yielded electricity into the equivalent thermal energyusing the average electricity-generation efficiency (commonly 38%[18]) at a coal-fired power plant.

COPPV=T ¼Qc;t þ Q e=0:38

Qc;eð49Þ

If taking the compressor’s isentropic efficiency into account, thesystem’s overall performance coefficient would be an even lowerfigure which could be re-defined as [40]

Table 3List of experimental testing and monitoring devices.

Device Specification Quantity Location/applicat

Solar simulator Tungsten halogen lamps(300W)

8 Above PV/LHP m

Pyranometer TQB-2 (Jinzhou Sunlight,China)

1 Module bracket

Anemometer EC-8SX (Jinzhou Sunlight,China)

1 Near PV/LHP mod

Power sensor WB1919B35-S andWBP112S91 (Weibo, China)

2 PV module outpu

Flow meter R025S116N (MicroMotion,USA)

1 Compressor outle

Resistancetemperaturedetector (RTD)

PT100 RTD probes, (China) 10 PV backplane, LHliquid line, heat e

Thermometer Testo 605-H1 (Germany) 1 Room temperatuData logger DT500 (DataTaker,

Australia)1 Record data with

COPPV=T�exp ¼Q c;t�exp þ Qe�exp=0:38

Q c;e=gsð50Þ

Where the compressor’s isentropic efficiency is given [40] as

gs ¼Q c;e

Qc;e�exp¼ H2s � H1

H2 � H1ð51Þ

4. Algorithm applied for the computer programming

The heat transfer processes will eventually achieve an energybalance when the system operates at the steady state conditionand each part of the system will establish a certain temperaturewhen in operation. The algorithm used for modelling set-up isshown schematically in Fig. 8 and illustrated as follow:

(i) Giving a pre-set external weather condition, system designand operating parameters.

(ii) Assuming the refrigerant mass flow rate mr and the refriger-ant evaporating heat gain, calculating Qe,t using Eq. (44).

(iii) Assuming the cell temperature tp, and taking up the follow-ing heat analyses:

A. Heat balance of the glazing cover could be analyzed byEqs. (2)–(13), which results in determination of the heatloss, QL;

B. Heat balance of the PV cells could be analyzed usingEqs. (1), (14), (15), and (17), which results in determi-nation of the converted solar electricity, Qe and heat,Qth.

C. Heat transfer from the PV cells to the heat pump evapora-tor could be analysed using Eqs. (20)–(43), which resultsin determination of the useful heat gain, Qu.

D. If (Qth � Qu)/Qth > 0.5% (error allowance), then increasingtp by 0.1 �C and return to step (iii) for re-calculation.

E. If (Qth � Qu)/Qth < �0.5% (error allowance), thendecreasing tp by 0.1 �C and return to step (iii) for re-calculation.

F. If �0.5% 6 (Qth � Qu)/Qth 6 0.5%, the system is consideredto have achieved heat balance.

(iv) If (Qu � Qe,t)/Qu > 0.5% (error allowance), then increasing mr

by 0.001 kg/s and returning to step (ii) for re-calculation.(v) If (Qu � Qe,t)/Qu < �0.5% (error allowance), then decreasing

mr by 0.001 kg/s and returning to step (iii) for re-calculation.

(vi) If �0.5% 6 (Qu � Qe,t)/Qu 6 0.5%, the system is considered tohave achieved heat balance.

ion

odule

ule

t (DC), compressor input (AC)

t

P evaporator section, vapour line, heat exchanger inlet/outlet (heat pipe side),xchanger inlet/outlet (refrigerant side), water tank

re and humiditycomputing unit

Table 4List of experimental testing modes.

Testingmode

Solar radiation (W/m2)

Air temperature (�C) Wind speed (m/s)

Heat pump evaporation temperature(�C)

Cover sheetnumber

Heat pipenumber

(1) 200 ± 20, 300 ± 20 20 ± 2 1 ± 0.2 10 1 1400 ± 20, 500 ± 20600 ± 20

(2) 600 ± 20 10 ± 2,15 ± 2, 20 ± 2,25 ± 2,

1 ± 0.2 10 1 1

(3) 600 ± 20 20 ± 2 1 ± 0.2, 3 ± 0.2 10 1 15 ± 0.2, 7 ± 0.2

(4) 600 ± 20 20 ± 2 1 ± 0.2 5, 10, 15, 20 1 1(5) 600 ± 20 20 ± 2 1 ± 0.2 10 0, 1, 2 1(6) 600 ± 20 20 ± 2 1 ± 0.2 10 1 0, 1, 2

1240 X. Zhang et al. / Applied Energy 102 (2013) 1229–1245

(vii) Calculating the module’s energy efficiencies, system work-back ratio and the overall performance coefficient of COPPV/

T using Eqs. (16), (18), (19), (45)–(50), (and) (51).(viii) Finally determining tp and stopping the program.

5. Experimental testing

5.1. Experimental rig set up

A prototype PV/LHP heat pump system was constructed andtested under the controlled indoor condition at Shanghai, China.The PV/LHP module, with an effective absorbing area of0.612 m2, was fixed to the 30� tilted frame, and fitted with thedetachable double/single glazing cover on top. The PV cells, con-sisting of totally 36 (4 � 9 array) pieces each with sizes of125 � 125 � 0.3 (mm �mm �mm), occupied nearly 90% of theabsorbing surface. Table 1 presents the values of the characteristic

200 300 400 500 600

31.2

35.1

39.0

42.9

10.14

10.40

10.66

10.92

32.4

35.1

37.8

40.5

8.61

9.02

9.43

9.84

t b (o C

)

(exp) (sim)r=0.991; e=3.28%

Solar radiation (w/m2)

η e (%

)

(exp) (sim)r=0.984; e=2.04%

η th (

%)

(exp) (sim)r=0.982; e=4.43%

CO

PP

V/T

(exp) (sim)r=0.979; e=3.02%

Fig. 10. Temperature at PV backplane, module efficiencies and COPPV/T as a functionof solar radiation.

parameters relating to the PV cells under the standard testing con-ditions. During the PV module making-up, a black 5052 aluminiumalloy sheet coated with 20 lm anodic oxidation film was used toreplace the conventional TPT base-board of the PV cells. A 5 mmthick aluminium X-type fin sheet, embracing a wicked pipe (with160 � 60 copper meshes), was adhered to the PV base-board usingthe silicon sealants. This pipe, when connecting to the liquid andvapour transportation lines and condensing heat exchanger,formed up a loop that was evacuated and then filled with 75 mlof water/glycol mixture (95%/5%) as the working fluid. The specifi-cations of the loop components including tube, fins, three-way fit-ting, liquid/vapour transportation lines and condensing heatexchanger are given in Table 2. Further, the system also employeda 1 kW rating heat pump cycle with the evaporation/condensationtemperatures of 10 �C/55 �C, which was charged with 300 g ofR134a refrigerant. A 100 l of water tank with built-in cooper heatexchanging coils was also installed and connected to the heat

10 15 20 25

39.2

40.6

42.0

43.4

10.08

10.22

10.36

10.50

31.2

35.1

39.0

42.9

8.28

8.74

9.20

9.66

(exp) (sim)r=0.985; e=3.29%

Air temperature (oC)

(exp) (sim) r=0.976; e=1.96%

(exp) (sim)r=0.996; e=4.64%

(exp) (sim)r=0.994; e=3.03%

t b (o C

)η e

(%)

η th (

%)

CO

PP

V/T

Fig. 11. Temperature at PV backplane, module efficiencies and COPPV/T as a functionof air temperature.

1 2 3 4 5 6 7

40.05

40.94

41.83

42.72

10.08

10.20

10.32

10.44

35.7

37.4

39.1

40.8

8.64

8.82

9.00

9.18

(exp) (sim)r=0.991; e=3.37%

Wind speed (m/s)

(exp) (sim)r=0.970; e=1.81%

(exp) (sim)r=0.992; e=4.53%

(exp) (sim)r=0.90; e=2.97%

t b (o C

)η e

(%)

η th (

%)

CO

PP

V/T

Fig. 12. Temperature at PV backplane, module efficiencies and COPPV/T as a functionof wind speed.

5 10 15 20

36

42

48

54

9.72

9.99

10.26

10.53

30.1

34.4

38.7

43.0

7.0

8.4

9.8

11.2

(exp) (sim)r=0.999; e=4.81%

Evaporation temperature (oC)

(exp) (sim)r=0.979; e=1.56%

(exp) (sim)r=0.999; e=5.07%

(exp) (sim)r=0.999; e=2.95%

t b (o C

)η e

(%)

η th (

%)

CO

PP

V/T

Fig. 13. Temperature at PV backplane, module efficiencies and COPPV/T as a functionof heat pump evaporation temperature.

X. Zhang et al. / Applied Energy 102 (2013) 1229–1245 1241

pump to obtain heat and store the heating water. The electricalparts of the system include a 12 V (10 A) controller, 500 W DC/AC inverter, a 100AH (12 V) battery, and the connection wires. Sev-eral insulation materials including the foamy polyurethane for pip-ing and polystyrene board for exchangers were also applied tominimise the heat loss of the system components. The image ofthe indoor experimental rig is presented in Fig. 9; a list of the test-ing instruments and devices are presented in Table 3.

5.2. Experimental processing

To enable characterization of the PV/LHP heat pump system,numerous sets of experiments, as shown in Table 4, were con-ducted over the duration from 8th November to 9th December2011 under the controlled indoor conditions. Each of the testsstarted at 9:00 am and run for sufficient hours to obtain the steadystate data. During the tests, the acquired data were recorded at the10-s interval.

5.3. Statistical analysis

To analyse the difference between the theoretical and experi-mental results, the correlation coefficient (r) and a root meansquare percentage deviation (e) were brought into use and theseparameters are defined as

r ¼ nP

XeXs � ðP

XeÞðP

XsÞffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffinP

X2e � ð

PXeÞ2

q ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffinP

X2s � ð

PXsÞ2

q ð52Þ

e ¼

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiP½100� ðXe � XsÞ=Xe�2

n

sð53Þ

6. Results and discussion

Using both established computer model and testing rig, para-metric studies of the PV/LHP heat pump system were conductedunder the pre-set operational conditions. This allows parallel com-parison of the modelling and test results of the system, thus lead-ing to verification or modification of the model accuracy inperformance prediction. Further, impact of the operational param-eters (e.g., solar radiation, air temperature, air velocity, heat-pump’s evaporation temperature, glazing cover, and number ofheat absorbing pipes) to the system performance (e.g. efficienciesof the PV/LHP module and system COPPV/T) was investigated indi-vidually. Finally, a brief comparison of such a system against theconventional solar/air energy systems was made. The results areillustrated as follows.

6.1. Impact of solar radiation

Varying the solar radiation from 200 to 600 W/m2 whileremaining other parameters constant, i.e., air temperature at20 �C, wind speed at 1 m/s, heat pump evaporating temperatureat 10 �C, single glazing cover and fixed single heat absorbing pipe,as indicated in Table 4 for mode 1 operation, simulation was car-ried out using the established computer programme. The simula-tion results were then compared with the testing resultsobtained under the same operational condition, i.e., mode (1) ad-dressed above, which yielded a comparing diagram of Fig. 10. Goodagreement between the modelling and experimental results wasobserved with the correlation coefficient and root mean squarepercentage deviation of 0.984 and 2.04% for the electrical effi-ciency, and 0.982 and 4.43% for the thermal efficiency. The minor

0 1 2

36.8

39.1

41.4

43.7

9.68

10.56

11.44

12.32

28.5

34.2

39.9

45.6

8.55

9.50

10.45

11.40

(exp) (sim)r=0.999; e=1.55%

Number of cover sheets

(exp) (sim) r=0.999; e=2.08%

(exp) (sim)r=0.999; e=2.37%

(exp) (sim)r=0.999; e=1.76%

t b (o C

)η e

(%)

η th (

%)

CO

PP

V/T

Fig. 14. Temperature at PV backplane, module efficiencies and COPPV/T as a functionof cover sheets.

0 1 2

27.6

36.8

46.0

55.2

9.40

9.87

10.34

10.81

0

15

30

45

8.75

9.10

9.45

9.80

(exp) (sim)r=0.999; e=1.80%

Number of heat absorbing pipes

(exp) (sim) r=0.997; e=1.46%

(exp) (sim)r=0.999; e=2.46%

(exp) (sim)r=0.999; e=2.03%

t b (o C

)η e

(%)

η th (

%)

CO

PP

V/T

Fig. 15. Temperature at PV backplane and module efficiencies as a function of heatpipe numbers.

1242 X. Zhang et al. / Applied Energy 102 (2013) 1229–1245

differences in existence may be caused by the simplified assump-tions, utilization of empirical formulas, measurement errors, andin particular, the compressor’s isentropic efficiency (less than88%), which caused slightly reduced practical performance coeffi-cient over the theoretical figure. It is found that increasing the solarradiation led to significant increase in temperature of PV absorber(from 31 �C to 41.8 �C) and in the module’s thermal efficiency(from 32.4% to 39.6%), and slight decreases in the module’s electri-cal efficiency (from 10.8% to 10.1%) and in the system’s overall per-formance coefficient (from 9.6 to 8.7). The phenomena could beexplained as follows: a higher solar radiation yielded an enhancedsolar heat transfer, which helped improve the solar heat gain andthe module’s thermal efficiency. Meanwhile, the heat gain wasaccumulated on the PV modules owing to the established thermalresistance of the LHP cycle; this led to increase in PV cells’ temper-ature and decrease in the module’s solar electrical efficiency. To re-move this amount of heat, the heat pump would operate at higherelectrical power condition, thus leading to reduced system overallperformance coefficient.

6.2. Impact of surrounding air temperature

Varying the surrounding air temperature from 10 to 25 �C whileremaining other parameters constant, as indicated in Table 4 formode 2 operation, simulation was carried out using the establishedcomputer programme, and the results of the simulation were thenput into parallel comparison against the experimental data, thusgiving a comparing image as shown in Fig. 11. Good agreementwas found between these two sets of results, giving the correlationcoefficient and root mean square percentage deviation of 0.976 and

1.96% for the electrical efficiency, and 0.996 and 4.64% for the ther-mal efficiency. It is found that increasing the surrounding air tem-perature resulted in increase in PVs’ temperature (from 38.7 �C to42.3 �C) and in associated thermal efficiency (from 31.5% to42.4%), and decrease in the module’s electrical efficiency (from10.3% to 10.1%) and in system’s overall performance coefficient(from 9.3 to 8.3). The phenomena could be explained as follows:the higher surrounding air temperature reduced the modules’ heatloss and thus increased their useful heat gain, resulting in increasein the system’s thermal efficiency. This, however, also led to in-creased PV temperature, resulting in reduced system electrical effi-ciency and net electricity generation. Consequently, the system’soverall performance coefficient somehow fell.

6.3. Impact of air velocity

Varying the surrounding air velocity from 1 to 7 m/s whileremaining other parameters constant, as indicated in Table 4 formode 3 operation, simulation was carried out using the establishedcomputer programme, and the results of the simulation were thenput into parallel comparison against the experimental data, givinga comparing image as shown in Fig. 12. The correlation coefficientand root mean square percentage deviation were found to be 0.970and 1.81% for the electrical efficiency and 0.992 and 4.53% for thethermal efficiency, indicating that good agreement between thesimulation and experiment has been achieved. It is found thatincreasing the air velocity led to slight decrease in temperatureof PV absorber (from 41.8 �C to 39.9 �C) and in the module’s ther-mal efficiency (from 39.6% to 35.5%), and increase in the module’selectrical efficiency (from 10.34% to 10.39%) and in the system’soverall performance coefficient (from 8.7 to 8.9). The phenomena

PV/LHP module (sim)

PV/LHP module (exp)

Thermal collector [42]

PV panels [41]

(a)

Average efficiency (%)

Thermal efficiency

Electrical efficiency

PV/LHP-HP (sim)

PV/LHP-HP (exp)

ISAHP [44]

ASHP [43]

0 5 10 15 20 25 30 35 40 45 50 55

0 2 4 6 8

(b)

System COP

System COP

Fig. 16. Comparison between the PV/LHP heat pump system and (a) independent PV panels, standard solar thermal collectors and (b) conventional air-source and solar-assisted heat pump water heating systems.

X. Zhang et al. / Applied Energy 102 (2013) 1229–1245 1243

could be explained as follows: the higher air velocity caused in-creased heat loss and reduced thermal efficiency. It, however, alsolowered temperature of PV modules, leading to the increase in themodules’ electrical efficiency. Consequently, the system’s overallperformance coefficient was higher.

6.4. Impact of the heat-pump’s evaporation temperature

Varying the heat-pump’s evaporation temperature from 5 to20 �C while remaining other parameters constant, as indicated inTable 4 for mode 4 operation, simulation was carried out usingthe established computer programme, and the results of the simu-lation were then put into parallel comparison against the experi-mental data, thus giving a comparing image as shown in Fig. 13.The correlation coefficient and root mean square percentage devi-ation were found to be 0.979 and 1.56% for the electrical efficiencyand 0.999 and 5.07% for the thermal efficiency, indicating that agood agreement has been achieved between the simulation andexperiment. It is found that increasing the evaporation tempera-ture led to increase in temperature of PV absorber (from 38.9 �Cto 52.1 �C) and in the system’s overall performance coefficient(from 7.6 to 11.6), and decrease in the modules’ electrical effi-ciency (from 10.4% to 9.7%) and in the modules’ thermal efficiency

(from 43% to 30%). The phenomena could be explained as follows:higher evaporation temperature reduced the temperature differ-ence between the PV/LHP absorber and condensing heat exchan-ger, thus resulting in the reduced heat transfer across the LHPand reduced modules’ thermal and electrical efficiency. It, how-ever, lowered the temperature difference between condensationand evaporation processes, which led to reduced compressorpower in the heat pump operation. As a result, the system’s overallperformance coefficient was higher.

6.5. Impact of the top glazing covers

Varying the layer of the top glazing covers from 0 to 2 whileremaining other parameters constant, as indicated in Table 4 formode 5 operation, simulation was carried out using the establishedcomputer programme, and the results of the simulation were thenput into parallel comparison against the experimental data, thusgiving a comparing image as shown in Fig. 14.

It is found that increasing number of the glazing covers led to in-crease in the PV-absorber temperature and in the thermal efficiency,and decrease in the electrical efficiency and in the system’s overallperformance coefficient. This is because that adding more glazingcovers helped reduce the overall heat losses and the amount of ab-

1244 X. Zhang et al. / Applied Energy 102 (2013) 1229–1245

sorbed solar energy due to its reflection and reduced transmittance;as a result, the modules’ thermal efficiency and the PV-absorbertemperature arose, and the modules’ electrical efficiency and thesystem’s overall performance coefficient fell. To minimizing heatloss and maximizing solar energy intake, the single glazing coverwas considered to be the most appropriate option.

6.6. Impact of heat absorbing pipe numbers

Varying the number of the heat absorbing pipes behind the PVlayer from 0 to 2 while remaining other parameters constant, asindicated in Table 4 for mode 6 operation, simulation was carriedout using the established computer programme, and the resultsof the simulation were then put into parallel comparison againstthe experimental data, thus giving a comparing image as shownin Fig. 15. It is found that increasing number of the heat absorbingpipes led to increase in all indicative parameters of the systemincluding thermal and electrical efficiencies, and COPPV/T. Applyingmore heat absorbing pipes underneath the PV laminations in-creased the standard fin efficiency, which enabled reduced PV sur-face temperature, increased solar heat gain, and enhanced thermaland electrical efficiencies; and consequently, the overall perfor-mance of the system was enhanced.

6.7. Comparison between the PV/LHP heat pump system andconventional PV panels, solar thermal collectors and solar/air heatpump water heating systems

Given a fixed operating condition, i.e. solar radiation of 600 W/m2, air temperature of 20 �C, wind speed of 1 m/s, heat pump evap-orating temperature of 10 �C, single glazing cover and single heatabsorbing pipe, simulations were carried out to enable compara-tive study of the performance of the above addressed systems, thusgiving the comparing image of Fig. 16. It is found that the electricaland thermal efficiency of the PV/LHP module were around 10% and40% respectively, giving 50% of overall efficiency which is muchhigher than the average values for the PV panels (around 10–12%[41]) and solar thermal collectors (around 40% [42]). The overallperformance coefficient of the system was about 8.7, which wasnearly fourfold of the conventional air-source heat pump waterheating system (ASHP) [43], and twice of the integral-type solar-assisted heat pump system (ISAHP) [44]. The work-back ratio ofthe system was around 15%, which implied that 85% of the electric-ity generated by the PV was used for powering the heat pumpoperation, while remaining could be utilized for other purposesin the building, or exported to the grid.

7. Conclusion

This paper introduced a novel solar photovoltaic/loop-heat-pipeheat pump water heating system, which has potential to overcomethe difficulties remaining in the existing PV/thermal technologiesand is expected to be low cost, building integrating, highly effi-ciency and aesthetically appealing. A computer model was devel-oped to simulate the performance of the PV/LHP heat pumpsystem on the basis of heat balances mechanism, which gave thepredicted values of PV surface temperature, PV modules’ solarthermal and electrical efficiencies, and the system’s overall perfor-mance coefficient (COPPV/T) at the specified operational conditions(e.g. solar radiation, air temperature, air velocity, heat pump evap-oration temperature, glazing covers, and absorbing heat pipe num-bers). A prototype solar PV/LHP heat pump water heating systemwas constructed, simulated and tested to examine its characteristicperformance under the fixed lab conditions. Parallel comparisonbetween the modelling and experimental results indicated that

these two sets of data were in good agreement and therefore, theestablished model was able to predict the system performance ata reasonable accuracy (average error less than 5%).

Relations between the system’s characteristic parameters (ther-mal, electrical efficiencies and COPPV/T) and operational/geometri-cal conditions were individually studied under different testingmodes. It is concluded that: (1) increasing the solar radiation ledto increase in the module’s thermal efficiency, and decrease inthe module’s electrical efficiency and in the system’s overall per-formance coefficient (COPPV/T); (2) increasing the surrounding airtemperature resulted in increase in the module’s thermal effi-ciency, and decrease in the module’s electrical efficiency and inthe system’s overall performance coefficient (COPPV/T); (3) increas-ing the surrounding air velocity led to slight decreases in the mod-ule’s thermal efficiency, and increase in the module’s electricalefficiency and in the system’s overall performance coefficient(COPPV/T); (4) increasing the heat pump’s evaporation temperatureled to decrease in the module’s electrical and thermal efficienciesand increase in the system’s overall performance coefficient(COPPV/T); (5) increasing number of the glazing covers led to in-crease in the module’s thermal efficiency and decrease in the mod-ule’s electrical efficiency and in the system’s overall performancecoefficient (COPPV/T); (6) increasing number of the heat absorbingpipes led to increase in the fin’s efficiency and in the system’s over-all performance coefficient. To achieve the better operational per-formance for the PV/LHP heat pump system, construction of thePV/LHP module should be made by (1) using a single glazing asthe cover of the module; (2) fixing two heat absorbing pipes under-neath the PV layer. During the operation of the system, the evapo-ration temperature of the heat pump is suggested to set to 5–10 �C.Further, the system is found to be able to obtain better perfor-mance at the moderate solar radiation (e.g., 400–600 W/m2), andmild surrounding air temperature (e.g., 15–20 �C), and lower airvelocity (e.g., 0–1 m/s).

The electrical, thermal and overall efficiency of the PV/LHPmodule at the given laboratory conditions were around 10%, 40%and 50% respectively. This hybrid technology enables the enhancedoverall solar conversion ratio over the independent solar photovol-taic panel and the typical solar thermal collector. The overall coef-ficient of system performance was measured at about 8.7, which isnearly two to four times higher than that for the conventional so-lar/air heat pump water heating systems. Operation of the heatpump system was proven to self-sustainable consuming around85% of PV generated electricity, while 15% of the PV generationcould be exported.

In overall, the research provided a method to determine thecharacteristic parameters of such a new PV/LHP heat pump systemand give useful clues on how to generate the best possible systemperformance in terms of the better geometrical settings andfavourite operational conditions. It obviously helps develop a solardriven (space or hot water) heating system with enhanced effi-ciency over conventional solar heating systems and thus contrib-ute to realisation of the energy saving and associated carbonemission targets set for buildings globally.

Acknowledgements

The authors would acknowledge our sincere appreciation to thefinancial supports from the De Montfort University, Shanghai Paci-fic Energy Centre, and EU Marie Curie International Research StaffExchange Scheme (R-D-SBES-R-269205).

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