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Annual simulation, energy and economic analysis of hybrid heat pump systems for residential buildings
Accepted Manuscript Title: Annual simulation, energy and economic analysis of hybrid heat pump systems for residential buildings Author: G. Bagarella, R. Lazzarin, M. Noro PII: DOI: Reference:
S1359-4311(16)30039-4 http://dx.doi.org/doi: 10.1016/j.applthermaleng.2016.01.089 ATE 7649
To appear in:
Applied Thermal Engineering
Received date: Accepted date:
11-11-2015 19-1-2016
Please cite this article as: G. Bagarella, R. Lazzarin, M. Noro, Annual simulation, energy and economic analysis of hybrid heat pump systems for residential buildings, Applied Thermal Engineering (2016), http://dx.doi.org/doi: 10.1016/j.applthermaleng.2016.01.089. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
ANNUAL SIMULATION, ENERGY AND ECONOMIC ANALYSIS OF HYBRID HEAT PUMP SYSTEMS FOR RESIDENTIAL BUILDINGS (*)
G. BAGARELLA, (**) R. LAZZARIN, (***) M. NORO
(*) (**) (***)
Department of Management and Engineering, University of Padua President of IIR Section E, Air conditioning, heat pumps, energy recovery Str.lla San Nicola 3, Vicenza, 36100, Italy (*) [email protected] - (**) [email protected] - (***) [email protected] (**)
A model of a hybrid system (heat pump and condensing boiler) was implemented Simulations were run with both a cold-humid and a mild-dry climate With a large size HP there are no benefits by setting a Tcut-off lower than BT With a low size HP, a bivalent parallel system can lead to energy savings
ABSTRACT Air source heat pumps have a well known unfavorable characteristic: in the coldest period of the year, when the building heating load is at maximum, the heat pump capacity is reduced. A possible solution is to size the heat pump to cover only a fraction of the peak load using a second heat generator in “hybrid heat pump system”. In this study a dynamic model of a hybrid system was built and several seasonal simulations were carried out (considering two different climates) to study how the choice of the cut off temperature can influence the annual efficiency of the system and to understand if a bivalent parallel plant can lead to energy savings compared to a bivalent alternative system. With both the considered climates, when a large size heat pump is selected (low bivalent temperature), there might be no benefits by setting a cut off temperature lower than the bivalent temperature, and the differences between the energy performances of the bivalent alternative system and those of the bivalent parallel system are negligible. On the other hand, when a low size heat pump is preferred (high bivalent temperature), the bivalent parallel system can lead to appreciable energy savings compared to a bivalent alternative plant. The main economic advantages of a hybrid system, with respect to the monovalent heat-pump plant, comes from the lower annual electric and gas energy needs.
KEYWORDS Hybrid system, heat pump, boiler, cut off temperature, bivalent temperature
NOMENCLATURE 1 Page 1 of 19
BAS = Bivalent Alternative System BPS = Bivalent Parallel System BT = Bivalent Temperature (°C) COP = Coefficient of Performance Eel_HP= Annual electric energy absorbed by the heat pump (kWh) Eth_boiler= Annual thermal energy produced by the boiler (kWh) Eth_building= Building annual thermal energy demand (kWh) EEV = Electronic Expansion Valve HP = Heat Pump Kv = Valve Flow Coefficient (m3 h-1) NPV = Net Present Value (€) Pth = Thermal Capacity (kW) RH = Relative Humidity SCOP = Seasonal Coefficient of Performance Tcut-off = Cut-off Temperature (°C) Tc = Condensation Temperature (°C) Te = Evaporation Temperature (°C) Text=Temperature of the external air (°C) Tset =Set-point Temperature (°C) Tw out =Temperature of the water at the tank outlet (°C) b = boiler efficiency s = system efficiency 1. INTRODUCTION Heat pumps are a promising approach to reduce the production of greenhouse gases in the building sector, because they use freely available ambient heat. In recent years the use of heat pumps for space heating has spread widely. This is mainly due to the evolution of the technical features of devices, of their performances and of the electricity generation system efficiency. Furthermore, in Europe the Directives 2010/31/EU on energy performance of buildings and 2012/27/EU on energy efficiency introduced the compulsoriness of producing increasing part of a building thermal energy needs by renewables; energy 2 Page 2 of 19
from the heat source (air, water or ground) used by heat pumps is considered renewable by the Directives and this is giving more and more chances to the heat pumps market. Outside air is the most widespread heat source for heat pumps. Outside air has a well known unfavorable characteristic: in the coldest period of the year, when the building heating load is at maximum, the heat pump capacity is reduced (Klein et al., 2014). Moreover, being the peak load required for a small fraction of the heating period (Lazzarin, 2012), if the heat pump unit is sized to cover the peak load when the air temperature is at the minimum value, the annual average on-off frequency of the unit might be really high, lowering the seasonal coefficient of performance (SCOP) due to cycling losses (Bagarella et al., 2013). A possible solution is to size the heat pump to cover only a fraction of the peak load. In this case, a second heat generator (usually electric heaters or a boiler) is required and the system is called “hybrid heat pump system”. When speaking about hybrid heat pump systems some definitions should be reminded. The bivalent temperature BT (Ente Nazionale Italiano di Normazione, 2012) may be defined as the temperature at which the thermal capacity of the heat pump equals the building thermal load. Thus, for a given building, the higher BT the smaller the heat pump size. The cut-off temperature (Tcut-off) is the temperature of the external air below which the heat pump is switched-off by the controller. If the temperature of the external air (Text) is higher than BT, the thermal power is produced only by the heat pump unit. If Tcut-off
cut-off temperature: should the choice of Tcut-off take into account the heat pump size and the climate? Moreover, to the authors’ best knowledge, there is still lack of information regarding the differences between a bivalent parallel system and a bivalent alternative plant, in terms of energy performance, for different climates. In this study a dynamic model of a hybrid system was built and several seasonal simulations were carried out to study how the choice of Tcut-off can influence the annual efficiency of the system and to understand if a bivalent parallel plant can lead to energy savings compared to a bivalent alternative plant. In order to generalize some of the main conclusions, simulations were run considering two different climates. 2. METHODOLOGY The model of the whole hybrid heat pump system was developed in TRNSYS 17 environment (Klein et al., 2010). As depicted in Figure 1, the air-to-water heat pump (HP) and the boiler operate in parallel mode, which is the most widespread solution in existing hybrid system applications (Klein et al., 2014). The system controller establishes whether to activate the heat pump only, the boiler only or both the heat generators, depending on the temperature of the water at the tank outlet (Tw out) and on the temperature of the external air. If Tw out is lower than the set-point (Tset) minus a dead-band (1.5 °C), at least one heat generator is activated. In particular, if Text is higher than the bivalent temperature, only the heat pump is activated. If Text is lower than BT but higher than the cut-off temperature, both the heat generators operate in parallel mode. Finally, if Text is lower than Tcut-off, only the boiler is activated. Different heat pump sizes were tested in the simulations, as well as different values of Tcut-off. On the contrary, the volume of the thermal storage was kept constant (0.3 m3). This volume allowed all the tested heat pump units to have a maximum on-off frequency lower than 5 switches per hour. The thermal efficiency of the condensing boiler was considered to be a function of the water temperature at the boiler inlet. In particular, the thermal efficiency profile suggested by Lazzarin (2012) was used. A model of a 10.8 kW (at Te= -10 °C, Tc = 55 °C, RH = 50 %) on-off air-to-water heat pump was firstly developed. Main information about each component are summarized in Table 2. The compressor was modelled basing on refrigerant mass flow rate and electric power profiles suggested by the compressor manufacturer (Emerson-Copeland, 2015). The plate condenser was simulated thanks to SSP-G7 calculation software (SWEP, 2015), developed by the condenser manufacturer. This software allows to simulate the thermal power exchanged by each plate condenser model varying the inlet conditions of both the fluids (water – R134a). The finned coil evaporator was modelled with the software EVAP-COND developed by NIST (National Institute of Standards and Technology, 2003). This software, also used in the past by Domansky et al. (2005), allows to design a finned coil evaporator and to simulate the thermal power 4 Page 4 of 19
exchanged in different test conditions. Moreover, the software estimates both the air temperature and air relative humidity in each point of the heat exchanger surface. Therefore, it was considered a useful tool in order to roughly simulate the evaporator frosting process, as the thermal resistance of both fins and tubes can be set, at each time-step, basing on the frost properties (thickness, density and conductivity). Being aware of the complexity of the frost formation problem, the same unit and the same test conditions considered by Guo et al. (2008) were firstly simulated, in order to validate our simplified evaporator model. In Figure 2, results experimentally obtained by Guo et al. (2008) are compared with those estimated by the simplified model. Three different tests were carried out, keeping the temperature of the external air constant (0 °C) with three values of the relative humidity (RH). Figure 2 shows the good agreement between the evaporator wall temperature and the frost thickness profiles estimated by the model and the values obtained by the survey of the real frost formation process, at least for the purposes of this study. As the approximation of the model becomes unacceptable when the average frost thickness is higher than 0.6 mm, the annual simulations have been carried out defrosting (by electric heaters) when the average frost thickness is close to that value. The defrosting efficiency (i.e. the ratio between the actual amount of energy required to melt the accumulated frost to the total amount of heat supply to the coil during the entire defrosting operation) was considered to be 60 %, as suggested by Dong et al. (2012). The models of the heat pump components were linked together as in Madani et al. (2011). To speed up the annual simulation of the unit (a one minute time step was required to simulate the evaporator frosting process with a sufficient approximation), superheating and subcooling were considered constant as in Cecchinato et al. (2010). REFPROP 9.0 was used to estimate the properties of the refrigerant in each operating condition. Cycling losses, defined as those inefficiencies which characterize the start-up period of a unit, were also considered basing on several measures reported in Bagarella et al. (2012). The same renovated building described by Klein et al. (2014) was considered, which consists in a typical European one family house, built in 1970s, with two floors, 132 m2 net living space and with typical energetic retrofitting interventions (14 cm insulation of outer walls, new coated double glazing windows and 18 cm non-walkable attic floor insulation). These interventions are aimed at reducing the energy needs of an old building without modifying the heating distribution system. More information about building properties and internal set points are described in that paper. The heating distribution system is made up of radiators and the heating curve proposed by Arteconi et al. (2013) was used (45 °C ≤Tset ≤ 55 °C as in Table 3). Among the possible European climates, a cold-humid climate (Copenhagen, Denmark) and a mild-dry climate (Carpentras, France) were here chosen (Figure 3). For the sake of brevity detailed results are here 5 Page 5 of 19
reported only for the cold-humid climate (where the issue of defrosting is especially important) while main results obtained with the mild-dry climate are summarized in section 3.4. 3.
RESULTS
3.1 Heat pump unit performances Some simulations were firstly run to characterize the performance of the heat pump unit. Figure 4 shows the COP profiles as a function of both the temperature of the external air and the relative humidity. To a better understanding of the heat pump performance, Tset was regarded as constant (60 °C) in Figure 4, while it was assumed a function of Tset in the annual simulations. COP profiles are coherent with those experimentally found by Di Perna et al. (2015) with a similar unit. When the frosting process does not occur (Text> 6 °C), the higher the RH the higher the COP, because the latent heat coming from air dehumidification increases the thermal capacity of the heat pump. Moreover, the higher the RH, the lower the maximum Text at which the frosting process occurs. For instance, if RH is 75% frosting appears when Text is lower than 6 °C, while it occurs at 4 °C if RH is 85%. In fact, the lower the RH the higher the contribution of the sensible heat over the total heat transferred by the air, thus the lower are both the temperature of the air at the evaporator outlet and the average temperature of the heat exchanger. When the frosting process occurs, the higher the RH the lower the COP, because energy is required for defrosting. Finally, simulations show that when RH is lower than 50%-55% the frosting process can be considered negligible. This happens either because the temperature of the evaporator surface is higher than the dew point or because the amount of vapour condensed is so little that the frost growth rate is negligible. Figure 4 also depicts both the heat pump thermal capacity and the building thermal load profiles. Thus, with this particular heat pump size, the bivalent temperature is close to -13 °C. 3.2 Annual simulations: energy analysis To better understand how the cut-off temperature affects the energy performance of the system, several simulation were run varying both Tcut-off and the bivalent temperature (thus the heat pump size). Figure 5 shows, for different Tcut-off, the annual primary energies required by both heat pump unit and boiler when BT is 0 °C (heat pump capacity 3.9 kW when Text= -13 °C and RH = 50 %). If Tcut-off is equal to BT (0 °C) the system operates as a bivalent alternative system (BAS).On the other hand, if Tcut-off is lower than BT, the system operates as a bivalent parallel system (BPS) and both the heat generators are activated when Tcut-off
the other hand, the lower Tcut-off, the lower the SCOP but the lower the thermal energy produced by the boiler. The minimum total annual primary energy consumption is obtained when Tcut-off is around 0 °C, thus when it is close to BT. Considering this cold-humid climate, when BT is close to 0 °C, the differences between the energy performance of a parallel or alternative system are negligible. The annual simulations were run again considering a smaller heat pump size (BT = 6 °C). Figure 7 shows that, when a high BT is chosen, the situation is different and there is an appreciable energy saving (up to 5 %) by adopting a bivalent parallel system (Tcut-off< 6 °C) instead of a bivalent alternative system (Tcut-off = 6 °C). The reason is that the bivalent parallel system can drastically decrease the annual thermal energy produced by the boiler, as in Figure 7. The minimum total annual primary energy is obtained when Tcut-off = -1 °C, while the best value was 0 °C when the considered BT was 0 °C. Thus, the best Tcut-off shifts toward lower values when BT increases. The reason is explained by Figure 6, which shows that the SCOP of the heat pump unit with BT = 6 °C is higher than the SCOP of the unit with BT= 0 °C. This is mainly the consequence of the lower number of on-off cycles (thus lower cycling losses) obtainable by lowering the heat pump size. In Figure 7, starting from high Tcut-off values, one could expect the total primary energy to rapidly increase once reached its minimum value (Tcut-off = -1 °C). This does not happen because the annual thermal energy required by the building when Text is low, e.g. lower than – 4 °C, is a small fraction (lower than 10 %) of the total annual thermal energy required, as depicted in Figure 3. Thus, considering this cold-humid climate, when BT is close to 6 °C, the differences between the energy performances of the bivalent parallel system with the best Tcut-off and the same system with Tcut-off set to really low values (e.g. -12 °C) are very small. Other simulations were carried out varying BT from -6 °C to 10 °C. The selection of these temperatures is finalized to study the behavior of different size heat pumps, recalling that a high BT means a low capacity heat pump and vice versa. In Figure 8, for each BT, only the primary energy consumption of the best solution (Tcut-off which minimizes the total primary energy consumption) is reported. Moreover, also the monovalent systems (heat pump only and boiler only) were considered. In this cold-humid climate, the solutions with higher system efficiency (ratio between thermal energy produced and primary energy consumed) are the bivalent systems with BT = -6 °C and BT = 0 °C. 3.3 Annual Simulation: economic analysis An economic analysis was carried out to compare the Net Present Values (NPV) of the alternatives. As in Table 4, the nominal thermal power of the boiler (in each alternative) was heightened by 8.5 kW to ensure 7 Page 7 of 19
a sufficient power also for the hot sanitary water production (this value was estimated by the aid of a dedicated tool developed by AiCARR (Italian Association for Air Conditioning, Heating and Refrigeration, 2015), considering 4 people, 50 l consumption per person and 140 l hot water storage tank). Heat pumps and boilers prices were set basing on a manufacturer catalogues and applying a 40% discount. Electric energy and natural gas prices were initially considered to be respectively 20 c€ kWh-1 and 77 c€ Sm-3 (Italian Authority for electricity, gas and water system, 2015). An annual incentive was considered for each heat pump solution as suggested by Italian legislation: the annual renewable energy used by each alternative was estimated (Eth_building – (Eth_boiler + Eel_HP)) and the annual incentive was calculated applying a 0.055 € kWh-1 enhancement factor (Italian Authority for electricity, gas and water system, 2015). The maintenance costs were estimated as in Busato et al. (2012). Finally, a 3% annual interest rate was used. Figure 9 summarizes the NPV (15 years) of each alternative. The best solutions, from an economic perspective, are the bivalent parallel systems with BT = -6 °C (5.3 kW heat pump size) and BT = 0 °C (3.9 kW heat pump size). The main contribution, with respect to the monovalent heat pump solution, comes from the reduction of both electricity and gas consumption. Moreover, another contribution comes from the lower initial investment, which is the consequence of the lower size of the heat pump. A sensitivity analysis was also carried out varying the prices of both electric energy and natural gas (±30%). Figure 10 and Figure 11 show that the best alternative, from an economic perspective, is strongly influenced by these prices. For instance, the monovalent boiler solution might be the best alternative with low prices of the natural gas, but becomes the worst solution when the price is higher than 85c€ Sm-3. Anyway, the bivalent parallel solution with BT= 0 °C is less sensitive to prices variations and it is always close to the best alternative. 3.4 Summary of energy and economic analysis with another climate As explained in section 2, simulations were run considering a mild-dry climate (Carpentras) as well. For the sake of brevity detailed results (including figures and tables) of simulations with Carpentras’ climate are not reported here. The aim of this chapter is to summarize the main results of these simulations and to explain how and why the different climate affects the system performances and choice of both the best BT and best Tcut-off. The most obvious consequence of the different climate (Figure 3) is that SCOPs of heat pump units are much higher in Carpentras. This is due to both the higher average Text during the winter period, and the lower energy required for the evaporator defrost (up to 9.8 % of the total electric energy in Copenhagen and up to 5.8 % in Carpentras) in the mild-dry climate. For instance, the SCOPs of the heat pump with BT = -6 °C and Tcut-off= -6 °C are 2.39 and 2.82 respectively in Copenhagen and Carpentras. Therefore, the advantages (in terms of primary energy consumption) of hybrid systems with low BT (BT ≤ 0 °C) with 8 Page 8 of 19
respect to the monovalent boiler system are much higher in Carpentras (up to 29.6 %) than in Copenhagen (up to 17.3 %). Moreover, for a given BT, the best Tcut-off is lower in the mild-dry climate than in the cold-humid one. For instance, when BT = 0 °C, the best Tcut-off is 0 °C in Copenhagen and -3 °C in Carpentras. This is a consequence of the lower average relative humidity in Carpentras during the winter period (73.8 %) than in Copenhagen (89.3 %). In fact, Figure 4 explains that for a given Text (lower than 6 °C), due to the reduction of the electric energy required for the evaporator defrost, the lower RH the higher the COP. Therefore, the lower the average RH during the winter period, the lower the best Tcut-off. Anyway, also with a mild-dry climate no (or negligible) advantages, in terms of primary energy consumption, are allowed by selecting a bivalent parallel system instead of a bivalent alternative one when BT is low (e.g. BT≤ 0 °C). Yet the efficiency of a bivalent parallel system is much higher than the efficiency of a bivalent alternative plant when BT is high (e.g. BT≥ 6 °C). Finally, in Carpentras, as well as in Copenhagen, the best alternative both from the energy and economic perspective (electric energy and natural gas prices respectively 20 c€ kWh-1 and 77 c€ Sm-3) is the bivalent system with BT = 0 °C, even if operating in parallel (Tcut-off = -3 °C) instead of alternative (Tcut-off = 0 °C) mode. 4. CONCLUSIONS
In this study a dynamic model of a hybrid system (heat pump and condensing boiler) was implemented and several annual simulations were carried out considering both a cold-humid and a mild-dry climate for a typical European family house built in 1970s with usual energetic renovation upgrades. The main goal of the simulations was to study how the choice of the cut-off temperature can influence the annual efficiency of the system and to understand if the best cut-off temperature, from an energy perspective, depends on heat-pump size and climate. With both the considered climates, when a large size heat pump is selected, there might be no (or not relevant) benefits by setting a Tcut-off lower than BT, and the differences between the energy performances of a bivalent alternative system and those of a bivalent parallel system are negligible. On the other hand, when a low size heat pump is preferred, a bivalent parallel system can lead to appreciable (up to 5 % in the cold-humid climate and up to 8 %in the mild-dry climate when BT = 6 °C) energy savings compared to a bivalent alternative plant. Moreover, the best Tcut-off shifts towards lower values for lower size heat pumps. From a primary energy perspective, the bivalent system with BT = 0 °C is the best solution with both the considered climates, even if the best Tcut-off is 0 °C in Copenhagen and -3 °C in Carpentras. In fact, for a given 9 Page 9 of 19
bivalent temperature of the system, the lower the average relative humidity in winter, the lower the best cut-off temperature. The main economic advantages of a hybrid system, with respect to the monovalent heat-pump plant, comes from the lower annual electric and gas energy needs. Anyway, an appreciable advantage is offered (up to 10.5 %) also in terms of initial investment as a lower size heat pump is required. REFERENCES AiCARR, Italian Association for Air Conditioning Heating and Refrigeration, 2015, Acqua calda sanitaria dimensionamento accumulo e potenze generatori/scambiatori Software, http://www.aicarr.org/Pages/Tools/non_soci_tools.aspx Arteconi, A., Newitt, N.J., Polonara, F., 2013, Domestic demand-side management (DSM): Role of heat pumps and thermal energy storage (TES) systems, Applied Thermal Engineering, 51: 155-165. Italian Authority for electricity, gas and water system,2015, http://www.autorita.energia.it/it/index.htm Bagarella, G., Lazzarin, R., Lamanna, B., 2013. Cycling losses in refrigeration equipment: An experimental evaluation, International Journal of Refrigeration, 36 (8): 2111-2118. Busato, F., Lazzarin, R., Noro, M., 2012, Energy and economic analysis of different heat pump systems for space heating, International Journal of Low-Carbon Technologies, 7: 104-112. Cecchinato, L., Chiarello, M., Corradi, M., 2010. A simplified method to evaluate the seasonal energy performance of water chillers, International Journal of Thermal Sciences, 49: 1776-1786. Di Perna, C., Magri, G., Giuliani, G., Serenelli, G., 2015, Experimental assessment and dynamic analysis of a hybrid generator composed of an air source heat pump coupled with a condensing boiler in a residential building, Applied Thermal Engineering, 76: 86-97. Domanski, P., Yashar, D., Kim, M., 2005, Performance of a finned-tube evaporator optimized for different refrigerants and its effect on system efficiency, International Journal of Refrigeration, 28 (6): 820-827. Dong, J., Deng, S., Jiang, Y., Xia, L., Yao, Y., 2012, An experimental study om defrosting heat supplies and energy consumptions during reverse cycle defrost for an air source heat pump, Applied Thermal Engineering, 37: 380-387. Emerson-Copeland, Selection Software v.7.11, 2015, http://www.emersonclimate.com/europe/en-eu/Resources/Software_Tools/Pages/Product_Selection_Soft ware.aspx Guo, X., Chen, Y., Wang, W., Chen, C., 2008, Experimental study of frost growth and dynamic performance of air source heat pump system, Applied Thermal Engineering, 28:n2267-2278. Lazzarin, R., 2012, Condensing boilers in buildings and plants refurbishment, Energy and Buildings, 47: 6167. 10 Page 10 of 19
Madani, H., Claesson, J., Lundqvist, P., 2011, Capacity control in ground source heat pump systems part I: modelling and simulation, International Journal of Refrigeration, 34: 1338-1347. Klein, K., Huchtemann, K., Muller, D., 2014, Numerical study on hybrid heat pump systems in existing buildings, Energy and Buildings, 69: 193-201. Klein, S.A. et al, 2010, TRNSYS 17: A Transient System Simulation Program, Solar Energy Laboratory, University of Wisconsin, Madison, USA, http://sel.me.wisc.edu/trnsys. NIST, National Institute of Standard and Technology, 2003, Simulation models for finned-tube heat exchangers, EVAP-COND, http://www2.bfrl.nist.gov/software/evap-cond/ SWEP, 2015, SSP-G7 software, http://www.swep.net/it/products_solutions/ssp_calculation_software/Pages/ssp_install.aspx Ente Nazionale Italiano di Normazione, 2012, UNI EN 14825: Air conditioners, liquid chilling packages and heat pumps, with electrically driven compressors, for space heating and cooling – Testing and raring at part load conditions and calculation of seasonal performance, Milan, Italy.
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Figure 1. Layout of the hybrid heat pump system developed in TRNSYS 17 environment.
Figure 2. Validation of the evaporator model. The average evaporator wall temperature and the average frost thickness profiles (dotted circles and arrows show to which axis each profile refers) were simulated and compared to the experimental values. A fixed temperature (0 °C) of the air at the evaporator inlet was considered while three different values of the relative humidity (RH) were tested.
12 Page 12 of 19
Figure 3. Cumulative profile of the building thermal energy demand and annual distribution of the temperature of the external air (Text) in Copenhagen (DK) and Carpentras (F) during the heating season.
Figure 4. Heat pump COP and thermal capacity profiles (dotted circles and arrows show to which axis each profile refers) as function of both the temperature (Text) and the relative humidity (RH) of the external air. The building thermal load for Copenhagen is also depicted.
13 Page 13 of 19
Figure 5. Annual primary electric energy, gas energy and total primary energy consumed by the system as function of the cut-off temperature (Tcut-off) when the bivalent temperature (BT) is 0 °C in Copenhagen (BAS = Bivalent Alternative System, BPS = Bivalent Parallel System).
Figure 6. Heat pump SCOP and annual number of on-off cycles as functions of the cut-off temperature (Tcut-off) when the bivalent temperature (BT) is 0 °C, 6 °C or 10 °C in Copenhagen.
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Figure 7. Annual primary electric energy, gas energy and total primary energy consumed by the system as function of the cut-off temperature (Tcut-off) when the bivalent temperature (BT) is 6 °C in Copenhagen (BAS = Bivalent Alternative System, BPS = Bivalent Parallel System).
Figure 8. Primary energy consumption of the hybrid system with different bivalent temperatures (BT). For each BT the cut-off temperature (Tcut-off) which minimizes the annual primary energy consumption in Copenhagen was chosen. Also monovalent heat pump (HP) and boiler systems were considered.b and s are respectively the annual average boiler and system efficiencies.
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Figure 9. Fifteen years Net Present Value (NPV) of both monovalent systems (HP and boiler) and hybrid systems with different bivalent temperatures (BT) in Copenhagen. Electric energy and gas prices are respectively 20 c€kWh-1 and 77 c€ Sm-3.
Figure 10. Fifteen years Net Present Value (NPV) of both monovalent systems (HP and boiler) and hybrid systems with different bivalent temperatures (BT) in Copenhagen. The gas price is 77 c€ Sm-3 while the electric energy price is varied.
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Figure 11. Fifteen years Net Present Value (NPV) of both monovalent systems (HP and boiler) and hybrid systems with different bivalent temperatures (BT) in Copenhagen. The electric energy price is 20 c€kWh-1while the gas price is varied.
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Table 1. Definition of the different hybrid heat pump systems Tcut-off
Tcut-off = BT(**)
Tcut-off ≤ Text
Text
HP ON ON OFF 2 heat OFF ON ON generator (*) Bivalent parallel operation - (**) Bivalent alternative operation
Text ≥ BT
Text
ON
OFF
OFF
ON
nd
Table 2. Main data of the modelled heat pump.
Evaporator
Condenser
Tube length (mm) Number of columns Number of rows
920 48 2
Air flow rate (m3 h-1) Fin Type Out tube diameter (mm) Fin thickness (mm) Fin pitch (mm) Tube spacing (mm)
6500 Wavy fin 10.0 0.15 1.95 25.4
Heat transfer area (m2) Number of channels Number of plates
2.52 41 42
Water flow rate (kg s-1)
0.7 -1
Compressor
Expansion Valve
Refrigerant flow rate (kg s ) Model Type Model
0.0861 Swep B25Tx42 Scroll Fixed Speed Copeland ZF24K4E_TWD
Internal Volume (dm3) Thermal Power (kW) Refrigerant Type
19.8 10.75 (Te = -10 °C, Tc = 55 °C) R134a EEV
Kv (m3 h-1)
0.066
Table 3. Main data of the considered building. Type
Single family house*
Net surface (m2)
132
Vertical walls transmittance (W m-2K-1) Roof transmittance (W m-2K-1) Ventilation Rate (h-1)
0.218 0.473 0.5 18 Page 18 of 19
Nominal Heating Load (Text = -13 °C) (kW)
8.0
Annual Thermal Demand (kWh)
Copenhagen: 21005 Carpentras: 10758
Heating distribution system
Radiators
*
Water supply temperature (°C)
Klein et al. (2014)
Heating curve**
**
Arteconi et al. (2013)
Table 4. Hypothesis of the economic analysis and annual costs of alternatives in Figure 8.
HP Cost (€) Boiler Cost (€) Total Investment (€)
HP 8.0 8.5 5759 1975 7735
BT= -6 °C 5.3 11.2 5319 2111 7430
BT = 0 °C 3.9 12.6 5076 2177 7252
BT = 6 °C 2.5 14.0 4846 2234 7080
BT = 10 °C 1.4 15.1 4670 2275 6945
boiler 0.0 16.5 0 2326 2326
Electric energy Cost (€ year-1)
1905
1712
1140
1002
764
97
Natural Gas Cost (€ year-1)
0
111
542
692
937
1617
Renewable Energy (kWh year )
11477
11850
9202
7800
5718
0
Incentives (€ year-1)
Pth HP (Text = -13 °C) (kW) Pth boiler (kW)
-1
631
652
506
429
315
0
-1
90
87
86
85
84
83
-1
1364
1259
1261
1349
1470
1797
Maintenance Cost (€ year ) Total annual Cost (€ year )
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S1359-4311(16)30039-4 http://dx.doi.org/doi: 10.1016/j.applthermaleng.2016.01.089 ATE 7649
To appear in:
Applied Thermal Engineering
Received date: Accepted date:
11-11-2015 19-1-2016
Please cite this article as: G. Bagarella, R. Lazzarin, M. Noro, Annual simulation, energy and economic analysis of hybrid heat pump systems for residential buildings, Applied Thermal Engineering (2016), http://dx.doi.org/doi: 10.1016/j.applthermaleng.2016.01.089. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
ANNUAL SIMULATION, ENERGY AND ECONOMIC ANALYSIS OF HYBRID HEAT PUMP SYSTEMS FOR RESIDENTIAL BUILDINGS (*)
G. BAGARELLA, (**) R. LAZZARIN, (***) M. NORO
(*) (**) (***)
Department of Management and Engineering, University of Padua President of IIR Section E, Air conditioning, heat pumps, energy recovery Str.lla San Nicola 3, Vicenza, 36100, Italy (*) [email protected] - (**) [email protected] - (***) [email protected] (**)
A model of a hybrid system (heat pump and condensing boiler) was implemented Simulations were run with both a cold-humid and a mild-dry climate With a large size HP there are no benefits by setting a Tcut-off lower than BT With a low size HP, a bivalent parallel system can lead to energy savings
ABSTRACT Air source heat pumps have a well known unfavorable characteristic: in the coldest period of the year, when the building heating load is at maximum, the heat pump capacity is reduced. A possible solution is to size the heat pump to cover only a fraction of the peak load using a second heat generator in “hybrid heat pump system”. In this study a dynamic model of a hybrid system was built and several seasonal simulations were carried out (considering two different climates) to study how the choice of the cut off temperature can influence the annual efficiency of the system and to understand if a bivalent parallel plant can lead to energy savings compared to a bivalent alternative system. With both the considered climates, when a large size heat pump is selected (low bivalent temperature), there might be no benefits by setting a cut off temperature lower than the bivalent temperature, and the differences between the energy performances of the bivalent alternative system and those of the bivalent parallel system are negligible. On the other hand, when a low size heat pump is preferred (high bivalent temperature), the bivalent parallel system can lead to appreciable energy savings compared to a bivalent alternative plant. The main economic advantages of a hybrid system, with respect to the monovalent heat-pump plant, comes from the lower annual electric and gas energy needs.
KEYWORDS Hybrid system, heat pump, boiler, cut off temperature, bivalent temperature
NOMENCLATURE 1 Page 1 of 19
BAS = Bivalent Alternative System BPS = Bivalent Parallel System BT = Bivalent Temperature (°C) COP = Coefficient of Performance Eel_HP= Annual electric energy absorbed by the heat pump (kWh) Eth_boiler= Annual thermal energy produced by the boiler (kWh) Eth_building= Building annual thermal energy demand (kWh) EEV = Electronic Expansion Valve HP = Heat Pump Kv = Valve Flow Coefficient (m3 h-1) NPV = Net Present Value (€) Pth = Thermal Capacity (kW) RH = Relative Humidity SCOP = Seasonal Coefficient of Performance Tcut-off = Cut-off Temperature (°C) Tc = Condensation Temperature (°C) Te = Evaporation Temperature (°C) Text=Temperature of the external air (°C) Tset =Set-point Temperature (°C) Tw out =Temperature of the water at the tank outlet (°C) b = boiler efficiency s = system efficiency 1. INTRODUCTION Heat pumps are a promising approach to reduce the production of greenhouse gases in the building sector, because they use freely available ambient heat. In recent years the use of heat pumps for space heating has spread widely. This is mainly due to the evolution of the technical features of devices, of their performances and of the electricity generation system efficiency. Furthermore, in Europe the Directives 2010/31/EU on energy performance of buildings and 2012/27/EU on energy efficiency introduced the compulsoriness of producing increasing part of a building thermal energy needs by renewables; energy 2 Page 2 of 19
from the heat source (air, water or ground) used by heat pumps is considered renewable by the Directives and this is giving more and more chances to the heat pumps market. Outside air is the most widespread heat source for heat pumps. Outside air has a well known unfavorable characteristic: in the coldest period of the year, when the building heating load is at maximum, the heat pump capacity is reduced (Klein et al., 2014). Moreover, being the peak load required for a small fraction of the heating period (Lazzarin, 2012), if the heat pump unit is sized to cover the peak load when the air temperature is at the minimum value, the annual average on-off frequency of the unit might be really high, lowering the seasonal coefficient of performance (SCOP) due to cycling losses (Bagarella et al., 2013). A possible solution is to size the heat pump to cover only a fraction of the peak load. In this case, a second heat generator (usually electric heaters or a boiler) is required and the system is called “hybrid heat pump system”. When speaking about hybrid heat pump systems some definitions should be reminded. The bivalent temperature BT (Ente Nazionale Italiano di Normazione, 2012) may be defined as the temperature at which the thermal capacity of the heat pump equals the building thermal load. Thus, for a given building, the higher BT the smaller the heat pump size. The cut-off temperature (Tcut-off) is the temperature of the external air below which the heat pump is switched-off by the controller. If the temperature of the external air (Text) is higher than BT, the thermal power is produced only by the heat pump unit. If Tcut-off
cut-off temperature: should the choice of Tcut-off take into account the heat pump size and the climate? Moreover, to the authors’ best knowledge, there is still lack of information regarding the differences between a bivalent parallel system and a bivalent alternative plant, in terms of energy performance, for different climates. In this study a dynamic model of a hybrid system was built and several seasonal simulations were carried out to study how the choice of Tcut-off can influence the annual efficiency of the system and to understand if a bivalent parallel plant can lead to energy savings compared to a bivalent alternative plant. In order to generalize some of the main conclusions, simulations were run considering two different climates. 2. METHODOLOGY The model of the whole hybrid heat pump system was developed in TRNSYS 17 environment (Klein et al., 2010). As depicted in Figure 1, the air-to-water heat pump (HP) and the boiler operate in parallel mode, which is the most widespread solution in existing hybrid system applications (Klein et al., 2014). The system controller establishes whether to activate the heat pump only, the boiler only or both the heat generators, depending on the temperature of the water at the tank outlet (Tw out) and on the temperature of the external air. If Tw out is lower than the set-point (Tset) minus a dead-band (1.5 °C), at least one heat generator is activated. In particular, if Text is higher than the bivalent temperature, only the heat pump is activated. If Text is lower than BT but higher than the cut-off temperature, both the heat generators operate in parallel mode. Finally, if Text is lower than Tcut-off, only the boiler is activated. Different heat pump sizes were tested in the simulations, as well as different values of Tcut-off. On the contrary, the volume of the thermal storage was kept constant (0.3 m3). This volume allowed all the tested heat pump units to have a maximum on-off frequency lower than 5 switches per hour. The thermal efficiency of the condensing boiler was considered to be a function of the water temperature at the boiler inlet. In particular, the thermal efficiency profile suggested by Lazzarin (2012) was used. A model of a 10.8 kW (at Te= -10 °C, Tc = 55 °C, RH = 50 %) on-off air-to-water heat pump was firstly developed. Main information about each component are summarized in Table 2. The compressor was modelled basing on refrigerant mass flow rate and electric power profiles suggested by the compressor manufacturer (Emerson-Copeland, 2015). The plate condenser was simulated thanks to SSP-G7 calculation software (SWEP, 2015), developed by the condenser manufacturer. This software allows to simulate the thermal power exchanged by each plate condenser model varying the inlet conditions of both the fluids (water – R134a). The finned coil evaporator was modelled with the software EVAP-COND developed by NIST (National Institute of Standards and Technology, 2003). This software, also used in the past by Domansky et al. (2005), allows to design a finned coil evaporator and to simulate the thermal power 4 Page 4 of 19
exchanged in different test conditions. Moreover, the software estimates both the air temperature and air relative humidity in each point of the heat exchanger surface. Therefore, it was considered a useful tool in order to roughly simulate the evaporator frosting process, as the thermal resistance of both fins and tubes can be set, at each time-step, basing on the frost properties (thickness, density and conductivity). Being aware of the complexity of the frost formation problem, the same unit and the same test conditions considered by Guo et al. (2008) were firstly simulated, in order to validate our simplified evaporator model. In Figure 2, results experimentally obtained by Guo et al. (2008) are compared with those estimated by the simplified model. Three different tests were carried out, keeping the temperature of the external air constant (0 °C) with three values of the relative humidity (RH). Figure 2 shows the good agreement between the evaporator wall temperature and the frost thickness profiles estimated by the model and the values obtained by the survey of the real frost formation process, at least for the purposes of this study. As the approximation of the model becomes unacceptable when the average frost thickness is higher than 0.6 mm, the annual simulations have been carried out defrosting (by electric heaters) when the average frost thickness is close to that value. The defrosting efficiency (i.e. the ratio between the actual amount of energy required to melt the accumulated frost to the total amount of heat supply to the coil during the entire defrosting operation) was considered to be 60 %, as suggested by Dong et al. (2012). The models of the heat pump components were linked together as in Madani et al. (2011). To speed up the annual simulation of the unit (a one minute time step was required to simulate the evaporator frosting process with a sufficient approximation), superheating and subcooling were considered constant as in Cecchinato et al. (2010). REFPROP 9.0 was used to estimate the properties of the refrigerant in each operating condition. Cycling losses, defined as those inefficiencies which characterize the start-up period of a unit, were also considered basing on several measures reported in Bagarella et al. (2012). The same renovated building described by Klein et al. (2014) was considered, which consists in a typical European one family house, built in 1970s, with two floors, 132 m2 net living space and with typical energetic retrofitting interventions (14 cm insulation of outer walls, new coated double glazing windows and 18 cm non-walkable attic floor insulation). These interventions are aimed at reducing the energy needs of an old building without modifying the heating distribution system. More information about building properties and internal set points are described in that paper. The heating distribution system is made up of radiators and the heating curve proposed by Arteconi et al. (2013) was used (45 °C ≤Tset ≤ 55 °C as in Table 3). Among the possible European climates, a cold-humid climate (Copenhagen, Denmark) and a mild-dry climate (Carpentras, France) were here chosen (Figure 3). For the sake of brevity detailed results are here 5 Page 5 of 19
reported only for the cold-humid climate (where the issue of defrosting is especially important) while main results obtained with the mild-dry climate are summarized in section 3.4. 3.
RESULTS
3.1 Heat pump unit performances Some simulations were firstly run to characterize the performance of the heat pump unit. Figure 4 shows the COP profiles as a function of both the temperature of the external air and the relative humidity. To a better understanding of the heat pump performance, Tset was regarded as constant (60 °C) in Figure 4, while it was assumed a function of Tset in the annual simulations. COP profiles are coherent with those experimentally found by Di Perna et al. (2015) with a similar unit. When the frosting process does not occur (Text> 6 °C), the higher the RH the higher the COP, because the latent heat coming from air dehumidification increases the thermal capacity of the heat pump. Moreover, the higher the RH, the lower the maximum Text at which the frosting process occurs. For instance, if RH is 75% frosting appears when Text is lower than 6 °C, while it occurs at 4 °C if RH is 85%. In fact, the lower the RH the higher the contribution of the sensible heat over the total heat transferred by the air, thus the lower are both the temperature of the air at the evaporator outlet and the average temperature of the heat exchanger. When the frosting process occurs, the higher the RH the lower the COP, because energy is required for defrosting. Finally, simulations show that when RH is lower than 50%-55% the frosting process can be considered negligible. This happens either because the temperature of the evaporator surface is higher than the dew point or because the amount of vapour condensed is so little that the frost growth rate is negligible. Figure 4 also depicts both the heat pump thermal capacity and the building thermal load profiles. Thus, with this particular heat pump size, the bivalent temperature is close to -13 °C. 3.2 Annual simulations: energy analysis To better understand how the cut-off temperature affects the energy performance of the system, several simulation were run varying both Tcut-off and the bivalent temperature (thus the heat pump size). Figure 5 shows, for different Tcut-off, the annual primary energies required by both heat pump unit and boiler when BT is 0 °C (heat pump capacity 3.9 kW when Text= -13 °C and RH = 50 %). If Tcut-off is equal to BT (0 °C) the system operates as a bivalent alternative system (BAS).On the other hand, if Tcut-off is lower than BT, the system operates as a bivalent parallel system (BPS) and both the heat generators are activated when Tcut-off
the other hand, the lower Tcut-off, the lower the SCOP but the lower the thermal energy produced by the boiler. The minimum total annual primary energy consumption is obtained when Tcut-off is around 0 °C, thus when it is close to BT. Considering this cold-humid climate, when BT is close to 0 °C, the differences between the energy performance of a parallel or alternative system are negligible. The annual simulations were run again considering a smaller heat pump size (BT = 6 °C). Figure 7 shows that, when a high BT is chosen, the situation is different and there is an appreciable energy saving (up to 5 %) by adopting a bivalent parallel system (Tcut-off< 6 °C) instead of a bivalent alternative system (Tcut-off = 6 °C). The reason is that the bivalent parallel system can drastically decrease the annual thermal energy produced by the boiler, as in Figure 7. The minimum total annual primary energy is obtained when Tcut-off = -1 °C, while the best value was 0 °C when the considered BT was 0 °C. Thus, the best Tcut-off shifts toward lower values when BT increases. The reason is explained by Figure 6, which shows that the SCOP of the heat pump unit with BT = 6 °C is higher than the SCOP of the unit with BT= 0 °C. This is mainly the consequence of the lower number of on-off cycles (thus lower cycling losses) obtainable by lowering the heat pump size. In Figure 7, starting from high Tcut-off values, one could expect the total primary energy to rapidly increase once reached its minimum value (Tcut-off = -1 °C). This does not happen because the annual thermal energy required by the building when Text is low, e.g. lower than – 4 °C, is a small fraction (lower than 10 %) of the total annual thermal energy required, as depicted in Figure 3. Thus, considering this cold-humid climate, when BT is close to 6 °C, the differences between the energy performances of the bivalent parallel system with the best Tcut-off and the same system with Tcut-off set to really low values (e.g. -12 °C) are very small. Other simulations were carried out varying BT from -6 °C to 10 °C. The selection of these temperatures is finalized to study the behavior of different size heat pumps, recalling that a high BT means a low capacity heat pump and vice versa. In Figure 8, for each BT, only the primary energy consumption of the best solution (Tcut-off which minimizes the total primary energy consumption) is reported. Moreover, also the monovalent systems (heat pump only and boiler only) were considered. In this cold-humid climate, the solutions with higher system efficiency (ratio between thermal energy produced and primary energy consumed) are the bivalent systems with BT = -6 °C and BT = 0 °C. 3.3 Annual Simulation: economic analysis An economic analysis was carried out to compare the Net Present Values (NPV) of the alternatives. As in Table 4, the nominal thermal power of the boiler (in each alternative) was heightened by 8.5 kW to ensure 7 Page 7 of 19
a sufficient power also for the hot sanitary water production (this value was estimated by the aid of a dedicated tool developed by AiCARR (Italian Association for Air Conditioning, Heating and Refrigeration, 2015), considering 4 people, 50 l consumption per person and 140 l hot water storage tank). Heat pumps and boilers prices were set basing on a manufacturer catalogues and applying a 40% discount. Electric energy and natural gas prices were initially considered to be respectively 20 c€ kWh-1 and 77 c€ Sm-3 (Italian Authority for electricity, gas and water system, 2015). An annual incentive was considered for each heat pump solution as suggested by Italian legislation: the annual renewable energy used by each alternative was estimated (Eth_building – (Eth_boiler + Eel_HP)) and the annual incentive was calculated applying a 0.055 € kWh-1 enhancement factor (Italian Authority for electricity, gas and water system, 2015). The maintenance costs were estimated as in Busato et al. (2012). Finally, a 3% annual interest rate was used. Figure 9 summarizes the NPV (15 years) of each alternative. The best solutions, from an economic perspective, are the bivalent parallel systems with BT = -6 °C (5.3 kW heat pump size) and BT = 0 °C (3.9 kW heat pump size). The main contribution, with respect to the monovalent heat pump solution, comes from the reduction of both electricity and gas consumption. Moreover, another contribution comes from the lower initial investment, which is the consequence of the lower size of the heat pump. A sensitivity analysis was also carried out varying the prices of both electric energy and natural gas (±30%). Figure 10 and Figure 11 show that the best alternative, from an economic perspective, is strongly influenced by these prices. For instance, the monovalent boiler solution might be the best alternative with low prices of the natural gas, but becomes the worst solution when the price is higher than 85c€ Sm-3. Anyway, the bivalent parallel solution with BT= 0 °C is less sensitive to prices variations and it is always close to the best alternative. 3.4 Summary of energy and economic analysis with another climate As explained in section 2, simulations were run considering a mild-dry climate (Carpentras) as well. For the sake of brevity detailed results (including figures and tables) of simulations with Carpentras’ climate are not reported here. The aim of this chapter is to summarize the main results of these simulations and to explain how and why the different climate affects the system performances and choice of both the best BT and best Tcut-off. The most obvious consequence of the different climate (Figure 3) is that SCOPs of heat pump units are much higher in Carpentras. This is due to both the higher average Text during the winter period, and the lower energy required for the evaporator defrost (up to 9.8 % of the total electric energy in Copenhagen and up to 5.8 % in Carpentras) in the mild-dry climate. For instance, the SCOPs of the heat pump with BT = -6 °C and Tcut-off= -6 °C are 2.39 and 2.82 respectively in Copenhagen and Carpentras. Therefore, the advantages (in terms of primary energy consumption) of hybrid systems with low BT (BT ≤ 0 °C) with 8 Page 8 of 19
respect to the monovalent boiler system are much higher in Carpentras (up to 29.6 %) than in Copenhagen (up to 17.3 %). Moreover, for a given BT, the best Tcut-off is lower in the mild-dry climate than in the cold-humid one. For instance, when BT = 0 °C, the best Tcut-off is 0 °C in Copenhagen and -3 °C in Carpentras. This is a consequence of the lower average relative humidity in Carpentras during the winter period (73.8 %) than in Copenhagen (89.3 %). In fact, Figure 4 explains that for a given Text (lower than 6 °C), due to the reduction of the electric energy required for the evaporator defrost, the lower RH the higher the COP. Therefore, the lower the average RH during the winter period, the lower the best Tcut-off. Anyway, also with a mild-dry climate no (or negligible) advantages, in terms of primary energy consumption, are allowed by selecting a bivalent parallel system instead of a bivalent alternative one when BT is low (e.g. BT≤ 0 °C). Yet the efficiency of a bivalent parallel system is much higher than the efficiency of a bivalent alternative plant when BT is high (e.g. BT≥ 6 °C). Finally, in Carpentras, as well as in Copenhagen, the best alternative both from the energy and economic perspective (electric energy and natural gas prices respectively 20 c€ kWh-1 and 77 c€ Sm-3) is the bivalent system with BT = 0 °C, even if operating in parallel (Tcut-off = -3 °C) instead of alternative (Tcut-off = 0 °C) mode. 4. CONCLUSIONS
In this study a dynamic model of a hybrid system (heat pump and condensing boiler) was implemented and several annual simulations were carried out considering both a cold-humid and a mild-dry climate for a typical European family house built in 1970s with usual energetic renovation upgrades. The main goal of the simulations was to study how the choice of the cut-off temperature can influence the annual efficiency of the system and to understand if the best cut-off temperature, from an energy perspective, depends on heat-pump size and climate. With both the considered climates, when a large size heat pump is selected, there might be no (or not relevant) benefits by setting a Tcut-off lower than BT, and the differences between the energy performances of a bivalent alternative system and those of a bivalent parallel system are negligible. On the other hand, when a low size heat pump is preferred, a bivalent parallel system can lead to appreciable (up to 5 % in the cold-humid climate and up to 8 %in the mild-dry climate when BT = 6 °C) energy savings compared to a bivalent alternative plant. Moreover, the best Tcut-off shifts towards lower values for lower size heat pumps. From a primary energy perspective, the bivalent system with BT = 0 °C is the best solution with both the considered climates, even if the best Tcut-off is 0 °C in Copenhagen and -3 °C in Carpentras. In fact, for a given 9 Page 9 of 19
bivalent temperature of the system, the lower the average relative humidity in winter, the lower the best cut-off temperature. The main economic advantages of a hybrid system, with respect to the monovalent heat-pump plant, comes from the lower annual electric and gas energy needs. Anyway, an appreciable advantage is offered (up to 10.5 %) also in terms of initial investment as a lower size heat pump is required. REFERENCES AiCARR, Italian Association for Air Conditioning Heating and Refrigeration, 2015, Acqua calda sanitaria dimensionamento accumulo e potenze generatori/scambiatori Software, http://www.aicarr.org/Pages/Tools/non_soci_tools.aspx Arteconi, A., Newitt, N.J., Polonara, F., 2013, Domestic demand-side management (DSM): Role of heat pumps and thermal energy storage (TES) systems, Applied Thermal Engineering, 51: 155-165. Italian Authority for electricity, gas and water system,2015, http://www.autorita.energia.it/it/index.htm Bagarella, G., Lazzarin, R., Lamanna, B., 2013. Cycling losses in refrigeration equipment: An experimental evaluation, International Journal of Refrigeration, 36 (8): 2111-2118. Busato, F., Lazzarin, R., Noro, M., 2012, Energy and economic analysis of different heat pump systems for space heating, International Journal of Low-Carbon Technologies, 7: 104-112. Cecchinato, L., Chiarello, M., Corradi, M., 2010. A simplified method to evaluate the seasonal energy performance of water chillers, International Journal of Thermal Sciences, 49: 1776-1786. Di Perna, C., Magri, G., Giuliani, G., Serenelli, G., 2015, Experimental assessment and dynamic analysis of a hybrid generator composed of an air source heat pump coupled with a condensing boiler in a residential building, Applied Thermal Engineering, 76: 86-97. Domanski, P., Yashar, D., Kim, M., 2005, Performance of a finned-tube evaporator optimized for different refrigerants and its effect on system efficiency, International Journal of Refrigeration, 28 (6): 820-827. Dong, J., Deng, S., Jiang, Y., Xia, L., Yao, Y., 2012, An experimental study om defrosting heat supplies and energy consumptions during reverse cycle defrost for an air source heat pump, Applied Thermal Engineering, 37: 380-387. Emerson-Copeland, Selection Software v.7.11, 2015, http://www.emersonclimate.com/europe/en-eu/Resources/Software_Tools/Pages/Product_Selection_Soft ware.aspx Guo, X., Chen, Y., Wang, W., Chen, C., 2008, Experimental study of frost growth and dynamic performance of air source heat pump system, Applied Thermal Engineering, 28:n2267-2278. Lazzarin, R., 2012, Condensing boilers in buildings and plants refurbishment, Energy and Buildings, 47: 6167. 10 Page 10 of 19
Madani, H., Claesson, J., Lundqvist, P., 2011, Capacity control in ground source heat pump systems part I: modelling and simulation, International Journal of Refrigeration, 34: 1338-1347. Klein, K., Huchtemann, K., Muller, D., 2014, Numerical study on hybrid heat pump systems in existing buildings, Energy and Buildings, 69: 193-201. Klein, S.A. et al, 2010, TRNSYS 17: A Transient System Simulation Program, Solar Energy Laboratory, University of Wisconsin, Madison, USA, http://sel.me.wisc.edu/trnsys. NIST, National Institute of Standard and Technology, 2003, Simulation models for finned-tube heat exchangers, EVAP-COND, http://www2.bfrl.nist.gov/software/evap-cond/ SWEP, 2015, SSP-G7 software, http://www.swep.net/it/products_solutions/ssp_calculation_software/Pages/ssp_install.aspx Ente Nazionale Italiano di Normazione, 2012, UNI EN 14825: Air conditioners, liquid chilling packages and heat pumps, with electrically driven compressors, for space heating and cooling – Testing and raring at part load conditions and calculation of seasonal performance, Milan, Italy.
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Figure 1. Layout of the hybrid heat pump system developed in TRNSYS 17 environment.
Figure 2. Validation of the evaporator model. The average evaporator wall temperature and the average frost thickness profiles (dotted circles and arrows show to which axis each profile refers) were simulated and compared to the experimental values. A fixed temperature (0 °C) of the air at the evaporator inlet was considered while three different values of the relative humidity (RH) were tested.
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Figure 3. Cumulative profile of the building thermal energy demand and annual distribution of the temperature of the external air (Text) in Copenhagen (DK) and Carpentras (F) during the heating season.
Figure 4. Heat pump COP and thermal capacity profiles (dotted circles and arrows show to which axis each profile refers) as function of both the temperature (Text) and the relative humidity (RH) of the external air. The building thermal load for Copenhagen is also depicted.
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Figure 5. Annual primary electric energy, gas energy and total primary energy consumed by the system as function of the cut-off temperature (Tcut-off) when the bivalent temperature (BT) is 0 °C in Copenhagen (BAS = Bivalent Alternative System, BPS = Bivalent Parallel System).
Figure 6. Heat pump SCOP and annual number of on-off cycles as functions of the cut-off temperature (Tcut-off) when the bivalent temperature (BT) is 0 °C, 6 °C or 10 °C in Copenhagen.
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Figure 7. Annual primary electric energy, gas energy and total primary energy consumed by the system as function of the cut-off temperature (Tcut-off) when the bivalent temperature (BT) is 6 °C in Copenhagen (BAS = Bivalent Alternative System, BPS = Bivalent Parallel System).
Figure 8. Primary energy consumption of the hybrid system with different bivalent temperatures (BT). For each BT the cut-off temperature (Tcut-off) which minimizes the annual primary energy consumption in Copenhagen was chosen. Also monovalent heat pump (HP) and boiler systems were considered.b and s are respectively the annual average boiler and system efficiencies.
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Figure 9. Fifteen years Net Present Value (NPV) of both monovalent systems (HP and boiler) and hybrid systems with different bivalent temperatures (BT) in Copenhagen. Electric energy and gas prices are respectively 20 c€kWh-1 and 77 c€ Sm-3.
Figure 10. Fifteen years Net Present Value (NPV) of both monovalent systems (HP and boiler) and hybrid systems with different bivalent temperatures (BT) in Copenhagen. The gas price is 77 c€ Sm-3 while the electric energy price is varied.
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Figure 11. Fifteen years Net Present Value (NPV) of both monovalent systems (HP and boiler) and hybrid systems with different bivalent temperatures (BT) in Copenhagen. The electric energy price is 20 c€kWh-1while the gas price is varied.
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Table 1. Definition of the different hybrid heat pump systems Tcut-off
Tcut-off = BT(**)
Tcut-off ≤ Text
Text
HP ON ON OFF 2 heat OFF ON ON generator (*) Bivalent parallel operation - (**) Bivalent alternative operation
Text ≥ BT
Text
ON
OFF
OFF
ON
nd
Table 2. Main data of the modelled heat pump.
Evaporator
Condenser
Tube length (mm) Number of columns Number of rows
920 48 2
Air flow rate (m3 h-1) Fin Type Out tube diameter (mm) Fin thickness (mm) Fin pitch (mm) Tube spacing (mm)
6500 Wavy fin 10.0 0.15 1.95 25.4
Heat transfer area (m2) Number of channels Number of plates
2.52 41 42
Water flow rate (kg s-1)
0.7 -1
Compressor
Expansion Valve
Refrigerant flow rate (kg s ) Model Type Model
0.0861 Swep B25Tx42 Scroll Fixed Speed Copeland ZF24K4E_TWD
Internal Volume (dm3) Thermal Power (kW) Refrigerant Type
19.8 10.75 (Te = -10 °C, Tc = 55 °C) R134a EEV
Kv (m3 h-1)
0.066
Table 3. Main data of the considered building. Type
Single family house*
Net surface (m2)
132
Vertical walls transmittance (W m-2K-1) Roof transmittance (W m-2K-1) Ventilation Rate (h-1)
0.218 0.473 0.5 18 Page 18 of 19
Nominal Heating Load (Text = -13 °C) (kW)
8.0
Annual Thermal Demand (kWh)
Copenhagen: 21005 Carpentras: 10758
Heating distribution system
Radiators
*
Water supply temperature (°C)
Klein et al. (2014)
Heating curve**
**
Arteconi et al. (2013)
Table 4. Hypothesis of the economic analysis and annual costs of alternatives in Figure 8.
HP Cost (€) Boiler Cost (€) Total Investment (€)
HP 8.0 8.5 5759 1975 7735
BT= -6 °C 5.3 11.2 5319 2111 7430
BT = 0 °C 3.9 12.6 5076 2177 7252
BT = 6 °C 2.5 14.0 4846 2234 7080
BT = 10 °C 1.4 15.1 4670 2275 6945
boiler 0.0 16.5 0 2326 2326
Electric energy Cost (€ year-1)
1905
1712
1140
1002
764
97
Natural Gas Cost (€ year-1)
0
111
542
692
937
1617
Renewable Energy (kWh year )
11477
11850
9202
7800
5718
0
Incentives (€ year-1)
Pth HP (Text = -13 °C) (kW) Pth boiler (kW)
-1
631
652
506
429
315
0
-1
90
87
86
85
84
83
-1
1364
1259
1261
1349
1470
1797
Maintenance Cost (€ year ) Total annual Cost (€ year )
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