Comparison of energy performance of heat pumps using a photovoltaic thermal evaporator with circular and triangular tube configurations

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BUILD SIMUL DOI 10.1007/s12273-015-0256-1 Comparison of energy performance of heat pumps using a photovoltaic thermal evaporator with circular and triangular tube configurations M. Mohanraj 1 ( ), N. Gunasekar 2, V. Velmurugan 3 1. Department of Mechanical Engineering, Hindusthan College of Engineering and Technology, Coimbatore-641032, India 2. Department of Mechanical Engineering, Sri Krishna College of Technology, Coimbatore-641042, India 3. Department of Mechanical Engineering, Sri Sakthi Engineering College, Coimbatore-641104, India Research Article Abstract In this work, the energy performance parameters of heat pumps using a photovoltaic thermal evaporator (PV TE) with circular and triangular tube evaporator configurations are compared. The experimental observations were made in a heat pump under the meteorological conditions of Coimbatore city in India. The standard energy performance parameters of a heat pump (using PV TE) such as, condenser heating capacity, coefficient of performance (COP), evaporator heat gain, solar energy input ratio and photovoltaic panel efficiency were calculated using experimental observations. The energy performance parameters of the heat pump were simulated using artificial neural networks to have a realistic energy performance comparison. The results show that triangular tube PV TE configuration has enhanced the major energy performance parameters of the heat pump such as, condenser heating capacity, COP and panel efficiency by 3% 7%, 3% 5% and 4% 13%, respectively when compared to the circular tube PV TE configuration. Keywords photovoltaic thermal evaporators, triangular tube configurations, heat pumps Article History Received: 27 May 2015 Revised: 12 August 2015 Accepted: 21 September 2015 Tsinghua University Press and Springer-Verlag Berlin Heidelberg 2015 1 Introduction The heat pump is an energy efficient device used for space heating, drying, water heating and desalination applications (Amin and Hawlader 2013). The performance of the heat pump can be improved by integrating it with renewable energy sources such as solar thermal, geothermal, solar thermal photovoltaic hybrid, solar ambient hybrid and solar geothermal hybrid sources (Ozgener and Hepbasli 2007). Out of all these five options, solar photovoltaic thermal hybrid system is identified as the feasible option for the small capacity direct expansion heat pumps, which provides both electrical and heat energy sources (Mastrullo and Renno 2010). The photovoltaic thermal evaporator (PV TE) is directly integrated with the heat pumps, where the refrigerant undergoes phase change from liquid to vapour by absorbing the heat from the panels (Ji et al. 2007). Many research investigations have been reported with different configurations of PV TE for heat pump applications. Ji et al. (2007) have developed and investigated the energy performance of a heat pump using circular tube PV TE and reported with a maximum COP of 10.0 and photovoltaic efficiency of 13.4%. In a similar work, Liu et al. (2009) theoretically simulated the energy performance of a heat pump using a PV TE with circular tube configuration. The COP of the heat pump was varied between 4.6 and 7.2 with an average value of 6.0. In another work, Xu et al. (2009) theoretically simulated the energy performance of a heat pump using multiport flat extruded aluminum tube PV TE. Their results concluded that, the modified PV TE has about 7% improved COP compared to the heat pump using conventional PV TE. In a further work, the energy performance of a heat pump was improved by PV TE using parabolic concentrators (Xu et al. 2011). Their results reported with an average COP of about 4.8. Similarly, a micro photovoltaic panel covered with vacuum glass tube was developed for a heat pump and investigated its energy performance (Chen et al. 2011). The COP of their system was varied between 2.9 and 4.6 with an increase in solar radiation. Further, the performance of a heat pump was improved using a heat pipe-assisted PV TE (Fu et al. 2012). Their results reported that the heat pipe-assisted PV TE Building Systems and Components E-mail: mohanrajrac@yahoo.co.in

2 Mohanraj et al. / Building Simulation List of symbols Clv climatic variations c p specific heat at constant pressure (J/(kg K)) h specific enthalpy (kj/kg) m a mass flow rate of air (kg/s) PV OP output power (W) PR performance rating of photovoltaic panel Q thermal capacity (W) R resistance T temperature ( C) W compressor power consumption (W) comp Subscripts 1 condition of refrigerant at the compressor suction 2 condition of refrigerant at the compressor discharge 3 condition of refrigerant at the condenser outlet 4 condition of refrigerant at the expansion valve outlet 5 condition of air at the condenser inlet 6 condition of air at the condenser outlet a air cao condenser air outlet cai condenser air inlet comp compressor cond condenser Ev OP p pl pv r Abbreviations ASHRAE ANN CGP COV COP DC LM MLFFN PV TE RMS SCG SEIR SPV THP evaporator output pressure plate photovoltaics refrigerant American Society for Heating, Refrigeration and Air-conditioning Engineers artificial neural network Pola Ribiere conjugate gradient coefficient of variance coefficient of performance direct current Levenberg Marquardt multilayer feed forward network photovoltaic thermal evaporator root mean square scaled conjugate gradient solar energy input ratio solar photovoltaic thermal heat pump has higher energy efficiency in the range between 62% and 82%. Zhang et al. (2013) tested the performance of a heat pump using heat pipe-assisted PV TE for water heating applications and reported with an average COP of about 5.5. The circular tubes used in the reported investigations have tangential thermal contact with the absorber plate. The tangential thermal contact has less heat absorption. Hence, the triangular configuration is proposed in this research work to improve the rate of heat absorption by increasing the thermal contact. The cited literature confirmed that there is no specific work reported on triangular PV TE configuration. The thermodynamic performance simulation of heat pumps at different climatic conditions is highly essential to have a realistic energy performance comparison when the system is operating with two different PV TE configurations. The theoretical simulations involve complicated equations and more assumptions, which are not strictly valid for the triangular tube configurations. Alternatively, the experiments with heat pumps using two PV TE configurations under similar ambient conditions are not possible. The ANN energy performance simulation overcomes these limitations by extracting the required information as training data, which has not required the system descriptions (Mohanraj et al. 2012, 2015). The MLFFN was successfully used for predicting the energy performance of direct expansion solar assisted heat pumps with reference to solar intensity and ambient temperatures (Mohanraj et al. 2008, 2009b,c). Further, the exergy performance of the solar-assisted heat pump was predicted using MLFFN (Mohanraj et al. 2009d, 2010). The influence of relative humidity and wind velocity was ignored in the reported investigations due to the presence of glazing surface over the absorber plate. However, the ambient wind velocity and relative humidity are also affecting the energy performance of the PV TE (Mekhilef et al. 2012). Hence, in this work, the energy performance of a heat pump was evaluated with reference to the four major ambient parameters such as, solar intensity, ambient temperature, ambient wind velocity and relative humidity. The cited literature confirmed that many research investigations have been reported with the energy performance of heat pumps using different PV TE configurations. However, there is no specific work reported on the application of triangular tube PV TE configuration for heat pumps. In this research work, the triangular tube PV TE was developed to increase the thermal contact of refrigerant with the absorber plate. The energy performance of the heat pump was evaluated using circular and triangular tube configurations. Two ANN models with four neurons in input layer (representing solar

Mohanraj et al. / Building Simulation 3 intensity, ambient temperature, ambient relative humidity and ambient wind velocity) and four neurons in output layer (representing energy performance parameters of heat pump and PV TE) were developed to simulate the energy performance of the heat pump and the PV TE for different climatic conditions to have realistic energy performance comparison. The ANN models were trained with experimental results observed in a heat pump. 2 Experiments The experimental observations were made in a heat pump under the meteorological conditions of Coimbatore city (latitude of 10.98 N and longitude of 76.96 E) in India during January 2014 to April 2014. 2.1 Experimental setup The schematic and photographic views of a heat pump are depicted in Fig. 1 and Fig. 2, respectively. The specifications are listed in Table 1. The heat pump consists of a R134a based hermetically sealed reciprocating compressor, a finned tube forced convection air cooled condenser, a liquid receiver, a sealed type refrigerant drier, a sight glass, a capillary tube expansion device and a PV TE of an area 1.92 m 2 (1.96 m 0.98 m) with integral suction line accumulator. The PV TE was designed to take the maximum heat load of 2.9 kw from the solar and ambient sources by maintaining the photovoltaic panel temperature around 25 C. The heat load acting on the PV TE varies with respect to ambient conditions and heat pump operating temperatures. The pictorial view of the PV TE is depicted in Fig. 3(a). The PV TE has 1.2 mm thick copper plate attached with 9.52 mm diameter circular tubes and 12 mm equilateral triangular tubes as shown in Fig. 3(b). From Fig. 3(b), it is confirmed that the refrigerant passing through the triangular tube configuration has more thermal contact with the absorber plate, which is capable of absorbing more thermal energy when compared to the circular tube configurations. The top surface of the absorber plate is covered with a photovoltaic panel. Bottom side of the PV TE was covered with 10 mm thick poly euro ethylene foam to reduce the heat interactions. The PV TE is tilted to an angle of about 20 with respect to the horizontal and oriented towards south to maximize the solar radiation incident on the collector (Shariah et al. 2002). The direct current (DC) output of the photovoltaic panel was connected to a DC load, while the condenser heat output can be utilized for heating applications. 2.2 Instrumentation The refrigerant temperatures at typical locations in a heat pump circuit (compressor suction, compressor discharge, condenser outlet and expansion valve outlet) were measured by four calibrated PT-100 temperature sensors having ±0.2 C accuracy. These temperature sensors were placed inside the insulated thermometer pockets to avoid the ambient heat interactions. Air temperatures at the entry and exit of the condenser and photovoltaic panel temperatures at three locations (as shown in Fig. 3) were measured by PT-100 temperature sensors with the same specifications and accuracy. All the temperature sensors are connected to the digital temperature indicator with 0.1 C resolution. Four Bourdon tube pressure gauges with ±2% accuracy were used to measure the refrigerant pressure at salient points in the refrigerant Fig. 1 Schematic diagram of experimental setup

4 Mohanraj et al. / Building Simulation Fig. 3 Sectional view of PV TE Fig. 2 Photographic view of experimental setup Table 1 Specifications of the SPV THP used for conducting experiments Component Compressor Condenser Expansion device Solar collector Specification Hermetically sealed reciprocating type, rated power input: 900 W, speed: 2800 rpm, protection against overload Finned tube air cooled type with 9.52 mm diameter copper tube Capillary tube Solar photovoltaic evaporator with circular and triangular tube configuration Photovoltaic panel Poly crystalline of total area 1.92 m 2 Refrigerant R134a loop. The instantaneous compressor power consumption was measured with the help of a digital Wattmeter having ±2 W accuracy. The ambient temperature was measured with the help of thermometer with ±0.3 C accuracy. The ambient relative humidity was measured with a help of sling psychrometer (two thermometers with ±0.3 C accuracy, one covered with wet wick and another with bare bulb). From the measured wet and dry bulb temperatures, the relative humidity was calculated using standard psychrometric chart. The photovoltaic output was measured using a multimeter. A pyranometer with ±5 W/m 2 accuracy was placed near the PV TE to measure the instantaneous solar intensity falls on the PV TE. The ambient wind velocity was measured with the help of cup type anemometer with ±0.2 m/s accuracy. The airflow rate through the condenser was measured with the help of orifice meter and manometer setup. The specifications of the instruments used in the experiments are consolidated in Table 2. The high pressure cutout was installed in the compressor discharge line to control the compressor discharge pressure within 13 bar. Similarly, the low pressure control was installed in the compressor suction to maintain the refrigerant quantity above 650 grams inside the heat pump. 2.3 Experimental procedure The moisture and foreign materials present inside the heat pump were flushed and leak tested with nitrogen gas. After the leak test, the heat pump is subjected to vacuum with the Table 2 Specifications of the measuring instruments Instrument Range Accuracy Pressure gauges 0 500 psi ±2% Temperature sensors PT-100 (0 200 C) ±0.2 C Wattmeter Digital type ±2 W Pyranometer 0 1200 W/m 2 ±5 W/m 2 Anemometer 0 20 m/s 0.2 m/s Sling psychrometer 0 100 C 0.3 C

Mohanraj et al. / Building Simulation 5 help of vacuum pump. The ozone friendly refrigerant R134a with good thermodynamic and thermo physical properties is used as a working fluid in this work (Mohanraj et al. 2009a). The refrigerant mass charge for the system using circular tube configuration was optimized by conducting the experiments with three different mass charges of 700 g, 750 g and 800 g of R134a during off sunshine hours. The COP of the system working with 700 g, 750 g and 800 g of refrigerant are 2.45, 2.6 and 2.5, respectively. A maximum COP of 2.6 was observed for a 750 g mass charge. Hence, 750 g of R134a was charged in the system when using circular tube PV TE configuration. Initially, the heat pump was connected with circular tube PV TE configuration. The experimental observations were made only during 9.00 to 17.00. Before starting the experimental observations, the heat pump was warmed up to avoid early transients. The surface of the PV TE was cleaned with a soft cloth to remove the dust accumulations and kept the surface clean and dry. During experimentation, the temperature and pressure of the refrigerant at salient points in a heat pump circuit and air temperatures at the entry and exit of the condenser were recorded manually for every 30 minutes interval. Similarly, the temperature of photovoltaic panel, photovoltaic outputs were recorded every 30 minutes. The solar intensity, ambient temperature, ambient relative humidity, ambient wind velocity and instantaneous compressor power consumption were measured at the same time interval. The airflow rate was maintained constant, which was measured using an orifice meter manometer setup. The experimental observations were made for the period of sixty days. More frequent experimental data observations are useful to study the nature of variations and to avoid erroneous data. After experimentation with circular tube PV TE, the PV TE configuration was changed to triangular tube mode with the help of control valve. The volume of PV TE using triangular configuration is about 5% higher when compared to the circular tube PV TE, which demands high refrigerant quantity. Hence, the refrigerant quantity was optimized for triangular configuration by conducting experiments with 700 g, 750 g, 800 g and 850 g mass charges during off sunshine hours. A maximum COP of 2.6 was observed for 800 g mass charge for triangular tube configuration. Hence, the refrigerant mass charge can be increased to 800 g during the experimentation with triangular tube configuration. The experimental observations were repeated with triangular configurations and the observations were made similar to the circular PV TE configuration. 3 Energy analysis of a heat pump The energy analysis based on the first law of thermodynamics is used to investigate the energy performance of a heat pump system. For steady state and steady flow processes, the mass and energy balance equations are used to find out the amount of heat input and the energy efficiency. The general mass balance equation can be expressed by the following equation: å m = m (1) in å out Here, m is the mass flow rate, the subscripts in and out stands for inlet and outlet of the particular component, respectively. The mass flow rate of the refrigerant is assumed to be constant at all the typical locations in the heat pump circuit. The general energy balance equation of a heat pump based on first law of thermodynamics is given by the following equation (Li et al. 2007): Qcond = Qe + W comp (2) Here, Q cond (W) is the condenser heating capacity, Q e (W) is the amount of heat gain in the evaporator, W comp is the compressor power consumption in (W). 3.1 Condenser heating capacity The condenser heating capacity (Q cond ) of a heat pump is estimated by the following equation: Q = m c ( T -T ) (3) cond a p cao cai Here, m a is the mass flow rate of air passing through the condenser coil (kg/s), c p is the specific heat of air (J/(kg K)), T cao temperature of condenser air outlet ( C), and T cai is the temperature of the condenser air inlet ( C). 3.2 Compressor power consumption The instantaneous compressor power consumption is theoretically calculated by the following equation (Kara et al. 2008): m ( h - h ) r 2 1 comp = (4) ηmech ηelec W Here, η mech is the compressor mechanical efficiency, η elec is the compressor electrical efficiency. h 1 and h 2 (in kj/kg) are the enthalpy of the refrigerant before and after leaving the compressor, respectively. The mechanical and electrical efficiency of the compressor are assumed as 0.85 and 0.9, respectively. The compressor power consumption was also measured using a Wattmeter. Both measured and theoretically calculated values are closer to each other. The refrigerant mass flow rate is calculated using the following equation (Kara et al. 2008): V η N dis vol m r = (5) 60v1

6 Mohanraj et al. / Building Simulation Here, v is the specific volume of the refrigerant (m 3 /kg) at the compressor suction, V dis is the volumetric displacement of compressor: 18.27 cm 3 /rev (based on compressor manufacturer catalogue), N is the compressor operating speed: 2800 rpm (based on compressor manufacturer catalogue), η vol is the volumetric efficiency. 3.3 Coefficient of performance The COP of a heat pump is estimated using the equation reported by Ji et al. (2008). The COP of the system is given by following relation: Q COP = cond + W PV 0.38 OP ( ) comp The condenser heating capacity and photovoltaic power out are taken out from the system. Here, W comp is the instantaneous compressor power consumption measured using Wattmeter (W), PV OP is the photovoltaic output (W). The instantaneous compressor power consumption and photovoltaic output were measured directly using a Wattmeter and a multimeter, respectively. 3.4 Photovoltaic thermal evaporator heat gain The capacity of a PV TE is rated in terms of evaporator heat gain ( Q PV-TE ). The QPV-TE is calculated by following equation (Ji et al. 2007): Q = m c ( T -T )- W = m ( h -h ) (7) PV-TE a p 6 5 comp r 1 4 Here, m is the mass flow rate of air (kg/s), c a p is the specific heat of air at constant pressure (J/(kg K)), T 5 and T 6 are the temperature of air at entry and exit of the air cooled condenser ( C) respectively, W comp is the compressor power consumption (W). The specific enthalpy of refrigerant at the entry and exit of PV TE is given by h 4 and h 1 (kj/kg), respectively. 3.5 Solar energy input ratio The solar thermal energy and ambient energy (due to the temperature difference between the PV TE surface and ambient air) are absorbed by the refrigerant through the PV TE. The amount of thermal energy absorbed by the PV TE is always higher than the solar thermal energy falls on the PV TE surface. The influence of solar thermal energy on energy performance of a PV TE is characterized by solar energy input ratio (SEIR). The solar energy input ratio is the ratio between the amount of solar energy absorbed to the total amount of energy absorbed by the refrigerant through (6) ambient source and solar source (Xu et al. 2006). The SEIR is given by the following relation: A I SEIR = (8) Q PV-TE Here, A is the panel area exposed to solar radiation (m 2 ), I is the solar intensity (W/m 2 ). 3.6 Photovoltaic panel efficiency The electricity conversion efficiency of PV TE is rated in terms of photovoltaic efficiency (η pv ). The η pv of a PV TE is the ratio between photovoltaic power output to the total amount of energy incident on the photovoltaic panel surface. The η pv is given by following relation: PVOP η = pv (9) A I Here, PV OP is the photovoltaic output (W), A is the area of photovoltaic panel (m 2 ), I is the solar intensity (W/m 2 ). 4 Uncertainty analysis The uncertainties of the measured and calculated parameters are estimated according to Holman (2007). In this work, refrigerant pressures and temperatures at the typical locations in a heat pump circuit, air temperatures at the inlet and exit of an air cooled condenser, solar intensity, instantaneous compressor power consumption and airflow rate were measured with the instruments as mentioned in the previous sub-section 2.2. The uncertainties in measured and calculated parameters are given by the following equation: 2 2 2 1/2 é R R R ù wr = ( w1) + ( w2) + + ( wn) (10) êë x1 x2 xn úû Here, R is a given function, w r is the total uncertainty, x 1, x 2,..., x n is the independent variables w 1, w 2,..., w n is the uncertainty in the independent variables. The detailed equations for predicting the uncertainty in condenser heating capacity, compressor instantaneous power consumption, photovoltaic power output and coefficient of performance are given by the following equations. The uncertainty in condenser heating capacity is given by δq cond é δm δt δt = + + êë m T T 2 2 2 1/2 a cao cai ù ( ) ( ) ( ) a cao cai úû (11) The uncertainty in compressor power consumption is given by 2 2 1/2 δer δr δw é ù comp = ( ) + ( ) (12) êë Er R úû

Mohanraj et al. / Building Simulation 7 Here, Er is the error in Wattmeter, R is the electrical resistance in electrical cables. The uncertainty in photovoltaic output is given by δpv OP é δpr = + êë PR 2 2 1/2 δclv Clv ù úû ( ) ( ) (13) Here, PR is the power rating specified by the manufacturers, Clv is the error due to climatic variation. The uncertainty in COP is given by δcop é 1/2 2 2 2 δqcond δpv δw OP comp = ê( ) + ( ) + ( Qcond PVOP W ) ú comp ë ù û (14) The uncertainties for Q PV-TE, SEIR, η pv and panel surface temperature are calculated using the following equations: δq PV-TE δseir é δm δt δt = ê + + + m ë W T T 1/2 2 δw 2 2 2 a comp cao cai ( ) ( ) ( ) ( ) é δm 2 a = ê ( ) + m a êë a comp cao cai δw δt δi δt + + + W T I T ù ú û (15) 1/2 2 2 2 2ù comp cao cai ( ) ( ) ( ) ( ) comp cao cai ú û (16) Here, m a is the mass flow rate of air through the condenser, P is the compressor power consumption, T cao is the condenser air temperature at the outlet, T cai is the condenser air temperature at the inlet. I represent the solar intensity measured. δη pv é δpr δclv δi = + + êë PR Clv I ù úû 2 2 2 1/2 ( ) ( ) ( ) (17) Here, PR is the power rating specified by the manufacturers, Clv is the error due to climatic variation. 2 2 2 δt ( ) ( ) ( ) 1/2 panel = é Acc + Er + R ù ê ë ú û (18) Acc is the accuracy of the sensor, Er is the error in the display indicator and R is the resistance loss. The uncertainties in condenser heating capacity, instantaneous compressor power consumption, photovoltaic power output and COP are ±4.2%, ±3.4%, ±4.1% and ±6.8%, respectively. The uncertainties for Q PV-TE, SEIR, η pv and panel surface temperature are ±5.4%, ±8.8%, ±8.11% and ±3.9%, respectively. 5 ANN modeling of heat pumps The basic information on ANN is not discussed in this paper, since it is very much available in the open literature (Mellit and Kalogirou 2008). The ANN modeling of heat pumps involves the experimental data for training the network, which has not required system specifications. The performance of a PV TE is influenced by ambient parameters such as, solar intensity, ambient temperature, ambient relative humidity, ambient wind velocity and dust accumulation over the photovoltaic surface (Rejeb et al. 2015). The surface of the PV TE was cleaned before experimental observations. Hence, in the present analysis, the effect of dust accumulation was ignored. In this work, the solar intensity, ambient temperature, ambient wind velocity and ambient relative humidity were taken as inputs. The energy performance parameters of a heat pump and PV TE were taken as outputs for the Network A and Network B, respectively. The results extracted from the series of steady state experiments (experimental observations after the system attained normal working temperature) conducted in a heat pump for the period of sixty days were used as training data for the network. The one hundred and thirty selected steady state experimental data were considered for discussion in this paper. The training data requirement was optimized by trial and error approach. The MLFFN with back propagation learning algorithm is the most widely used network architecture for the performance prediction of heat pumps (Mohanraj et al. 2012). In this work, two MLFFN models with four neurons in input layer and four neurons in output layer were developed for the performance prediction of the heat pump and the PV TE. The first network was used for simulating the energy performance parameters of the heat pump, such as condenser heating capacity, compressor power consumption, photovoltaic power output and coefficient of performance. The second MLEFFN is used for simulating the energy performance of a PV TE, such as solar energy input ratio, evaporator heat gain, photovoltaic efficiency and photovoltaic panel temperature. The network parameters such as the number of hidden layers, number of training data, momentum factor and learning rate were optimized by trial and error approach. The number of neurons in the hidden layer was optimized by the following equation (Kalogirou and Bojic 2000): 1 Number of hidden neurons = ( Input neurons + Output neurons) 2 + Number of training data (19) Since the hidden neuron requirement obtained from the above equation is in fractional value, it is necessary to find out the exact number of neurons in hidden layers. After enough trials with different configurations by changing the number of neurons in hidden layer (14, 15 and 16), changing the number of input output training data, changing the transfer function (log-sigmoid and tangent sigmoid), changing the variant (LM, SCG and CGP), momentum factor (between 0.7 and 0.9) and learning rate (between 0.7 and 0.9), it is decided that the network consists of one hidden layer with

8 Mohanraj et al. / Building Simulation fifteen neurons, one hundred twenty training data, LM variant, momentum factor of 0.9 and learning rate of 0.8. The network configurations were optimized to 4 15 4 configuration. The input output data were normalized in the range of 0 1. The log-sigmoid transfer function was used for hidden and output layers. The configurations of MLFFN with their names of inputs and outputs for predicting the energy performance of heat pump and PV TE are depicted in Figs. 4 and 5, respectively. The steps involved in ANN simulation is illustrated in Fig. 6. ANN used for energy performance prediction was made in MATLAB (version 7.0) environment using neural network toolbox. The back propagation learning algorithm optimizes the weight connection by allowing the error to spread from output layers towards the lower layers (hidden layer and input layer). The output of the network was compared with desired output at each presentation and errors were computed. These errors were then back propagated to the ANN for adjusting the weight such that the errors decreased with each iteration and ANN model approximated the desired output. The reliability of the results predicted using MLFFN Fig. 6 Flow chart of ANN simulation procedure Fig. 4 MLFFN configuration for energy performance prediction of a heat pump (Network A) modeling of a PV TE was evaluated by the fraction of absolute variance (R 2 ), root mean square error (RMS) and coefficient of variance (COV) values, which are given by following equations: R 2 = 1- n å m= 1 ( y -t ) pre, m n å m= 1 ( t ) mea, m 2 mea, m 2 (20) RMS = n å m= 1 mea,avg ( y -t ) pre, m n mea, m 2 (21) RMS COV = 100% (22) t Fig. 5 MLFFN configuration for energy performance prediction of a PV TE (Network B) Here, n is the number of data pattern in the independent data set, y pre,m indicates the predicted by ANN, t mea,m is the measured value of one dated point, t mea,avg is the mean value of all measured data points.

Mohanraj et al. / Building Simulation 9 6 Results and discussions The energy performance results obtained in the heat pump using circular and triangular tube PV TE configurations are compared in this section. 6.1 Experimental observations The refrigerant temperature variations at salient points in a heat pump circuit with respect to solar intensity are depicted in Fig. 7. It is observed that refrigerant temperatures at the compressor suction and compressor discharge are varied from 9.3 to 14.7 C with average temperature of 11.9 C and from 48.3 to 80.2 C with an average temperature of 64.5 C, respectively. Similarly, the refrigerant temperature is varied between 27.3 and 38.4 C with an average value of 32.8 C at the inlet of an expansion valve. The temperature of refrigerant at expansion valve outlet is varied between 5.3 and 11.7 C with an average temperature of 9.2 C. The compressor discharge temperature is the major parameter influencing the compressor life. In this work, the maximum compressor discharge temperature observed is less than 80.2 C, which ensures a better compressor life. The variations in air temperatures at entry and exit of the condenser are plotted in Fig. 8. The ambient temperature varies between 27.2 and 35.3 C with an average temperature of 31.5 C. The air temperature at the condenser exit is varied between 33.1 and 43.8 C with an average temperature of 38.1 C. As expected, all the temperatures measured at typical locations in the refrigerant flow path and airflow path get increased with increase in solar intensity. Similarly, the refrigerant temperature variations at salient points in a heat pump circuit using triangular tube configurations are depicted in Fig. 9. It is observed that refrigerant temperature variations at the compressor suction and discharge are 9.6 15.3 C and 48.8 81.1 C, respectively. The respective average temperatures are 12.5 and 65.1 C. Similarly, the refrigerant temperature lies between 27.7 and 38.9 C with an average value of 33.1 C at the inlet of an expansion valve. The temperature of refrigerant at expansion valve outlet ranges from 5.4 to 12 C with an average temperature of 9.7 C. The variations in air temperatures at entry and exit of the condenser are plotted in Fig. 10. Ambient temperature varies between 27.7 and 34.7 C with an average temperature of 32.7 C. The air temperature at the condenser exit is varied between 33.4 and 45.3 C with an average temperature of 39.4 C. Air temperature at condenser outlet is about 2 C higher when compared to the system operating Fig. 7 Variation of refrigerant temperature in a heat pump circuit with reference to solar intensity for circular tubes Fig. 9 Variation of refrigerant temperature in a heat pump circuit with reference to solar intensity for triangular tubes Fig. 8 Variation of air temperatures at inlet and outlet of a condenser with reference to solar intensity for circular Fig. 10 Variation of air temperatures at inlet and outlet of a condenser with reference to solar intensity for triangular tubes

10 Mohanraj et al. / Building Simulation with circular tube configuration. As expected, all the temperatures at all the typical locations in the heat pump circuit and in the airflow path are increased with the increase in solar intensity, which is similar to the circular tube PV TE. 6.2 Ambient conditions considered for performance comparison The ambient conditions considered for the energy performance comparison of heat pump were measured on a clear sunny day with maximum solar intensity of 950 W/m 2 and minimum fluctuations of solar intensity (16.04.2014: Wednesday) under the meteorological conditions of Coimbatore city in India. The variation of solar intensity and ambient temperature observed from 8:00 to 17:00 are shown in Fig. 11. It is observed that both solar intensity and ambient temperature increase in the forenoon and decreases in the afternoon. The solar intensity increases from about 215 W/m 2 to a maximum solar intensity of about 955 W/m 2. Similarly, the ambient temperature varies from about 27 C to about 36 C. A maximum solar intensity and ambient temperature are observed during 13:00. Similarly in Fig. 12, the variations of relative humidity and ambient wind velocity are depicted. It is observed that the ambient relative humidity varies between 55% and 71%. The ambient wind velocity varies between 2 and 4 m/s. 6.3 Network performance The statistical performance parameters such as, fraction of absolute variance (R 2 ), coefficient of variance (COV) and root mean square errors (RMS) of Network A and Network B with fifteen neurons in the hidden layer are listed in Table 3 and Table 4, respectively. From Table 3 and Table 4, it is confirmed that experimental values are closer to the ANN predicted results with maximum fraction of absolute variance for all the energy performance parameters of the heat pump and PV TE using circular and triangular tube configurations. The RMS errors and coefficient of variance values are found to be low for all the ANN predicted values. 6.4 Energy performance comparison The energy performance of heat pumps using circular and triangular tube configurations in PV TE is compared in this section. 6.4.1 Condenser heating capacity Fig. 11 Variation of solar intensity and ambient temperature considered for performance comparison Fig. 12 Variation of ambient relative humidity and ambient wind velocity considered for performance comparison The condenser heating capacity of a heat pump using circular and triangular tube configurations are depicted in Fig. 13. The condenser heating capacity is influenced by the major ambient parameters such as, solar intensity and ambient temperature (Mohanraj et al. 2008, 2009b,c). It is observed that the heat pump using triangular tube PV TE configuration enhances the condenser heating capacity by 3% 7% when compared to the heat pump system using circular tube PV TE configuration. The condenser heating capacity of the heat pump using circular tube varies between 2440 and 3900 W with an average value of about 3200 W. Whereas, the condenser heating capacity of the heat pump using triangular tube configuration varies from 2600 to 4000 W with an average value of about 3350 W. The heat absorption in PV TE using triangular tube configuration is higher when compared to the circular tube configuration due to the increase in surface contact with the absorber plate. The triangular tube configuration has about 2 3 C higher superheating effect compared to the circular tube configuration. An increase in evaporator heat gain by 4% 9% and increase in compressor consumption by 1% 2% in the case of triangular tube PV TE configuration have increased the refrigerant temperature at the compressor discharge enhances the condenser heating capacity by 3% 7% when compared to the circular tube configuration.

Mohanraj et al. / Building Simulation 11 Table 3 Performance of Network A Circular tube Triangular tube Parameter R 2 COV RMS R 2 COV RMS W comp (W) 0.9993 0.0184 0.1897 0.9989 0.2339 2.910 Q cond (W) 0.9998 0.1215 3.8677 0.9965 0.2187 4.230 PV OP (W) 0.9997 0.1644 0.2245 0.9919 0.2684 0.1911 COP 0.9998 0.0600 0.0020 0.9938 0.1443 0.0249 Table 4 Performance of Network B Circular tube Triangular tube Parameter R 2 COV RMS R 2 COV RMS SEIR 0.9999 0.0477 0.4743 0.9911 0.3264 0.2803 Q e (W) 0.9999 0.0932 2.036 0.9927 0.4289 3.2011 pv 0.9999 0.1228 0.0139 0.9919 0.6092 0.280 Tpl ( C) 0.9999 0.0545 0.0200 0.9921 0.3328 0.2148 Fig. 13 Variation of system condenser heating capacity using circular and triangular tube configurations 6.4.2 Compressor power consumption The compressor power consumption of a heat pump using triangular and circular tube configurations is compared in Fig. 14. The power consumption for both the configurations increases with increase in solar intensity and ambient temperature due to the increase in refrigerant super heat and evaporator pressure (Mohanraj et al. 2008, 2009b, c). From Fig. 14, it is observed that the compressor power consumption of a heat pump using circular tube varies between 935 and 1090 W with an average power consumption of about 1025 W. The compressor power consumption of a heat pump using triangular tube configuration varies between 945 and 1100 W with average value of about 1035 W. The compressor power consumption is 1% 2% higher in the heat pump working with triangular tube PV TE configuration when compared to the heat pump system using circular tube PV TE configuration. About 5% higher pressure drop in the triangular tube evaporator configuration and increase in refrigerant quantity by 50 grams have increased the compressor power consumption by 1% 2%. Fig. 14 Variation of system compressor power consumption using circular and triangular tube configurations 6.4.3 Photovoltaic output The performance of a photovoltaic panel is influenced by five ambient parameters such as solar intensity, ambient temperature, ambient relative humidity, ambient wind velocity and dust accumulations (Mekhilef et al. 2012). The panel surface was cleaned before the experimentation. Hence, the influence of dust accumulation was ignored in this work. The comparison between the panel output of PV TE using circular tube configuration and triangular tube configuration are compared in Fig. 15. The photovoltaic power output is varied from 50 to 175 W with an average output of 130 W for circular tube PV TE and the output is varied between 55 and 205 W for triangular tube configuration. It is observed that the photovoltaic power output of triangular tube PV TE is 8% 14% higher than that of circular tube PV T evaporator. Higher panel output is observed with triangular tube configuration due to absorption of more heat from the photovoltaic panel by the refrigerant. It maintains the temperature of the panel around 25 C. High power output is observed

12 Mohanraj et al. / Building Simulation Fig. 15 Variation of photovoltaic power output using circular and triangular tube configurations during peak sunshine hours due to the availability of more solar energy. The photovoltaic power output is relatively low during early morning and late afternoon due to poor tracking of solar illuminations. The average electricity generation per day is around 1.3 and 1.5 kwh for the circular tube PV TE and triangular tube PV TE, respectively. 6.4.4 Coefficient of performance The coefficient of performance is the main factor considered for rating the energy performance of heat pumps. The COP of a heat pump using PV TE evaporators depends on the photovoltaic power output, compressor power consumption, evaporator heat gain and condenser heating capacity. The four ambient parameters such as, solar intensity, ambient temperature, ambient wind velocity and ambient relative humidity are directly influencing the energy performance of heat pumps. In Fig. 16, the COP of a heat pump using circular and triangular tubes is compared. The COP of a heat pump using circular tube configuration is varied between 2.80 and 3.90. Whereas, the COP of heat pump using triangular tube configuration is varied from 2.90 to 4.10. It is observed that the COP of the heat pump using a triangular tube evaporator configuration is 3% 5% higher when compared to the heat pump using a circular tube evaporator. The compressor power consumption of a heat pump using triangular tube configuration is about 2% higher than that of circular tube configuration. However, the heat gain in the PV TE is 4% 10% higher in the case of triangular tube configuration when compared to the circular tube configuration. The photovoltaic power output in the case of triangular tube configuration is also 9% 15% higher when compared to the circular tube configuration. The high evaporator heat gain and photovoltaic power output have influenced the COP enhancement by 3% 5%. 6.4.5 Photovoltaic evaporator heat gain The heat gain of PV TE using circular and triangular tube configurations in a heat pump is compared in Fig. 17. The PV TE heat gain depends on solar intensity, ambient temperature and ambient wind velocity. The heat gain in PV TE is varied from 1500 to 2800 W with an average evaporator heat gain 2160 W for circular tube configuration. Whereas, the evaporator heat gain in the case of triangular tube configuration is varied from 1650 to 2900 W with an average value of 2350 W. More amount of heat is absorbed during the peak sunshine hours due to the absorption of solar thermal energy and ambient energy due to the temperature difference between absorber plate and ambient conditions. Amount of heat gain in the triangular tube PV TE configuration is 4% 9% higher when compared to the circular tube PV TE configuration. The pressure drop in the case of triangular tube PV TE is around 5% higher due to the presence of sharp edges when compared to the circular tube PV TE, which influences more heat absorption. 6.4.6 Solar energy input ratio The SEIR of PV TE using circular and triangular tube configurations in a heat pump is compared in Fig. 18. The SEIR is an important parameter considered for evaluating the energy performance of the evaporators used in solarassisted heat pumps (Xu et al. 2006; Mohanraj et al. 2009c). Fig. 16 Variation of coefficient of performance using circular and triangular tube configurations Fig. 17 Variation of evaporator heat gain using circular and triangular tube configurations

Mohanraj et al. / Building Simulation 13 Fig. 18 Variation of solar energy input ratio using circular and triangular tube configurations From Fig. 18, it is confirmed that the SEIR varies between 0.29 and 0.7 for the circular tube configuration PV TE. The maximum solar energy absorbed by the PV TE is around 70% of the total heat load during peak sunshine hours for the circular tube PV TE configuration. The remaining 30% of the energy was absorbed from the ambient sources due to the temperature difference between the atmospheric air and absorber plate of PV TE. Whereas, the SEIR varies from 0.25 to 0.65 in the case of triangular tube PV TE configuration. The maximum solar energy absorbed is around 65% of total energy absorbed by the PV TE. The remaining about 35% of the energy has absorbed from the ambient sources due to the temperature difference between the atmospheric air and absorber plate. The SEIR for the circular tube configuration PV TE is 4% 10% higher than that of triangular tube configuration. The remaining heat is absorbed from the ambient sources due to the temperature difference between the absorber surface and the ambient conditions. 6.4.7 Photovoltaic efficiency The photovoltaic panel electrical efficiency of the PV TE using circular and triangular tubes is compared in Fig. 19. It is observed that panel efficiency is varied between 9.7% and 13.1% for circular tube PV TE configuration, whereas, the panel efficiency is varied between 11.1% and 13.7% for triangular tube configuration under the influence of solar intensity variations from 215 to 955 W/m 2. The electrical conversion efficiency of a PV TE using triangular tube configuration has about 4% 13% higher when compared to the circular tube PV TE configuration. The triangular tube configuration provides more contact area with the absorber plate when compared to the circular tube configuration as depicted in Fig. 3(b). The triangular tube PV TE configuration recovers 4% 9% higher evaporator heat gain and enhances the panel electrical efficiency by 4% 13% when compared to circular tube configuration. The panel efficiency gets reduced with the increase in solar intensity, which is similar Fig. 19 Variation of photovoltaic panel efficiency using circular and triangular tube configurations to the earlier reported work (Shan et al. 2014). The drop in panel efficiency is due to the losses in the photovoltaic panel at elevated ambient temperature. 6.4.8 Photovoltaic panel temperature The temperature of the photovoltaic panel should be maintained around 25 C to achieve high photovoltaic efficiency. However, if the panel surface is maintained below the dew point temperature of the ambient air, the moisture present in atmospheric air starts condensing over the panel surface. The condensed moisture may infiltrate into the cell enclosure of the panel. Hence, it is suggested to maintain the temperature of the panel above the dew point temperature of the air. Generally, the temperature of photovoltaic panel should be maintained around 25 C to achieve better energy performance. The photovoltaic panel temperature of a PV TE using circular and triangular tube configurations are compared in Fig. 20. The temperature of the panel using circular tube PV TE is varied between 21 and 29 C, whereas, the temperature of panel using triangular tube is varied between 20 and 27 C. The temperature of the panel using triangular tube PV TE is about 2 C lower when compared to the panel using circular tube PV TE, which enhances the photovoltaic efficiency by 4% 13%. Fig. 20 Variation of photovoltaic panel temperature using circular and triangular tube configurations

14 Mohanraj et al. / Building Simulation 6.5 Parametric analysis The energy performance of a heat pump under the influence ambient parameters (such as, solar intensity, ambient temperature, ambient relative humidity and ambient wind velocity) and system operating parameters (such as, condenser and evaporator operating pressures, degree of sub-cooling and superheating) are discussed in this section. 6.5.1 Influence of ambient parameters The solar radiation fall on PV TE is the major influencing parameter affecting the thermodynamic performance of a heat pump. The condenser heat output and photovoltaic output get increased from 2600 to 4000 W and from 55 to 205 W with the increase in solar radiation from 215 to 955 W/m 2, respectively. The compressor power consumption is increased from 945 to 1100 W with the increase in solar intensity and increase in ambient temperature from 27 to 36 C. The photovoltaic efficiency gets decreased with the increase in solar radiation, which is similar to the previous work (Shan et al. 2014). Even though the photovoltaic power output has reached the maximum value of 205 W during the peak sunshine hours around 900 W/m 2, the losses in the photovoltaic panels have reduced the photovoltaic electrical conversion efficiency to 9.7% in the case of circular tube configuration and 11.1% in the case of triangular tube configuration. An increase in condenser heating capacity and photovoltaic power output will enhance the heat pump COP from 2.8 to 3.9 for the circular tube configuration with increase in solar intensity and ambient temperature from 215 to 955 W/m 2 and 27 to 36 C, respectively. Whereas, the COP value varied between 2.9 and 4.10 for the heat pump system working with triangular tube configuration with variations in ambient conditions. Additionally, the ambient temperature has influenced the thermal load on PV TE due to its temperature difference between the ambient air and the panel surface. The ambient wind velocity is providing the effective convective cooling of photovoltaic panels, which maximize the photovoltaic output (Mekhilef et al. 2012; Rejeb et al. 2015). But, the thermal load acting on PV TE slightly reduces with the increase in ambient wind velocity, which results in loss of heat pump COP. The high ambient moisture content (above 70%) in the ambient air will have a chance to condense over the photovoltaic cells. The condensed moisture may infiltrate into the cell enclosure and reduce the photovoltaic performance. Hence, it is essential to seal the panel to avoid the moisture infiltration into the cell enclosure (Mekhilef et al. 2012; Rejeb et al. 2015). The instantaneous energy performance of heat pump systems is not affected by humidity. However, the long term energy performance of heat pumps is highly affected due to the moisture infiltration into the cell enclosure. The surface of photovoltaic panel is cleaned with a soft cloth and kept the surface clean and dry before the experimental observations. However, the influence of moisture and dust accumulation on the PV TE of a heat pump needs further research investigation. 6.5.2 Influence of system operating parameters During peak sunshine hours (above 750 W/m 2 ), the condenser and evaporator pressures are getting increased from 8.5 to 12 bar and from 0.9 to 1.3 bar, respectively. An increase in operating pressures results in higher compressor pressure ratio, which will reduce the compressor volumetric efficiency. Hence, it is essential to control the compressor discharge pressure within 12.5 bar and refrigerant quantity around 650 grams by integrating the pressure controls in the compressor discharge and in the compressor suction, respectively. The superheating effect in the evaporator is also influencing the compressor pressure ratio, which results in loss of compressor volumetric efficiency. Hence, the evaporator superheating effect is controlled within 5 C by a thermostat installed at the compressor suction. Further, the operating pressures and superheating effect in the evaporator are also controlled by integrating a liquid receiver after the condenser. 7 Conclusions The energy performance of a heat pump was evaluated using circular and triangular tube PV TE configurations. The specific conclusions arrived in this research work are as follows: The condenser heating capacity of the heat pump using triangular tube configuration is 3% 7% higher when compared to the heat pump using circular tube configuration due to more evaporator heat gain. The compressor power consumption using triangular tube PV TE configuration is 1% 2% higher when compared to the heat pump system using circular tube PV TE due to the increase in refrigerant mass charge by 50 grams. The photovoltaic power output of a triangular tube PV TE is 9% 15% higher when compared to the circular tube PV TE due to the absorption more heat. The COP of a heat pump using triangular tube PV TE is 3% 5% higher when compared to the heat pump using circular tube PV TE due to 4% 9% higher evaporator heat gain and 1% 2% higher compressor power consumption in the case of triangular tube configuration. The evaporator heat gain in triangular tube configuration is 4% 9% higher than that of circular tube PV TE configuration due to the absorption of more thermal energy. The solar energy input ratio for the circular tube PV TE configuration is 4% 10% higher than that of triangular tube PV TE configuration.