thermal collectors: Dynamic simulation and thermoeconomic optimization

Energy 95 (2016) 346e366

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A novel solar-assisted heat pump driven by photovoltaic/thermal collectors: Dynamic simulation and thermoeconomic optimization Francesco Calise a, Massimo Dentice d'Accadia a, Rafal Damian Figaj b, *, Laura Vanoli b a b

Department of Industrial Engineering, University of Naples Federico II, P.le Tecchio 80, 80125, Naples, Italy Department of Engineering, University of Naples Parthenope, Centro Direzionale. IS.C4, Naples, Italy

a r t i c l e i n f o

a b s t r a c t

Article history: Received 2 September 2015 Received in revised form 6 November 2015 Accepted 29 November 2015 Available online xxx

This paper presents a dynamic simulation model and a thermo-economic analysis of a novel polygeneration system based on a solar-assisted heat pump and an adsorption chiller, both driven by PVT (photovoltaic/thermal) collectors. The aim of this work is to design and dynamically simulate a novel ultra-high efficient solar heating and cooling system. The overall plant layout is designed to supply electricity, space heating and cooling and domestic hot water for a small residential building. The system combines solar cooling, solar-assisted heat pump and photovoltaic/thermal collector technologies in a novel solar polygeneration system. In fact, the polygeneration system is based on a PVT solar field, coupled with a water-to-water electric heat pump or to an adsorption chiller. PVT collectors simultaneously produce electricity and thermal energy. During the winter, hot water produced by PVT collectors primarily supplies the evaporator of the heat pump, whereas in summer, solar energy supplies an adsorption chiller providing the required space cooling. All year long, solar thermal energy in excess is converted into DHW (domestic hot water). The system model was developed in TRNSYS environment. 1year dynamic simulations are performed for different case studies in various weather conditions. The results are analysed on different time bases presenting energetic, environmental and economic performance data. Finally, a sensitivity analysis and a thermoeconomic optimization were performed, in order to determine the set of system design/control parameters that minimize the simple pay-back period. The results showed a total energy efficiency of the PVT of 49%, a heat pump yearly coefficient of performance for heating mode above 4 and a coefficient of performance of the adsorption chiller of 0.55. Finally, it is also concluded that system performance is highly sensitive to the PVT field area. The system is profitable when a capital investment subsidy of 50% is considered. © 2015 Elsevier Ltd. All rights reserved.

Keywords: Solar energy Solar cooling Heat pump Adsorption chiller Heating and cooling system PVT (photovoltaic/thermal)

1. Introduction In the past decades, a number of Countries, especially in EU, massively supported conventional solar technologies (solar thermal and solar photovoltaic) by public funding [1]. This resulted in a significant reduction of their capital cost, especially for PV (photovoltaic) systems (from 7.0 kV/kWp to 1.3 kV/kWp). Simultaneously, more and more attention has been paid to some innovative solar technologies other than PV and solar thermal, namely: solar power [2,3], solar heating and cooling [4,5], hybrid PVT (photovoltaic/thermal) solar collectors [6e9]. The majority of these novel solar technologies are implemented in the novel solar

* Corresponding author. Tel.: þ39 0815476709; fax: þ39 0815476777. E-mail address: rafal.fi[email protected] (R.D. Figaj). http://dx.doi.org/10.1016/j.energy.2015.11.071 0360-5442/© 2015 Elsevier Ltd. All rights reserved.

polygeneration [10,11] system included in this paper. In the followings, a brief description of such technologies is provided. PVT (photovoltaic/thermal) collectors are particularly promising, since they combine in a single component conventional photovoltaic collectors, PV, and conventional thermal SC (solar collectors) [12,13]. PVT collectors simultaneously produce electricity and heat [14]. PVT collectors are typically manufactured by covering the absorber of a conventional thermal collector with a suitable PV layer. Thermal energy is distributed to a fluid (typically air [15] or water [16e18]), whereas the PV layer produces electricity [12,13]. The overall result is the simultaneous production of electricity and heat [19]. In addition, the electrical efficiency of a PVT collector may be even higher than that of a conventional PV module, at least in case of low PVT operating temperatures [12,13,20]. The most common PVT configuration is the “sheet-andtube” [21], where a conventional SC thermal collector is equipped

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with a PV layer encapsulated in the absorber, although several alternative configurations are currently under investigation [12,13,22,23]. PVT electrical efficiency decreases when the operating temperature increases. Usually, the performance drop-off due to temperature increase is not too high: it is typically around 0.45%/ K for silicon cells [20], and even lower in case of novel PV materials, such as multi-junction solar cells which can approach a nominal efficiency of 40% [24,25]. In any case, a lower operating temperature of PVT collectors obviously leads to an improvement of both electrical and thermal efficiencies [26]. For this reason, PVT systems are typically used for low-temperature applications, such as domestic hot water production, floor heating, desiccant cooling [27,28]. A possible alternative for increasing the outlet temperature of the working fluid without decreasing the PVT electrical efficiency may consist in the use of a heat pump (mainly driven by the PVT electric output) [12,13,22]. The combination of PVT collectors and heat pumps is a special case of the category of the SAHP (solar-assisted heat pumps), in which solar thermal energy is used to enhance the COP (coefficient of performance) of a conventional HP (heat pump) providing space heating during the winter season [29,30]. SAHP typically include conventional flat-plate solar thermal collectors producing a hot fluid supplying heat to the evaporator of an electric vaporcompression heat pump. When SAHP system include PVT collectors, the compressor of the HP can be driven by the renewable electrical energy produced by the collectors, further enhancing the overall efficiency of the system [31]. SAHP systems are typically classified into two groups: direct and indirect coupled. In the first case, the refrigerant directly flows inside the PVT collectors. In the second case, PVT and HP working fluids are decoupled by means of a heat exchanger. Higher conversion efficiencies and lower capital costs are achieved by the direct configuration. Conversely, the indirect one is much more simple and flexible, especially during the summer season [32,33]. The majority of the studies available in literature only analyse the conventional direct SAHP arrangement based on an electric heat pump and flat-plate solar collectors. The combination with PVT systems is scarcely analyzed. In particular, Gorozable Chata et al. [32] investigated the thermal performance of a direct expansion solar-assisted heat pump with several refrigerants, using two different configurations for the collector, bare and with cover. The results showed that R-12 produces the highest value of COP, followed by R-22 and R-134A. A similar study was presented by Zhang et al. [34], who investigated the effects of the refrigerant charge on the performance of the system. In the optimum configuration, good system performance and feasible costs can be achieved [34]. Tagliafico et al. [35] proposed to use solar thermal collectors as thermal exchange units (evaporators) in a heat pump system, in order to improve the efficiency and the economic profitability of the system. The authors used a simplified approach and they found results consistent with those available in literature, with a mean primary energy saving of about 50% with respect to a standard gas burner. A similar approach was also used by Scarpa et al. [36], using a model developed around the fluidindependent Carnot cycle. The system produces hot sanitary water and it is equipped with an auxiliary gas burner. The authors found results in accordance with those available in the literature [36]. The direct expansion SAHP, including solar thermal collectors, was also investigated by Chow et al. [37]. The authors presented a numerical model of the system. Then, a simulation was performed using the TMY (typical meteorological year) weather data of Hong Kong: a year average COP (coefficient of performance) of 6.46 was found. A more complex direct SAHP was presented by Chaturvedi et al. [38], including a double-stage compression for high temperature applications. Results are presented and compared with those of a single-stage direct expansion solar-assisted heat pump. The

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authors concluded that significant improvement of the thermal performance is achieved at high condensing temperatures when using the double-stage SAHP system. However, a higher capital cost (due to the larger solar field area) must be taken into account. A more comprehensive analysis of the papers available in literature investigating direct SAHP system was presented by Kara et al. [39], also including energy and exergy models. From this study, it is clear that only a couple of papers investigate PVT collectors, whereas the remaining ones focus on conventional solar thermal collectors [39]. One of the few studies investigating SAHP including PVT collectors was presented by Chow et al. [40]. The authors modeled a PV-SAHP (photovoltaic-integrated solar heat pump) system, using a dynamic simulation model and the TMY weather data of Hong Kong. It was found that the proposed system with R-134a is able to achieve a year average COP of 5.93 and a PV efficiency of 12.1%; the energy output is therefore considerably higher than that of a conventional heat pump plus PV “side-by-side” system. The indirect SAHP configuration was investigated by Sterling et al. [41]. The authors designed and modeled an indirect SAHP for DHW (domestic hot water) production, which was compared to a conventional SDHW (solar domestic hot water) system and to an electric domestic hot water system. The simulations were performed by TRNSYS software, using a simple scheme and basic control strategies. It was found that the best performance, considering the electrical consumption and the operating cost, was achieved by the SAHP system. In this study, a solar field equipped with conventional solar thermal collectors was considered [41]. An indirect SAHP configuration also including PVT collectors was recently presented by Hazi and Hazi [33]. The authors presented a comparative study of indirect photovoltaic thermal solar-assisted heat pump systems for water heating in industry is presented. Both steam ejector heat pumps and mechanical compression heat pumps were evaluated, for an application in a paper mill. A numerical model was implemented including energy, exergy and economic balances for a Romanian climatic condition. The authors concluded that in winter the operation time of the heat pump is shorter than the duration of the solar radiation, while in summer, when the air temperature and the solar radiation are higher, such operation time equals the duration of the solar radiation. SAHP systems are also studied in a plurality of applications, such as: swimming pools [42,43], solar heating and cooling [21], desalination [44], geothermal heat pumps [45], etc. A considerable number of studies are also available in literature presenting different experimental analyses of different SAHP configurations [46e54], including a number of different solar devices (flat-plate collectors, PVT collectors, evacuate tube collectors, variable speed compressors, etc). All the papers investigating the experimental performance of SAHP systems are based on small systems, and mainly focus on the direct configuration arrangement. As mentioned before, SAHP systems are specialized in the conversion of solar heat space heating energy during the winter season. Conversely, in summer, solar energy can be converted in space cooling energy by solar cooling systems [55,56]. In particular, this technology is particularly promising since the availability of solar radiation is simultaneous with building space cooling demand. Conversely, in winter time, the maximum heating demand often occurs in case of extremely scarce solar radiation [57]. The basic principle is simple since solar heat is delivered to a heat-driven chiller (absorption [5], adsorption [58], desiccant [59], etc.) converting such heat into cooling energy [5,56]. The majority of the literature studies and commercial systems are based on evacuated tubes solar thermal collectors and a single-effect absorption chiller [1,60]. In climates where the availability of solar beam radiation is extremely high, the combination of concentrating solar collectors and a double-effect absorption chiller may be profitable [57]. In this case, a higher solar collector outlet temperature is required to drive

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the chiller. Conversely, for colder climates solar collector temperature is lower and an adsorption chiller is usually employed [61] since such chillers can be activated for lower hot fluid temperatures. Literature review regarding solar cooling systems equipped with adsorption chiller is very rich [56,58,61,62]. A number of papers are available in literature presenting both numerical [63,64] and experimental analyses [65,66] of solar cooling systems equipped with adsorption chillers. For example, Gonzales et al. developed a solar adsorption cooling system based on CPC (compound parabolic collectors) where the tube includes the sorption bed. The system was designed and experimented, showing that the measured solar COP ranged from 0.078 to 0.096 [67]. The same collector technology (CPC) was investigated by Lu et al. [68]. They coupled the CPC collectors to an adsorption chiller and found that the average solar COP was 0.16. Such figure may increase up to 0.19, when the hot water temperature is 125  C [68]. Similarly a number of novel solar adsorption prototypes are diffusely discussed in literature [69e71]. In the recent few years, a new layout of solar cooling system has been also designed. This novel layout replaces the conventional solar thermal collectors by PVT collectors. Such arrangement allows one to enhance the overall system energetic efficiency since a certain amount of electricity is also produced in addition to the conventional space heating and cooling energy [21,24,44,72e74]. In particular, Calise et al. designed and simulated a hybrid solar trigeneration system including flat-plate PVT collectors and an absorption chiller [21]. They found that the outlet temperature of the PVT collectors were adequate to drive the absorption chiller, whereas a significant decrease of the electrical efficiency was achieved as a consequence of the increase of the operating temperature. System performance significantly improved when highly efficient CPVT (concentrating PVT) collectors, using III-V PV cells, are employed [72,74]. In case of a dish CPVT collector a double-stage absorption chiller can be driven [72], whereas using linear parabolic CPVT collectors a single-stage absorption chiller must be considered [74]. All these studies regarding solar trigeneration systems are based on absorption chiller. The possibility to combine CPVT and adsorption chiller was investigated only by Garcia-Heller et al. [75]. The authors performed an exergoeconomic analysis of a 10 MW 2000 suns CPVT solar field. Authors investigated the possibilities to use CPVT heat to drive an adsorption chiller, pointing out that a possible increase of the COP of this device is crucial for achieving good energetic, exergetic and economic performance [75]. As a summary, a large number of papers is available in literature presenting both numerical and experimental analyses of SAHP, PVT and adsorption solar cooling systems. As expected, the papers investigating the system from the experimental point of view only focus on simple system layouts, often simulating user load demands. As for numerical studies, the majority of the papers found in literature only focus on the direct SAHP configuration, whereas only few studies are available in which the indirect configuration is investigated. Such studies are often based on simple models and/or they just consider solar thermal collectors, rather than PVT collectors. Therefore, the indirect configuration is scarcely studied, especially when coupled with PVT collectors. In this case, no study was found in literature presenting a 1-year simulation model of an indirect SAHP system including PVT collectors, based on realistic load demands. Moreover, no paper was found in literature investigating the possible integration of SAHP systems with an adsorption chiller in order to supply also space cooling energy in addition to the space heating one. In particular, no study was found in literature investigating the novel layout presented in this paper, where PVT collectors drive a SAHP and an adsorption chiller, respectively for the winter and summer operation. In this paper,

authors aim at improving the knowledge on this topic presenting a dynamic model of this indirect SAHP system based on PVT and adsorption chiller technologies. With respect to the papers mentioned above [33,41], a much more complex system layout, developed with TRNSYS software, is proposed, including all the components required for operating the system (mixers, valves, controllers, heat exchangers, tanks, collectors, etc.). In addition, some innovative control strategies are here presented, aiming at maximizing the utilization of solar energy in all the operating conditions. Finally, differently from the majority of the studies available in literature, this work presents a much more detailed methodology for the assessments of building loads. In fact, system space heating and cooling and domestic hot water demands are calculated on the basis of detailed simulation models of a specific building, developed in TRNBUILD, included in TRNSYS package. The model is also completed by a detailed thermo-economic optimization aiming at calculating the capital and operating costs in all the configurations investigated, in order to determine the costoptimal system configuration. 2. System layout A simplified layout of the system under investigation in the paper is shown in Fig. 1, where only the main components are displayed. The proposed system is used to provide space heating during the winter (from 16th November to 31st March) and space cooling during the summer (from 1st June to 31st October). Furthermore, the system produces DHW (domestic hot water) and electric energy (all year long). The proposed plant is a small size PVT system installed on a residential use building. Here, the overall amount of electric energy provided by a small or medium size PV system in a year is much lower than the corresponding demand of the building, even in the absence of electric powered heating and cooling equipment (heat pumps or chillers). Thus, it is assumed that the electric energy provided by the PVT system is completely consumed by building electric equipment. The system layout includes several operating fluids and main components, reported in Tables 1 and 2, respectively. Additional BOP (balance of the plant) components (not displayed in Fig. 1, for brevity) have been also implemented in the simulation model, such as: pipes, sensors, controllers (proportional and on/off with hysteresis), schedulers (daily, weekly and seasonal). In particular, the pipes were used in order to simulate the ductwork. Other mandatory components are: weather database, plotters, printers, integrators, etc. The design parameters for all the main components of the system were accurately selected in order to allow the system to operate properly, meeting the energy demands of the building. Such values are summarized in Table 3. In particular, set-point temperatures and temperature deadbands for the components were selected in order to minimize the number of on-off events. The presented ST (solar trigeneration) system was dynamically simulated in TRNSYS environment [48]. The basic operating principle of the system can be here summarized as follows. Solar irradiation is converted into electrical energy and thermal energy through the PVT field. The produced electrical energy is completely consumed by the user, due to the operation of the reversible heat pump and the remaining electric equipment of the building. The solar thermal energy produced is used to heat up the SCF (solar collector fluid) circulating through the PVT field. The water flow rate of the variable speed pump P1 is regulated by a proportional controller activated when the radiation is higher than 10 W/m2 [76] (see Fig. 2). This controller varies continuously P1 flow rate, in order to achieve the desired PVT outlet temperature.

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Fig. 1. Simplified layout of the proposed system including the main components and loops.

Table 1 Description of the system operating fluids. Operating fluid

Description

SCF e solar collector fluid HF e hot fluid

Water flowing from the solar field to the source sides of the tanks Water flowing from the load side of tank to the evaporator of the heat pump (winter) and to the hot side of the adsorption chiller (summer) Water flowing from the surface aquifer to the condenser of the heat pump or to the cooling side of the adsorption chiller (summer) Water used for domestic devices

CW e cooling water

DHW e domestic hot water CHW e chilled water HEW e heated water

Circulating chilled water in the building hydronic system Circulating heated water in the building hydronic system

The set point temperature at the outlet of the PVT depends on the tank to be charged (TK1 and TK2) and on the season (winter and summer). In fact, TK1 has to provide heat to the evaporator of the water-to-water heat pump and to the hot side of the adsorption chiller in winter and in summer respectively, whereas TK2 is used for DHW production all year long. In particular, in winter PVT set point temperature is 30  C when TK1 is supplied by the collectors. Conversely, in summer when the same TK1 has to be charged, the set point temperature is set at 70  C. Moreover, when TK2 has to be supplied by solar thermal energy, the set point temperature is set at 55  C, independently from the season. This strategy is performed in order to maximize PVT electric efficiency, increasing in case of low operating temperatures, operating at the lowest temperature compatible with user demands (HP in winter and ADS in summer).

Table 2 Description of the system main components. System main component

Description

PVT

Photovoltaic/thermal collector field with water-cooled sheet-and-tube planar solar collectors with a cover glass and a photovoltaic panel placed above the absorber, whose inclination is 30 and with south orientation Thermal storage system that supplies thermal energy to the evaporator of the heat pump (winter) and to the hot side of the adsorption chiller (summer), and consists of a vertical stratified hot storage tank equipped with inlet stratification devices Thermal storage system that stores the eventual solar heat in excess (when TK1 is thermally loaded) for producing DHW, and consists of a vertical stratified hot storage tank with an internal heat exchanger Plate-fin heat exchanger in the solar loop used to dissipate the solar thermal energy when the solar irradiation is higher than the HP (winter) or the adsorption chiller, ADS, (summer) thermal demand and when both tanks are thermally loaded Gas-fired burner used to produce the building DHW when TK2 top temperature is below the required set-point Reversible heat pump with a reversing valve used to reverse the refrigerant flow from the compressor through the condenser and evaporation coils; HP supplies space heating energy in winter and auxiliary space cooling energy in summer when ADS capacity is lower than building demand Adsorption chiller with zeolite adsorption material, powered thermally by the HF (summer) Cold water surface aquifer well that provides cooling water CW to the condenser of the HP or to the cooling side of the ADS in the summer operation mode Inertial hot/chilled water storage tank used in order to reduce the number of on/off events of the reversible heat pump HP and the adsorption chiller ADS Fan coil system that supplies thermal energy for space heating and cooling, as needed Variable-speed pump a variable-speed pump, P1, for the SCF loop Fixed-volume pumps for the HF, CHW/HEW and CW loops Flow diverters and mixers used in order to properly manage the fluid circulating within the loops

TK1

TK2 HE

GB HP

ADS WE TK3 FC P1 P2/3/4/5 D, M

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Table 3 System design parameters. Parameter

Value [unit]

PVT field area Reference temperature for PV efficiency PVT slope PVT azimuth PVT set point temperature (when TK1 is supplied) PVT set point temperature (when TK2 is supplied) Nominal P1 flow rate per PVT area TK1 volume/PVT area TK1 height TK1 set point temperature TK1 activation deadband TK2 volume/PVT area TK2 height TK2 set point temperature TK2 activation deadband TK3 volume TK3 height TK3 set point temperature winter TK3 activation deadband winter TK3 set point temperature summer TK3 activation deadband summer HE activation temperature TK1, TK2, TK3 loss coefficient P3 flow P4 flow P5 flow HP rated heating capacity HP rated cooling capacity Space heating set point temperature Space cooling set point temperature

20 [m2] 25 [ C] 30 [ ] 0 [ ] 30/70 [ C] 55 [ C] 50 [kg/(h m2)] 100 [l/m2] 2 [m] 25/65 [ C] 5 [ C] 50 [l/m2] 1.6 [m] 50 [ C] 3 [ C] 1 [m3] 1 [m] 47 [ C] 2 [ C] 6 [ C] 2 [ C] 70 [ C] 0.278 [W/(m2  C)] 2400 [kg/h] 1500 [kg/h] 3600 [kg/h] 8 [kW] 7 [kW] 20 [ C] 26 [ C]

The running logic of the solar loop operation is shown in Fig. 2. PVT outlet fluid (SCF) is managed through two diverters and two mixers, D1, D2 and M1, M2 respectively. In this way, PVT field can supply alternatively TK1 and TK2. A sensor measures the temperature at the top of tank TK1; in winter, when such temperature falls outside the fixed range (25e20  C), the control system switches the flow of D2 to TK1 (see Fig. 2). In summer, this same

operation is done when the TK1 top temperature is not within the fixed deadband (60e65  C) [77] in order to achieve a proper ADS operation. The control system manages the solar thermal energy in order to supply primarily TK1, only when this tank is thermally loaded, SCF is moved to the TK2 internal heat exchanger, in order to produce DHW. If the DHW energy demand is not satisfied by solar energy, and therefore TK2 top temperature is below 45  C, the control system turns GB (gas burner) on. In this way, even in case of scarce solar irradiation, the system provides the heat required for matching DHW demand. In winter and summer, when both tanks are thermally loaded, SCF is bypassed, in order to avoid TK2 overheating of. In this operation, the SCF flow entering D1 is supplied straight to M1 (see Fig. 2), therefore the SCF is re-circulated through the closed loop constituted by PVT collectors, HE and P1. The by-pass operation is also performed in order to avoid TK2 cooling down when SCF temperature is lower than TK2 top temperature. In fact, SCF is bypassed as long as its temperature is not 2  C higher than TK2 top temperature. Then, SCF is supplied to the TK2 internal heat exchanger in order to produce DHW (see Fig. 2). Note that PVT outlet temperature may overcome the set point when the user thermal energy demand is low and/or solar radiation is high. In such case, in order to avoid SCF overheating and a consequent decrease of PVT electric efficiency, PVT fluid is cooled down by the heat exchanger HE, dissipating the solar thermal energy in excess. Conversely, when the user thermal energy demand is high and the solar irradiation is low, PVT outlet temperature may be lower than the fixed set point. The running logic of the ADS and HP circuits is show in Fig. 3. In the winter period, HF (hot fluid) is supplied by the load side of tank TK1 to P2, that pumps the flow across D3 and M5 towards the evaporator side of the reversible HP. HF temperature is lowered by the heat exchange within HP evaporator. Then, HF returns to the TK1 tank across D5 and M3. In order to meet the building space heating demand, HP is activated in order to maintain the top temperature of the inertial tank TK3 within the fixed deadband

Fig. 2. Control strategy: management of solar thermal energy, and tanks operation represented by a flowchart diagram.

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Fig. 3. Control strategy: management of heat pump (HP) and adsorption chiller (ADS) operation represented by a flowchart diagram.

(48-45  C) (see Fig. 3). When TK3 top temperature drops to 45  C, HP is turned on and HEW (heated water) is pumped by P3 to the condenser of the reversible HP. HEW passed through the HP condenser is supplied to the TK3 source side across M6. In this way, when the room temperature falls below 19  C, the hot fluid stored in TK3 is supplied by P4 to the fan-coils. In this case, the room temperature rises up to 21  C. In the summer period, HF is supplied by the load side of tank TK1 to P2, that pumps the flow across D3 towards the hot side of the adsorption chiller ADS (see Fig. 3). In order to dissipate the heat rejected by ADS, CW is pumped from WE by P5 towards the cooling side of ADS across D4. CW exiting ADS passes through M4 and is rejected to the sewer. The space cooling is provided by activating ADS in order to control the top temperature of the inertial tank TK3 within a fixed deadband (12e14  C). When the TK3 top temperature rises to 14  C ADS is turned on and CHW is pumped by P3 to the chiller side of ADS. CHW passed through the ADS is supplied to the TK3 source side across M6. In this way, when the room temperature exceeds 27  C, the chilled water stored in TK3 supplies the fan-coils inside the building. In this case, the room temperature decreases to 25  C. If TK1 top temperature is not high enough (below 55  C), no solar thermal energy is supplied to ADS. In this case, in order to maintain the TK3 top temperature within the fixed deadband (12e14  C), the reversible HP is turned on (see Fig. 3). Therefore, the condenser of the reversible heat pump HP is connected to the CW cooling circuit through D4, M5 and D5, M4, in order to dissipate the heat rejected by the condenser. Thus, the space cooling is provided by activating the HP, as a chiller, in order to control the top temperature of the inertial tank TK3. Moreover, when the space cooling demand is high and the TK1 top temperature is low (60  C) the TK3 top temperature may rise above 14  C even if the ADS is activated. If the inertial tank top temperature rises over 16  C ADS is deactivated and the reversible heat HP is turned on in order to cool down the tank to the lower deadband value (12  C). In this way, even in case of scarce solar irradiation and high space cooling demand the system provides required cooling power. It is worth noting that the set points temperatures of the components are related to the optimum operating range of the equipment, e.g. heat pump or adsorption chiller. In particular, the set point temperatures were selected in order to achieve the nominal capacities/performance of these components. Moreover, PVT set point temperature is maintained as low as possible in order to enhance collectors efficiency and to provide a proper equipment operation. However, in Section 5 a sensitivity analysis is also presented in order to investigate the effect of the variation of such set-point temperatures on the overall system energetic and economic results.

Summarizing the system operation (Table 4). - in winter, the heat produced by the PVT field is supplied primarily to the reversible heat pump for space heating, while the remaining part is used for DHW production; - in summer, the heat produced by the PVT field is supplied primarily to the hot side of the adsorption chiller for space cooling (with a surface aquifer for heat rejection to the ambient and the reversible heat pump in cooling mode as backup equipment), while the remaining part is used for DHW production.

3. Simulation model The system described in the previous section was dynamically simulated in TRNSYS. The simulation tool allows one to calculate energy and mass flows and temperature profiles for the components of the system for whatever period of the year. Both energetic and economic performance of the system under investigation can be analyzed by means of integrated data on whatever time period (hours, days, weeks, months or years). Moreover, the software includes a library of built-in components often based on experimentally validated data [78]. The entire simulation model of the proposed system includes several submodels, simulating all the components included in the system and listed in the previous section. It is worth noting, that all the submodels are linked to each other in order to perform the overall system simulation. In particular, the majority of the models (e.g. pumps, mixers, diverters, valves, controllers, gas burner, reversible water to water heat pump, adsorption chiller, building, etc.) were taken from the software library. On the contrary, specific models have been also developed for the management of the system (Control System) and for the calculations of primary energy savings and for the evaluation of the economic profitability of the system (Primary Energy Calculator and Economic Costs Calculator). The models of the components taken from TRNSYS library were

Table 4 Working modes of the system for different seasons. Season

Working mode

Winter

Summer

Space heating Space cooling

PVT þ HP e

DHW

PVT þ GB

e PVT þ ADS HP þ WE (as auxiliary device) PVT þ GB

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previously validated vs experimental data, as shown in TRNSYS reference [78]. Other components are based on a data lookup approach where the performance is obtained by manufacturer data. Therefore such components are intrinsically validated. Nevertheless, the validation of the overall performance of the system as a whole is not possible since, in authors' knowledge, a similar solar system has never been experimented. However, considering that all the sub-models of the components were previously validated vs experimental data and/or based on manufacturer's data, it may be concluded that the overall results are highly reliable due to robustness and consolidated use of the models of all the components of the system. In the followings, for sake of brevity, only a brief description is provided only for the PVT solar collector, the reversible heat pump and the adsorption chiller; a detailed description of the models used for the remaining components is available in the TRNSYS reference [78]: type 3b for the singe velocity pumps, type 110 for the variable speed pump P1, type 751 for GB, type 31 for the pipes, type 4d for TK1, type 340 for TK2, type 60f for TK3. 3.1. PVT collector The solar collector field consists of flat-plate photovoltaic thermal solar collectors, PVT. Sheet-and-tube collectors [12,13,51] are considered: they consist of a conventional solar thermal device including a PV film covering the absorber. PVT consists of a set of components: a glass cover, an absorber encapsulating the PV film, flow channels for the cooling fluid and thermal insulation. The PVT collector was simulated by adopting the component Type 50 included in the TRNSYS library. Here, PVT collectors are simulated by using a modified version of flat-plate thermal collectors in which a PV module has been added [78]. Moreover, constant values for the overall energy loss coefficient, the glass transmittance and the absorbance of the absorber [79] were assumed in the model. The model calculates PVT electrical efficiency as a function of the temperature of the PV cell average temperature, Tcell:

hPVT

h  i ¼ hPVT;ref pf 1  bPV Tcell  Tref

(1)

The temperature of the cell is calculated through an energy balance, accordingly to the modified Hottel eWillier e Bliss approach [78]. Thermal energy per unit area produced by the PVT is calculated as follows:

Ul* ¼ Ul  tg Itot hPVT ðTcell  Tamb Þ

FR ¼ m_ f cf

fp U * APVT l m_ f cf

1e Ul* APVT

  Q_ u ¼ Itot FRtg aPVT Ul* Tf ;in  Tamb

Table 5 PVT design parameters. Parameter

Symbol

Value

Unit

PVT gross area PVT fin efficiency factor Fluid specific heat Collector plate absorptance Collector loss coefficient Cover glass transmittance Temperature coefficient of PV efficiency Reference temperature for PV efficiency PVT packing factor: absorber area/PV area PV electrical efficiency at Tref

APVT fp cf

20 0.96 4.1877 0.92 16 0.89 0.0032 25 0.8 0.16

m2 kJ/kg K e kJ/h m2 K e e e  C e e

aPVT

Ul

tg bPV

Tref pf

hPVT,ref

the component included in the type 669 of the TRNSYS library. This component is based on user-supplied data files, which consists of manufacturer catalog data, including heating/chilling capacity and power consumption for the different operating conditions. The data files were generated on the basis of the load value, the source temperatures and liquid mass flow rates. Thus, the model adopted by the authors for the component is validated using manufacturer data. In particular, an AERMEC® WRL 41 reversible heat pump was selected, with heating and cooling capacities equal to 8 and 7 kW, respectively. The model is based on well-known thermodynamic equations. The heat pump COP in heating mode is given by:

COPHP;heat ¼

NCheat P_

(5)

heat

The thermal energy absorbed from the source liquid stream in heating mode is:

Q_ absor ¼ NCheat  P_ heat

(6)

The source and load outlet temperatures of the two liquid streams are calculated as follows:

Tsource;out ¼ Tsource;in 

Tload;out ¼ Tload;in 

Q_ absor m_ source cP source

NCheating _ mload cP load

(7)

(8)

The governing equations in the chilling mode are similar to those used for the heating mode, namely:

(2) COPHP;chill ¼

NCchill P_

(9)

chill

(3)

Q_ rejec ¼ NCchill þ P_ chill

(4)

Tsource;out ¼ Tsource;in þ

(10) Q_ rejec m_ source cP source

(11)

Finally, energy balances are used to calculate the outlet temperature of the fluid (Tf,out), the PV average temperature (Tcell) and the electric energy production. The PVT design parameters for the case study analyzed in the paper, taken from Ref. [21], are listed in Table 5.

Tload;out ¼ Tload;in þ

3.2. Reversible heat pump

3.3. Adsorption chiller

A reversible heat pump, HP, was used for space heating purposes and as space cooling auxiliary equipment. In particular, a singlestage water-to-water heat pump was considered [80]. In the model, the reversible heat pump was simulated by implementing

An adsorption chiller, ADS, was used for space cooling purposes in the summer season. In particular, a hot water powered adsorption chiller with solid desiccant zeolite matrix was used [80]. In the model, this component was simulated by the component

NCchill m_ load cP load

(12)

F. Calise et al. / Energy 95 (2016) 346e366

represented by type 909. As for the heat pump, the component is based on user-supplied data files, which consists of manufacturer catalog data, including cooling capacity and the coefficient of performance for the different operating conditions. The data files were generated on the basis of the hot, chilled and cooling side inlet temperatures and liquid mass flow rates. In particular, an InvenSor® LTC 10 vario adsorption chiller was selected, with a cooling nominal capacity equal to 10 kW. The thermodynamic model of the adsorption chiller is based on well-known equations. The thermal energy removed from the chilled water is computed as:

    Q_ chw ¼ MIN Q_ capacity ; m_ chw cP;chw Tchw;in  Tchw;set

(13)

where Tchw,set is the set point temperature for the outlet chilled water stream and is an input to the model. The adsorption chiller operates with the regeneration of a desiccant matrix, the energy required for this process is:

Q_ chw Q_ hw ¼ COPADS

(14)

In addition, adsorption chiller operation requires that thermal energy must be rejected by the cooling water stream. This energy flow is calculated as follows:

Q_ cw ¼ Q_ chw þ Q_ hw þ Q_ aux

COPADS ¼

353

Q_ chw

Q_ aux þ Q_ hw

(19)

3.4. Energy savings and economic analysis The energy analysis was performed evaluating the eventual savings in terms of primary energy achieved by the PS (proposed system), compared to a conventional system assumed as a reference (RS). In order to compare the systems on the same base of final primary energy demand, it was assumed that both PS and RS produce the same amount of space heating, cooling and domestic hot water. Moreover, according to the assumption that all the produced electric energy by the collectors is totally consumed by the user, RS electric energy produced by the PVT field must be entirely provided by the public grid. The reference system was assumed to include a natural gas boiler for DHW production (boiler system efficiency, hRS,DHW ¼ 0.85), a reversible air-to-air heat pump for space heating and cooling (coefficient of performance in heating and cooling mode COPRS,heat ¼ 3.0 and COPRS,chill ¼ 2.5, respectively). Moreover, in such reference system, the public grid is used for matching the electric demand of the building (average national electrical efficiency is hRS,el ¼ 0.46) [81]. Thus, the primary energy saving achieved by the PS is calculated as:

(15)

DPE ¼ PERS  PEPS PERS ¼ PEPS

QDHW Qheat Qchill þ þ hRS;DHW COPRS;heat hRS;el COPRS;chill hRS;el

(20)

EPVT;el þ EP5;el QDHW;GB Qchill;HP Qchill;ADS Qheat ¼ þ þ þ þ hRS;el hRS;DHW COPPS;HP;heat hRS;el COPPS;HP;chill hRS;el COPRS;chill hRS;el

where Q_ aux is the auxiliary energy required to run the pumps and controllers of the chiller. Once all the energy flow terms are defined (hot, chilled and cooling sides of the adsorption chiller), the outlet water streams temperatures can be calculated as:

Tchw;out ¼

Tcw;out

8 > T > < chw;set

   if Q_ capacity  m_ chw cP;chw Tchw;in  Tchw;set

Q_ capacity > > : Tchw;in  _ mchw cP;chw

Q_ cw ¼ Tcw;in þ m_ cw cP;cw

Thw;out ¼ Thw;in 

Note that, the electric energy consumption of P1 and P2 was neglected because these values are two orders of magnitude lower than the electric energy produced by the PVT. Conversely, the P5 electric energy consumption must be considered because

Q_ hw _ mhw cP;hw

   if Q_ capacity < m_ chw cP;chw Tchw;in  Tchw;set

(17)

(18)

Finally, the coefficient of performance is calculated as follows:

(16)

it is not negligible with respect to the PVT electric energy production. An economic analysis was also implemented in the simulation tool, calculating investment costs and economic savings. Cost functions for all the components of the system were introduced in order to calculate their capital costs. In order to estimate the components costs, the cost functions were derived from manufacturers’ data, on the basis of the selected capacities (Table 3).

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Note that in the previous equation, COP of HP is accurately calculated on the basis of the dynamic simulations as the ratio between the annual cooling/heating energy and the electrical one. Conversely, for the Reference System, RS, average yearly COP are considered. The economic performance of the proposed system is evaluated using the SPB (Simple Pay-Back period) index, defined as follows

SPB ¼ Fig. 4. 3D model of the one-floor building considered in the simulations.

PVT collectors and tank costs are evaluated as follows [44]:

CPVT ¼ 600APVT

(21)

Ctank ¼ 494:9 þ 808:0Vtank

(22)

The reversible heat pump cost was obtained by manufacturers' data, in the range of 5e35 kW; thus, it is evaluated as follows: 2 CHP ¼ 4:7108PHP þ 139:69PHP þ 3845:7

(24)

For the cost of Balance of Plant components required in the system (pipes, pumps, valves, controllers, etc.), a 20% of the overall cost estimated for PVT, tanks, HP and ADS was assumed. Thus, the total capital cost of the proposed system was estimated as:

CPS;tot ¼ 1:2ðCPVT þ Ctank þ CHP þ CADS Þ

(25)

For the RS, the following operating costs were assumed: a cost of 0.20 V/kWhel for the electric energy; a unit cost of 0.80 V/Sm3 (referred to a lower heating value of 9.59 kWh/Sm3) for the natural gas; differences in maintenance costs between PS and RS were neglected. Moreover, as mentioned before, it was assumed that the building consumes all the electric energy produced by the PVTsystem, also due to a net-metering contract with the public grid manager. The operating costs for the PS are not negligible and they are only due to the operation of the electrical devices, whereas the savings due to the following energy flows were considered in the economic analysis: i) electric energy produced by the PVT and self-consumed by the building; ii) electric energy savings due to a higher COP of the water-to-water heat pump with respect to the air-to-air one, for both heating and cooling modes; iii) thermal energy produced by the PVT and used for the DHW production. So, the annual savings of the PS with respect to the RS were calculated as:

(27)

where CRS,tot is the total cost of the reference system, assumed equal to the cost of the reversible heat pump incremented by 20%. This cost estimation was achieved analyzing manufacturers cost data for an HVAC (heating, ventilating and air conditioning) system based on reversible heat pumps of cooling and heating capacity similar to those involved in the proposed system. In order to consider a possible public funding strategy for the installation of a new HVAC system in a residential building a capital investment subsidy equal to the 50% of the total capital cost for both RS and PS was considered. This kind of subsidy is common in Italy for new HVAC systems. Under these conditions, the Simple Pay Back index is calculated as follows:

(23)

The specific cost per kW of cooling power of the adsorption chiller was estimated in 500 V/kWcool [82], thus the cost of the thermally driven chiller is:

CADS ¼ 500Qchill;ADS

CPS;tot  CRS;tot Jtot

SPBcis ¼

  0:5 CPS;tot  CRS;tot Jtot

(28)

4. Case study A base case study was developed for the system, using Meteonorm weather data of Naples, South of Italy. The building coupled with the SAHP system is a one-floor apartment located at the top floor of residential building. The building includes a flat roof and 4 zones (Fig. 4). The house floor and roof have an area of 100 m2 and the height is 2.70 m; the glazed area per each facade is 3.60 m2. Moreover, the house is divided into four rooms with independent fan coils that provide the required space heating or cooling energy. It was assumed that during the heating season space heating operates from 7 to 11 am and from 2 pm to 8 pm, while space cooling is activated from 8 am to 6 pm in the cooling season. In this paper, the Google SketchUP tool for designing 3D buildings was used in order to implement the case study building model. Thus, the first step of the work focused on the implementation of building model in the TRNSYS software. In particular, the model was imported in TRNSYS by using the TRNSYS3d plug-in Ref. [83]. Fig. 4 shows the building 3D model, whereas Table 6 presents transmittances of building envelope components. In particular, each component (wall, roof and floor) was modeled with a set of layers, not reported here for sake of brevity. Furthermore, the assumed building user DHW demand is reported in Fig. 5. The main design and operating parameters of the proposed system are reported in Table 3. In particular, for the PVT a tilt of

Jtot ¼ Cop;RS  Cop;PS Cop;RS ¼ Cop;PS

QDHW Qheat Qchill j þ j þ j LHVNG hRS;DHW NG COPRS;heat el COPRS;chill el 

 ¼  EPVT;el þ EP5;el jel þ

Qchill;HP Qchill;ADS QDHW;GB Qheat j þ j þ j þ j LHVNG hRS;DHW NG COPPS;HP;heat el COPPS;HP;chill el COPRS;chill el

(26)

F. Calise et al. / Energy 95 (2016) 346e366 Table 6 Transmittances of building envelope components. Element

Value [W/(m2 K)]

External wall External roof Floor Internal wall Window

0.830 0.816 1.048 2.856 2.830

30 and a field area of 20 m2 area were assumed. The selection of collector slope is consistent with the data reported in works [73,74,84]. In fact, using Meteonorm data for Naples [78], the optimal slope for an annual operation (without tracking system) is found at about 30 . For sake of brevity, only the main parameters used in the simulations are reported in Tables 3 and 5 due to the high number of components required to run the simulation. Moreover, a parametric design of the system components was performed, as a consequence some of the design parameters shown in the previous table are related to each other. In particular, the implemented algorithm selects automatically the capacities of the components on the basis of some main system parameters. For example, the PVT field area determines the P1 capacity, TK1 and TK2 volumes because the parameters determining the size of these components are referred to the unit of PVT field area (Table 3). It is worth noting that the capacities of the system components were selected in order to cover all the heating and cooling demand of the building and a part of the DHW demand all year long. For each simulation, the TRNSYS model provides several data, such as temperature, load profiles, and integrated values of the same variables of a fixed time basis (for example, the weekly, monthly or yearly energy demand). As mentioned above, the dynamic simulation tool developed in this work allows the use of whatever time basis (seconds, days, weeks, months or one year) to integrate the results. In particular, this feature is used for a better interpretation of the results because oscillations achieved during the dynamic operation are mitigated with the integration process. The system simulation was performed for one year (from 0 h to 8760 h) and the time step was 0.05 h. The relative tolerance convergence and integration values of 0.001 were adopted. In particular, specifying a relative convergence tolerance indicates that TRNSYS should not move on to the next time step until all connected outputs are changing by less than one-tenth of one percent of their absolute value and all integrated outputs are changing by one-tenth of one percent of their absolute value. In addition, a small time step (0.05 h) is required in order to

355

appropriately simulate controllers operation, since the time step is representative of controllers sampling time. In addition, small time steps are recommended in order to promote convergence in capacitive components (e.g. tanks) and to improve results accuracy. The dynamic behavior of the system is presented in terms of temperature and energy profiles for two days, representative of winter and summer system operation. The variation of the most important energy parameters during the year is presented on a weekly base. The yearly results are presented with an annual integration for the base case. The yearly results are also presented for additional two locations, in particular the system simulation was performed for the weather data of Palermo and Milan in order to investigate the system performance in different climatic conditions. Finally, a parametric sensitivity analysis and a thermo-economic optimization were performed. In particular, the first analysis was performed in order to study the dependence of the performance and economical profitability of the PVT system as a function of main system parameters. 5. Results 5.1. Winter day The selected representative winter day of the daily results provided by the simulation in the heating season is January 27th (from 600 h to 624 h). Fig. 6 shows the main temperatures in the system for the selected day, where the operation of the system is outlined. In particular, it is clearly shown the operation of the controllers managing the pump P1, mixers (M1 and M2) and diverters (D1 and D2) installed in order to manage the SCF flow to TK1 and to TK2 internal heat exchanger. It is worth noting that none operation occurs in the system until 7 am, because there is no demand for space heating or DHW. During the night hours (from 0.00 am to 7.00 am) there is no solar thermal energy produced by the PVT collectors (QPVT) and both tanks (TK1 and TK2) are not heated. Therefore, a decrease of the temperature inside such tanks (TTK1,top and TTK2,top) is reported, due to their thermal losses. At 7 am, the sensor of the radiation detects irradiance higher than 10 W/m2; as a consequence, P1 is activated and the SCF starts to circulate within the solar loop. Thus, PVT outlet flow temperature (TPVT,out) rises due to the solar thermal gain. At the same time, TK1 top temperature decreases below 20  C. Thus, according to the system control strategy, during the first hours of the morning SCF is pumped from the bottom of TK1 to the

Fig. 5. Daily DHW demand for the simulated building.

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F. Calise et al. / Energy 95 (2016) 346e366

Fig. 6. Temperatures, winter day 27th January (600e624 h).

PVT field by means of P1. In this phase, inlet PVT temperature is equal to the bottom TK1 temperature (TTK1,bottom), thus the solar thermal energy produced by the PVT field is supplied to TK1. This operation occurs until its top temperature reaches the fixed setpoint temperature (25  C). Moreover, during this operation, PVT set point temperature is set to 30  C by the control system. At the same time, TK2 top temperature continues to decrease, due to DHW demand and thermal losses. As a consequence, TK2 temperature is not high enough to produce DHW using only solar energy, thus the auxiliary heater is activated in order to provide DHW at 45  C. As previously mentioned, the building space heating system is activated at 7:00 am. At the same time, TK3 temperature (TTK3,top) is about 45  C, due to the thermal losses occurred in the night. Thus, at 7:00 am the reversible heat pump is turned on and heats the TK3 up to the set point temperature (47  C). Therefore, TK1 top temperature decreases due to the HF supply to the source side of HP. TK1 reaches the set point temperature approximatively just before 12:00 am and during the previous part of day TK1 top temperature remains below 25  C and no solar thermal energy is supplied to TK2. Therefore, the heat produced by the PVT field is entirely delivered to TK1, until its top temperature reaches the fixed set-point temperature (25  C). When TK1 is thermally loaded, D2 deviates the SCF to the TK2 internal heat exchanger, for DHW production. At the same time, the control system sets the PVT set point temperature to 55  C. The DHW tank is supplied with SCF approximately until just after 4.00 pm; during this operation the TK3 tank set point

temperature is not reached due to the low PVT outlet temperature and to the DHW demand. Therefore, the solar thermal energy supplied to TK2 is not enough to produce DHW exclusively by solar energy, thus the auxiliary heater is activated again in order to provide DHW at 45  C. During the morning, TK3 top temperature decreases due to tank thermal losses and FC (fan coil) system heat demand. Therefore, just after 10:00 am, the temperature at the top of TK3 drops below the lower deadband (45  C); so, HP is again activated in order to rise the temperature to the set point temperature (47  C). Such decrease of TK3 temperature is due to the high space heating demand of the building in the morning hours. In the same period, TK1 is still supplied by solar energy, however, its top temperature drops to approximatively 19.5  C due to the thermal energy supply to the HP source side. Furthermore, just before 7:00 p.m. the HP is again activated in order to rise the TK3 top temperature to the 47  C set point. Moreover, from 4:00 pm to 9.00 pm the TK2 top temperature decreases mainly due to the DHW demand of building users. It is worth noting that, the PVT outlet temperature do not exceeds 70  C, thus the operation of HE is not required. The system energy flows can be also used to explain the dynamic behavior of the proposed system. In particular, energy flow rate trends are shown in Fig. 7. The trend of the solar radiation (IPVT) significantly affects the profiles of the heat and power (QPVT and PPVT) produced by the PVT field, obviously the power produced by the system is directly connected to the solar radiation. Moreover, the thermal powers supplied to TK1 and TK2 (QTK1 and QTK2) are also shown. Note that, during the first part of the day, thermal energy produced by the PVT collectors is completely

Fig. 7. Thermal and electric flow rates, winter day 27th January (600e624 h).

F. Calise et al. / Energy 95 (2016) 346e366

supplied to TK1. Then, during the remaining hours of the day, solar energy supplies TK2. In the same figure, the activation of HP is also clearly represented, showing that, for the selected day, space heating demand occurs in the morning and late evening hours. It is also worth noting that the thermal power supplied by HP (QHP,heat) during the TK3 heating decreases due to the TK1 top temperature decrease.

5.2. Summer day As an example of the daily results provided by the simulation in the cooling season, the representative summer day of June 19th (from 4032 h to 4056 h) was selected. Fig. 8 shows the main temperature profiles for the selected day. Here, it is clearly shown that the solar loop is activated at 5:00 am. Obviously, the solar loop in summer is activated earlier with respect to winter case due to the higher sunlight hours. In the early morning hours, TK1 top temperature is lower than 60  C. Thus, SCF is pumped from P1 to TK1 (across D1, D2, M1 and M2). In the same period, solar radiation is not high enough to rise TK1 top temperature to the desired set point (65  C). Moreover, TK1 top temperature decreases due to thermal losses. The space cooling system is activated at 8:00 am, as a consequence the inertial tank temperature (TTK3,top) increases until it reaches the value of 14  C, approximately at 8:15 am. Then, the adsorption chiller turns on and supplies chilled water CHW (chilled water) to cool down the TK3 to the desired temperature (12  C). In this operation, CW is supplied the cooling side of ADS by means of P5 and the cooling circuit mixers and diverters. At the same time, TK1 top temperature decreases to 55  C, minimum allowable temperature for ADS operation. Thus, the control system turns ADS off and as a consequence there is no thermal power supply from TK1. After the ADS deactivation, TK3 top temperature increases due to the FC system cooling demand. When this temperature reaches 16  C, HP is activated in order to match the cooling demand. In this operation, D4 and M5 split CW to the condenser side of HP in order to reject the produced heat. Therefore, the inertial tank temperature is lowered to the set point value of 12  C. This takes about 45 min. During the morning hours, TK2 top temperature is lower than 45  C due to DHW demand of the previous day, and decreases due to thermal losses. Thus, GB activation is mandatory required in order to match the DHW demand. After the ADS deactivation, approximatively at 8:40 am, TK1 top temperature starts to increase due to the solar thermal energy supplied by the PVT collectors. Therefore, approximately at

357

11:15 am the temperature at the top of TK1 increases to the setpoint value, 65  C. Then, TK1 is thermally loaded and the control system switches D1 and D2 flow to the TK2 internal heat exchanger in order to produce DHW. In the same period, the inertial tank temperature increases due to the building cooling demand. Thus, just before 12:00 am the activation of ADS is performed and the TK3 temperature is lowered from 14  C to 12  C. This implies a decrease of TK1 temperature to 60  C approximatively at 12:20 am. As a consequence, the control system switches SCF flow to the source side of TK1 in order to heat it again to 65  C. Therefore, from 12:20 am to about 1:10 pm the solar thermal energy is supplied to TK1 instead of TK2. Only when TK1 temperature reaches again 65  C, SCF can be supplied to TK2. Note that, the described operation of TK1 and TK2 heating is repeated again during the day, in particular in the early evening hours. It is worth noting that, during the considered day, TK2 top temperature does not reach the set point value of 50  C, thus GB activation is activated in order to match DHW demand. In Fig. 8 it is clearly shown as the temperature at the top of TK3 slightly increases after ADS or HP deactivation due to the FC system operation. Moreover, during the space cooling period (8 ame6 pm), TK3 top temperature oscillates mainly between 12 and 14  C, in order to properly provide the required inlet temperature to the FC system. Only when ADS cannot be supplied by solar thermal energy (TK1 temperature <60  C), TK3 temperature rises to 16  C. In this case, to prevent a further temperature increase the reversible heat pump is activated in cooling mode. As for the winter day, the dynamic behavior of the system can be also and better interpreted observing the heat flows shown in Fig. 9. Here, the system operation discussed above is clearly represented. Note that, that the thermal power produced by the PVT collectors increases when the TK2 internal heat exchanger is supplied. This occurs because the outlet TK2 internal heat exchanger is lower with respect to the TK1 bottom one, as a consequence the PVT heat exchange is higher. Furthermore, the chilling power supplied by ADS (QADS,chill) is more sensitive to the temperature variation of the thermal source with respect to the HP one (QHP,chill). Moreover, during the system operation no heat is dissipated by HE because the storage capacity of the proposed system is not fulfilled, thus the outlet PVT temperature is always lower than 70  C (Fig. 8).

5.3. Weekly analysis The proposed system was also analyzed with the integration of the results on a weekly basis in order to show the variation of

Fig. 8. Temperatures, summer day 19th June (4032e4056 h).

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F. Calise et al. / Energy 95 (2016) 346e366

Fig. 9. Thermal and electric flow rates, summer day 19th June (4032e4056 h).

energy flows and the system performance during all year. In particular, this analysis was performed for a better understanding of the results due to the mitigation of the fluctuations of the energy flow trends shown in the previous subsection of the paper. In Fig. 10 thermal and electrical energy flows (QPVT and PPVT) related to the PVT field for each week are reported. Here it is clearly shown that system performance is significantly affected by the different availability of solar radiation in summer and winter. Solar thermal energy supplied by both tanks (QTK1 and QTK2) and the thermal energy dissipated (QHE) are also plotted in Fig. 10. It is clearly shown that the last one is negligible in the winter period with respect to the solar energy supplied by both tanks, because PVT outlet temperature is often below 70  C and the solar energy is completely supplied to TK1 and TK2. Only in intermediate seasons (when space heating/cooling demand is very scarce or null), HE dissipates the solar thermal energy. Chilled energy supplied in summer by the adsorption chiller (QADS,chill) is higher than the heat pump one (QHP,chill), this is due to the strategy adopted for HP activation as backup system in order to satisfy the cooling demand. Space cooling energy is demanded from mid-June to midSeptember and it is higher than the heating one. The utilization of the solar thermal energy among the thermal devices, TK1, TK2 and HE, is better shown in Fig. 11. Here, the solar utilization factor (f) for the three thermal devices is reported. The solar utilization factors is defined as the ratio between solar thermal energy supplied to each component (TK1, TK2 and HE,

respectively) and the total solar thermal energy provided by the PVT. In winter and summer, solar thermal energy produced by the PVT collectors is mainly supplied to the solar tank TK1, in order to satisfy the thermal energy demand of the reversible heat pump and the adsorption chiller, in winter and summer respectively. Note that in week 2 the thermal energy produced by the solar field is completely supplied to TK1, because the solar radiation is low and the heat demand is high. Moreover, during the summer period the heat dissipated by HE is null, due to the high demand of thermal energy of the adsorption chiller and DHW tank. In particular, the maximum DHW production is achieved during the intermediate seasons due to high thermal energy produced by the PVT field and to low system cooling and heating demands. The operation of the chilling equipment is reported in Fig. 12 as utilization factors. It is defined as the ratio between thermal energy supplied to the user in terms of chilling water by each component (HP and ADS) and the total one. In the first week of cooling season operation the thermal energy demand of the fan coil system is almost totally supplied by the reversible heat pump. In fact, a small amount of solar thermal heat is provided to TK1 in this week, thus a negligible thermal energy is supplied to ADS. In all the remaining weeks of cooling season the utilization factor of the ADS is higher than the HP one. Note that the utilization of ADS is higher in the firsts and lasts weeks of cooling season, instead of the central ones. In the central weeks, solar energy availability is high (Fig. 10), however the adsorption chiller contribution decreases due to high

Fig. 10. Thermal and electrical energies, weekly analysis between 1st and 52 nd week of the year.

F. Calise et al. / Energy 95 (2016) 346e366

359

Fig. 11. Solar energy utilization factor between 1st and 52 nd week of the year.

Fig. 12. Utilization factor of ADS and HP between 22nd and 44th week of the year (cooling season).

space cooling demand. As a consequence, the reversible heat pump is often activated as a chiller in order to satisfy the user demand. In the last weeks of the cooling season, the cooling demand is entirely supplied by ADS because of the middle season climatic conditions that produce a low space cooling demand. 5.4. Yearly results The results of the annual simulations (from 0 h to 8760 h) for the base case are summarized in Table 7, where both thermal and electrical energies are shown. Moreover, in such table annual results for the weather conditions of Palermo and Milan are also presented. According to the system behavior discussed in previous sections, the energy supplied to the building for space cooling (Qchill,req) is higher than that needed for space heating (Qheat,req). This occurs for Naples and Palermo because the weather conditions for these two locations are similar. Conversely, for Milan the heating demand is higher than the space cooling one due to colder climatic conditions compared to Naples and Palermo. Moreover, Table 7 shows that solar thermal energy utilized by the system is lower than the one demanded by the user in term of space heating/ cooling and DHW demands. In fact, only a part of the cooling ad DHW demand (QDHW) is supplied by solar thermal energy (QDHW,sol). This condition enhances the full utilization of the solar energy. As expected, PVT thermal energy (QPVT) produced is higher in Palermo and Naples with respect to Milan due to the different

Table 7 Annual results: thermal and electrical energies. Parameter

IPVT QPVT PPVT QHE QTK1 QDHW,sol QDHW QHP,heat PHP,heat QHP,chill PHP,chill PHP QADS,hot QADS,chill QTK3,heat QTK3,chill Qheat,req Qchill,req PE

Value (kWh/year) Naples

Palermo

Milan

3.46Eþ04 1.34Eþ04 3.65Eþ03 1.18Eþ03 6.24Eþ03 4.89Eþ03 8.92Eþ03 1.94Eþ03 4.61Eþ02 8.03Eþ02 1.37Eþ02 5.99Eþ02 3.85Eþ03 2.10Eþ03 1.68Eþ03 2.78Eþ03 1.38Eþ03 2.71Eþ03 1.45Eþ04

3.75Eþ04 1.44Eþ04 3.90Eþ03 1.74Eþ03 6.45Eþ03 5.15Eþ03 8.92Eþ03 5.60Eþ02 1.24Eþ02 1.12Eþ03 1.92Eþ02 3.16Eþ02 5.14Eþ03 2.81Eþ03 2.88Eþ02 3.81Eþ03 1.85Eþ02 3.72Eþ03 1.52Eþ04

2.71Eþ04 9.87Eþ03 2.92Eþ03 4.89Eþ02 5.68Eþ03 2.96Eþ03 8.92Eþ03 4.84Eþ03 1.70Eþ03 3.51Eþ02 6.01Eþ01 1.76Eþ03 2.00Eþ03 1.09Eþ03 4.57Eþ03 1.32Eþ03 4.12Eþ03 1.28Eþ03 9.84Eþ03

availability of the solar total radiation. Solar thermal energy supplied by TK1 (QTK1) do not show significant variations among the three locations, ranging around 5e6 MWh. Moreover, HP electric consumption (PHP) is mainly due to space heating operation because in summer the heat pump is only activated as backup

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Table 8 Comparative analysis of efficiency, economic parameters and annual energy results for the cities of Naples, Palermo and Milan. Parameter

Naples

Palermo

Milan

Unit

hth,PVT hel,PVT

0.387 0.105 0.466 0.365 0.088 4.214 5.841 0.546 0.724 0.276 1.41Eþ03 2.71Eþ04 16.26 8.13

0.383 0.104 0.449 0.359 0.121 4.502 5.841 0.547 0.715 0.285 1.52Eþ03 2.71Eþ04 15.06 7.53

0.364 0.107 0.576 0.300 0.050 2.838 5.841 0.544 0.756 0.244 9.39Eþ02 2.71Eþ04 24.43 12.22

e e e e e e e e e e V/year V years years

fsol,TK1 fsol,TK2 fsol,HE COPHP,heat COPHP,chill COPADS fchill,ADS fchill,HP Jtot CPS,tot SPB SPBca

device. Furthermore, yearly demand of DHW in the building is covered by more than 50% by the solar collectors in Naples and Palermo, conversely in Milan the solar contribution to the DHW demand is only about 30%. The high utilization factor of the available solar energy all over the year leads to a significant primary energy saving (PE), equal to 14.3 MWh/year in the base case study. The main performance and economic parameters of the proposed system are summarized in Table 8. PVT thermal efficiency (hth,PVT), calculated with respect to the total solar radiation, is between 37 and 39% for the three localities. The efficiency is higher in Naples and Palermo because the average ambient temperature is higher compared to Milan, thus the thermal losses of the PVT collectors are lower. The low set point temperatures of the PVT collectors (30  C, 55  C and 70  C, depending on the operating conditions) leads to an electrical efficiency more than 10%. The electric efficiency is higher in Milan with respect to Naples and Palermo due to a lower average PVT collector temperature related to colder climatic conditions. This trend is opposite with respect to the one shown for the thermal efficiency. Furthermore, the total energy efficiency of the PVT system oscillates from 47 to 49% for the three cities used in the simulation. Moreover, Table 8 also reports TK1, TK2 and HE solar utilization factors (fTK1, fTK2 and fHE). Solar energy produced in the year is mainly supplied to TK1, in order to supply HP and ADS in winter and summer, respectively. For the base case, the energy dissipated by HE is lower than 9%. This percentage is higher in Palermo due to higher irradiation in summer period and lower in Milan due to the lower availability of solar energy source. Considering the thermal losses and the heat dissipated by HE, more than 80% of the solar thermal energy produced by PVT collectors is supplied by TK1 and TK2, for all the localities. Furthermore, the space cooling demand is mainly supplied by the adsorption chiller. In fact, the utilization factor of ADS is higher than 70% in all the locations considered. The yearly mean COP of the water-to-water heat pump for space heating (COPHP,heat) is above 4, for Naples and Palermo. This is due to the TK1 top temperature, that is relatively high (20e25  C) when the HP source side is supplied in the heating season. It is worth noting that the COP of the reversible heat pump is relatively low (2.835) in the operational conditions of Milan, because in winter solar irradiation is scarce in order to maintain TK1 top temperature within the fixed optimal deadbands. In the summer period, the COP of the same HP, operating as a chiller, (COPHP,chill) is above 5, due to operation of the aquifer well that supplies cooling water to the source side of HP. Moreover, the COP of the adsorption chiller (COPADS) is near the nominal one (0.60), for all the locations. This occurs because the hot side inlet temperature of ADS in summer is maintained mostly between 60 and 65  C, for a proper performance of the thermally powered chiller.

In Table 8, an investment cost of the proposed system of 27.1 kV is reported. The economic analysis returns a SPB period of about 16 years, for the system localization in Naples. This value was obtained with a saving of about 1.41 kV/year, an initial cost for the reference system of 4.15 kV and without any public funding. In order to consider a capital investment subsidy (similar to that currently adopted in Italy for renewable energy and energy efficiency systems, also including solar energy and heat pumps) a 50% discount value was considered for both PS and RS. In this case, the SPB decreases and it halves to about 8 years. Obviously, SPB period is lower in Palermo due to the higher savings achievable with a higher radiation with respect to Naples. However, SPB without incentives remains high (more than 15 years). Finally, the low savings achieved by the system in Milan implies a very high SPB period, even when a capital investment subsidy of 50% is considered. Thus, this kind of system is not economically feasible in climatic conditions similar to those of the city of Milan, where the heat demand is high and the solar irradiation is low. In fact, for such city the simple geothermal heat pump, coupled with the aquifer, represents the most favorable option. 5.5. Parametric analysis A parametric analysis was also developed, in order to complete the study of the proposed system. In particular, the sensitivity of the performance on the most significant design variables and/or boundary conditions was analyzed, when all the remaining parameters remain fixed. The following parameters were chosen on order to perform the parametric analysis: PVT field area (APVT), TK1 tank specific volume per unit of PVT area (Vtank,PVT), flow rate of P1 pump per unit of PVT, area (qP1), set point temperature of PVT in winter (Tset,PVT,wint), set point temperature of PVT in summer (Tset,PVT,summ), TK1 set point temperature in winter (Tset,TK1,wint), TK1 set point temperature in summer (Tset,TK1,summ), TK3 set point temperature in winter (Tset,TK3,wint), TK3 set point temperature in summer (Tset,TK3,summ). The most important design parameter of a solar system is the field area, such parameter is related to the energy production. This is outlined in Fig. 13. Here, the thermal and electric energy flows produced by the solar field and the primary energy savings are directly proportional to the solar collector area. The graph also shows the solar thermal energy supplied by TK1, TK2 and dissipated by HE. The amount of solar heat delivered to the tank TK1 and TK2 is almost constant for PVT area higher than 40 m2. Only a slight increase is detected. Thus, for PVT area higher than 40 m2 a maximum utilization of solar thermal heat is achieved by the system. Furthermore, only a slight increase of the thermal energy supplied by TK1 to HP is reported for PVT area from 10 to 30 m2, over this last value the energy supplied is almost constant. This occurs because the energy required on the source side of HP in winter is already satisfied even in case of low PVT area. Moreover, the thermal energy supplied to the hot side of ADS (QADS,hot) increases significantly only from 10 to 40 m2, for higher values of PVT area the energy supplied is almost constant. It is worth noting that the energy dissipated by HE increases proportionally for PVT area higher than 40 m2, assuming the same trend of the thermal energy produced by PVT collectors. Therefore, over this threshold, solar energy supplied by TK1 and TK2 is constant and the thermal energy produced by the system is mainly dissipated. Thus, there is no energetic convenience to increase the PVT area to values higher than 40 m2. The variations of the main system energetic and economic performance parameters as a function of the PVT area are shown in Fig. 14. Thermal efficiency is sensitive to the PVT area: higher thermal efficiency for lower PVT field sizes would be achieved, due to the lower mean temperature of the SCF entering the collectors.

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Fig. 13. Parametric analysis: energies, solar collector area.

Fig. 14. Parametric analysis: efficiency and economic parameters, solar collector area.

On the contrary, the electric efficiency is scarcely sensible to the PVT field size: only a slight decrease can be observed, due to the increase of the mean PVT inlet temperature with a PVT area increment. In Fig. 14, the previously defined solar utilization factors (referred to TK1, TK2 and HE), are also shown. An increase of the PVT area implies a lower solar energy utilization factor for TK1. This is due to a constant space heating and cooling demand and to a consequent constant solar heat supplied by TK1 for higher PVT area. For TK2 tank, such factor rapidly moving from 10 m2 to 30 m2; then, a decrease rate appears. This one for a PVT area higher than 40 m2

becomes almost constant, since DHW is almost completely supplied by the solar thermal energy. Moreover, the increase of thermal energy produced by the PVT collectors and not utilized by the building (see Fig. 13), determines an increase of the solar utilization factor related to the heat dissipation by HE. Furthermore, the COP related to HP heating mode increases for larger PVT solar fields, due to TK1 higher average temperature. In particular, a significant increase from 10 to 40 m2 is reported. Obviously, COP of HP in chilling mode is not reported, because it is independent of the field size, only depending on the operation of the aquifer well. Moreover, the COP of the adsorption chiller is

Fig. 15. Parametric analysis: energies, specific TK1 volume.

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almost constant at different PVT area values, due to the constant operation conditions of the thermally driven chiller (optimal inlet hot side temperature) ensured by the control system. Fig. 14 shows the Simple Pay-back period, SPB. In particular, the minimum SPB is obtained for a PVT area of 30 m2. However, in that case SPB is too high (about 15 years) and the system is not economically profitable. In case of a capital investment subsidy of 50%, this value is halved. Fig. 15 shows the parametric analysis as a function of the ratio between TK1 volume and PVT area. As a consequence, an increase of this parameter determines a corresponding increase of the solar thermal storage capacity of TK1. Results show that this parameter scarcely affects the energy flows involved in the system. It is worth noting that the thermal energy produced by the PVT field decreases in case of higher values of this parameter, due to a higher PVT inlet temperature. In particular, this is due to a larger energy supply from SCF to PVT field by TK1; TK1 mean bottom temperature is higher compared to the TK2 internal heat exchanger outlet one. In fact, the produced PVT electric energy also decreases with higher PVT area. Moreover, the solar energy supplied by TK2 decreases due to the higher thermal energy supplied to TK1 and to the lower PVT thermal output, both achieved for TK1 volume increment. Results show that both electrical and thermal efficiencies of the PVT are scarcely sensitive to the increase of TK1 volumes (Fig. 16). According to the energy flows reported in Fig. 15, TK1 solar utilization factor increases. Thus, the utilization factor of ADS increases too. As a consequence, HP utilization for space cooling decreases. Furthermore, as shown in Fig. 16, the COP of the heat pump heating mode is almost constant at TK1 specific volume area higher than 50 l/m2, only a slightly decrease is detected for lower TK1 specific volumes. This is due to the lower storage capacity of TK1 that causes a lower temperature of the tank in case of low radiation. Furthermore, SPB period increases linearly with the increase of TK1 volume, due to the increase of the capital cost along with a decrease of the produced thermal and electric energy. In addition, the performance of the system as a function of the specific flow rate of P1 pump was analyzed. The results show that the system performance is scarcely affected by the variation of this parameter over 25 kg/(h m2). Moreover, the heat dissipated by HE increases for low specific flow rates of P1 (lover than 20 kg/(h m2)), due to the recurrent high PVT peak outlet temperature. The parametric analysis plots as a function of the specific flow rate of P1 are omitted for sake of brevity. Moreover, Table 9 shows that system performance is almost independent with respect to the variation of set point temperatures of PVT, TK1 and TK3, in winter and summer. It is only shown that COP of HP is higher for higher set point temperatures of TK1 in

winter. Conversely, COP of the adsorption chiller decreases for higher TK1 set point temperatures in summer, according to the data sheet of the producer. Moreover, SPB period is sensitive to the summer TK3 set point temperature decrease; this is due to the lower solar heat supplied to TK2. As a consequence, lower savings by the solar production of DHW are achieved for lower TK3 set point temperature. 5.6. Thermoeconomic optimization In order to complete the study, a thermo-economic optimization of the case study for the city of Naples is also performed. The analysis was implemented using TRNOPT plug-in tool included in TRNSYS package, where complex mathematical algorithms are adopted in order to perform the optimization. TRNOPT tool is designed to link the optimization algorithm and the dynamic simulation. In particular, this tool uses the algorithms included in the GENOPT package developed by Lawrence Berkeley National Laboratory [85]. For this analysis, the Generalized Search Method was used for the optimization process. The performed method avoids the calculation of partial derivatives in the optimal value calculation procedure. In particular, the HookeeJeeves [86] modified algorithm was performed. The structure of this algorithm avoids the achievement of local minimum points and takes into account the TRNSYS solving technique for the approximation of the objective function. This robust optimization method allows one to perform realistic computational times and to obtain the optimum value in a relatively low number of simulations. The optimization was performed considering only the main design variables, namely: PVT field area, TK1 specific volume, TK3 volume and TK1 winter/summer set point temperatures. Moreover, the SPB was selected as optimization objective function. Fig. 17 shows SPB and independent variables as a function of optimization iteration. The optimization process converged in about 100 iterations. Results of the thermo-economic optimization are consistent with the ones obtained by the sensitivity analysis. The optimum SPB value of 14.38 years is obtained for a PVT area of 24.25 m2, TK1 specific volume of 50.00 l/m2, TK1 winter and summer set-point temperature of 20.00 and 65.00  C, respectively, and a TK3 volume of 1.00 m3. Note that in Fig. 17, the TK3 volume optimization variable is not reported because during the optimization process the variation of such parameter is negligible. Moreover, it is worth noting that, the optimization algorithm was performed limiting the lower value of TK1 specific volume to 50.00 l/m2 in order to limit the number of on/off events of the solar loop.

Fig. 16. Parametric analysis: efficiency and economic parameters, specific TK1 volume.

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Table 9 Parametric analysis: efficiency and economic parameters vs set point temperatures. Parameter

Tset,PVT,wint

Tset,PVT,summ

Tset,TK1,wint

Tset,TK1,summ

Tset,TK3,wint

Tset,TK3,summ

30 32 34 36 38 40 70 72 74 76 78 80 25 26 27 28 29 30 65 66 67 68 69 70 45 46 47 48 49 50 10 11 12 13 14 15

hth,PVT

hel,PVT

fTK1

fTK2

fHE1

fADS

fHP

COPHP,heat

COPADS

SPB

SPBcis

0.387 0.387 0.387 0.387 0.387 0.387 0.387 0.384 0.384 0.382 0.379 0.379 0.387 0.387 0.387 0.387 0.387 0.387 0.387 0.384 0.382 0.379 0.376 0.373 0.382 0.382 0.382 0.382 0.382 0.382 0.379 0.382 0.384 0.384 0.382 0.382

0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105 0.105

0.466 0.465 0.464 0.464 0.464 0.464 0.466 0.468 0.469 0.471 0.477 0.473 0.466 0.466 0.466 0.469 0.471 0.472 0.466 0.484 0.497 0.506 0.512 0.516 0.471 0.471 0.473 0.471 0.474 0.473 0.533 0.513 0.498 0.482 0.473 0.461

0.365 0.366 0.367 0.366 0.367 0.368 0.365 0.367 0.366 0.371 0.372 0.376 0.365 0.363 0.361 0.357 0.354 0.351 0.365 0.344 0.337 0.326 0.322 0.322 0.372 0.372 0.369 0.372 0.368 0.371 0.312 0.333 0.341 0.356 0.369 0.381

0.088 0.088 0.088 0.088 0.088 0.088 0.088 0.084 0.080 0.075 0.071 0.067 0.088 0.087 0.087 0.087 0.087 0.086 0.088 0.086 0.084 0.080 0.079 0.076 0.077 0.077 0.077 0.077 0.077 0.077 0.074 0.075 0.077 0.077 0.077 0.078

0.723 0.720 0.720 0.720 0.720 0.720 0.723 0.723 0.726 0.720 0.725 0.719 0.723 0.721 0.716 0.722 0.722 0.723 0.723 0.753 0.768 0.776 0.779 0.781 0.721 0.722 0.723 0.718 0.725 0.722 0.788 0.781 0.769 0.738 0.723 0.716

0.277 0.280 0.280 0.280 0.280 0.280 0.277 0.277 0.274 0.280 0.275 0.281 0.277 0.279 0.284 0.278 0.278 0.277 0.277 0.247 0.232 0.224 0.221 0.219 0.279 0.278 0.277 0.282 0.275 0.278 0.212 0.219 0.231 0.262 0.277 0.284

4.21 4.21 4.20 4.20 4.20 4.19 4.21 4.21 4.21 4.21 4.21 4.21 4.21 4.26 4.31 4.35 4.40 4.44 4.21 4.21 4.22 4.21 4.21 4.22 4.37 4.30 4.21 4.14 4.05 3.97 4.22 4.21 4.21 4.21 4.21 4.21

0.545 0.544 0.544 0.544 0.544 0.544 0.545 0.545 0.547 0.546 0.548 0.549 0.545 0.545 0.545 0.545 0.545 0.545 0.545 0.542 0.537 0.533 0.534 0.535 0.546 0.548 0.545 0.547 0.547 0.548 0.525 0.527 0.532 0.537 0.545 0.561

16.3 16.2 16.2 16.2 16.2 16.2 16.3 16.3 16.3 16.3 16.3 16.3 16.3 16.3 16.3 16.3 16.4 16.4 16.3 16.6 16.7 17.0 17.1 17.1 16.2 16.2 16.3 16.3 16.3 16.3 17.8 16.9 16.7 16.4 16.3 16.1

8.13 8.12 8.12 8.12 8.12 8.11 8.13 8.14 8.15 8.14 8.15 8.13 8.13 8.14 8.15 8.17 8.19 8.21 8.13 8.30 8.37 8.48 8.54 8.56 8.11 8.12 8.15 8.14 8.17 8.16 8.88 8.45 8.33 8.22 8.15 8.06

Fig. 17. Thermo-economic optimization: SPB objective function and optimization variables.

6. Conclusion The dynamic simulation model of a trigeneration system including a solar-assisted heat pump coupled with an adsorption chiller is presented. In particular, the system under investigation is based on a PVT collector field coupled with a reversible heat pump and a zeolite adsorption chiller. The electric energy produced by the PVT field is totally consumed by the system users, while the thermal energy is used primarily to supply the solar

assisted heat pump, in winter, and the adsorption chiller, in summer, and secondarily to produce DHW. A base case study was developed and discussed, referred to a building located in Naples, in the South of Italy. The results of the dynamic simulation showed that: - PVT thermal energy is fluctuating due to the external radiation and temperature, thus an auxiliary system is always mandatory in order to satisfy the DHW and space cooling demand;

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- in summer period, the combination of PVT and an adsorption chiller allows one to maximize the utilization of the solar energy because no thermal energy produced by the PVT is dissipated; - the utilization of the produced solar thermal energy is about 83%; - the space heating demand is entirely supplied with the solar assisted heat pump, and the cooling demand is supplied by higher than 70% by the adsorption chiller for the base case study; - the system energetic and economic performance is better in localities with high irradiance availability, as Naples and Palermo; - the system may is not profitable without public incentive (SPB higher than 16 years), became profitable with capital investment subsidy of 50%. - for the base case, the system optimal performance in terms of economic profitability is achieved with a collector area of 24.25 m2; - the system performance is scarcely dependent on the set point temperatures of the PVT collectors, solar storage tank and inertial tank of the hydronic system. Nomenclature A c C NC E f I j J P PE P Q SPB T U

area, m2 specific heat, J/(kg K) cost, V nominal capacity, kW energy, kWh utilization factor, solar utilization factor, e solar irradiance, kW/m2 specific cost-price, V/kWh savings, V/year pressure, kPa primary energy, kWh electric power, kW thermal power, kW simple pay back, years temperature,  C or K heat transfer coefficient, kW/(m2  C)

Greek symbols a absorbance coefficient, adim t transmissivity coefficient, adim Subscripts and superscripts amb ambient air absor absorbed ADS referred to adsorption chiller aux auxiliaries capacity capacity cell referred to photovoltaic cell CHW chilled water cis capital investment subsidy cw cooling water cool cooling DHW domestic hot water el electric f fluid g glass heat heating hw hot water GB referred to GB gas burner HP referred to heat pump in input

load NG op out PS pump PV PVT ref rej RS source tank th TK1 TK2 tot

referred to the load side of heat pump natural gas operational output proposed system referred to pump photovoltaic PVT solar collectors reference condition rejected reference system referred to the source side of heat pump tank thermal referred to TK1 tank referred to TK2 tank total

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