2.2. Layout of the System and Operation Strategy
The investigated system is based on the integration of flat-plate collectors, concentrator, and adsorption chiller in a traditional DHW, space heating, and cooling system adopting a natural gas boiler and a reversible heat pump system for space conditioning. Solar thermal collectors and the concentrator are used for providing DHW all year long and space heating in the heating season by means of a reversible heat pump with water-to-water loops, while the adsorption chiller is used in the cooling season for providing air conditioning with the produced solar heat.
In the proposed arrangement of solar devices, the concentrator is used in order to increase the temperature of the working medium boosting up the thermal energy production of the system. Thus, the concentrator operation allows one to reach more frequently the desired temperature inside the tank in order to supply both heat pump and thermally driven chiller in winter and summer, respectively, and this, as a consequence, allows one to use more frequently the produced solar heat.
In winter, the dry cooler system is used as a heat source when the solar thermal energy is scarce, while in summer it is used to reject heat from the adsorption chiller and the heat pump operating as a chiller. An auxiliary heating boiler is used to match the DHW thermal demand in case of low solar radiation and when the produced heat by the solar devices and stored in the thermal storage is not sufficient to supply the adsorption chiller. Moreover, the reversible heat pump is also adopted in summer as an additional backup air conditioning device in case of low-temperature level in the thermal energy storage. The layout of the system has been shown in
Figure 2.
The hybrid SHC system consisted of seven loops designed to manage properly the thermal energy flows:
Hot Water, HW: water-glycol mixture heated by the solar thermal collectors and the concentrator, stored in the thermal storage and the domestic hot water tank and used as a heat source for the evaporator of the reversible heat pump during winter and the hot side of the adsorption chiller during summer.
Chilled Water, CHW: chilled water produced by the evaporator of the adsorption chiller, supplying the fan-coil system in cooling season;
Cooling Water, CW: water-glycol mixture circulating from the dry cooler to the condenser of the adsorption chiller and the reversible heat pump in summer, or to the evaporator of the heat pump in winter;
Heating-Cooling Water, HCW: hot or cold water supplying the fan-coil system for space cooling purposes;
Domestic Hot Water, DHW: sanitary water used by in the household;
Aqueduct Water, AW: mains water used to produce DHW;
Gas Boiler Water, GBF: water heated up in the gas boiler for purposes of DHW preparation.
The main components of the system are the following:
Solar collectors, SC: flat plate selective solar collectors used to produce thermal energy from the total solar radiation;
Concentrator, CONC: a parabolic dish concentrator with receiver equipped with a two-axis tracking system;
Thermal storage tank, TK1: an insulated tank with thermal stratification used to accumulate thermal energy generated by the solar loop;
Domestic hot water tank, TK2: a tank integrating two internal heat exchangers dedicated to the heating of water by the produced solar thermal energy and by the auxiliary heating device;
Natural gas boiler, GB: a natural gas-fired boiler used to heat TK2 and to heat up the hot water in order to run the adsorption chiller;
Sorption chiller, ACH: a thermally driven chiller consisting of a LiBr single-stage absorption chiller of a zeolite matrix-based adsorption chiller, used to produce chilled water;
Reversible heat pump, RHP: a vapor compression reversible unit used to produce heating water for space cooling and chilled water for space cooling when the adsorption chiller operation is not possible;
Dry cooler, DC: an air to fluid heat exchanger used to provide thermal energy for the heat pump in winter in case of low thermal energy stored in TK1, or used to dissipate the thermal energy rejected by the adsorption and vapor compression chiller when operating;
Hydraulic separator, HS: double inlet-outlet vessel used to separate the primary and the secondary heating system;
Fan coils, FC: water to indoor air heat exchangers used to provide heat and cool to the rooms of the building.
The main system components were connected to each other by means of pipes, diverters, mixers, and pumps, while the control and management model was developed using virtual temperature sensors and several controllers. The logic of the system operation has been reported below.
The working medium in the solar part of the HW loop is pumped by P1, activated when the solar radiation reaches 10 W/m
2 [
32]. Solar radiation is transformed in thermal energy with the use of both solar collectors and the concentrator, whose operation causes the increase of the water-glycol mixture temperature. The supply of HW to solar collectors and concentrator is controlled by double by-pass connections (M1/D1 and M2/D2), used to avoid the possibility of cooling of the working medium during circulating within the devices. The control system activates by-pass connections in case of a situation when the temperature at the SC or CONC outlet is lower than the inlet temperature.
During the heating season, the solar loop supplies HW to TK1 and TK2 in order to provide thermal energy for space conditioning and DHW production. TK1 tank has valves D3 and M3, which allow to manage the supply of HW. In the situation when HW temperature is higher than TK1 top temperature by 2 °C, the control system allows to supply HW to TK1. In a scenario when the HW temperature is lower than the TK1 top temperature, the by-pass D3/M3 is activated. This same control strategy is adopted for TK2 by means of M4/D4 in order to prevent the cooling of the bottom part of the tank. In this case, the water temperature in the proximity of internal heat exchanger of TK2 is used for the control strategy. In winter, the heating operation of the top part of TK1 is performed when its temperature drops to 15 °C, while it is stopped when the temperature increases to 25 °C. It is worth noticing that the TK1 temperature may decrease below the fixed set point due to the variation of weather conditions. From TK1, HW is supplied to the evaporator side of the reversible heat pump, which operates in order to keep a proper temperature for the FC system. In particular, RHP operates in order to keep the temperature at the outlet of HS between 40 and 42 °C. The set point temperature for the heated water by RHP was set to 40 °C. Moreover, in order to achieve the best possible performance of RHP, outside air may be supplied to DC for RHP operation. In fact, TK1 stops to supply heat to RHP when the outside air temperature increases by 2 °C above TK1 top temperature, while this operation stops once the air temperature decreases to the one inside TK1.
DHW is provided to the user at constant temperature. In the situation when TK2 top temperature exceeds 45 °C, mains water is mixed with the tank water in order to provide the required DHW temperature which is equal to 45 °C. In case of an insufficient amount of solar radiation, which is connected to TK2 top temperature below 55 °C, the control system activates gas boiler GB, which heats the water at the top part of TK2 to the set point of 65 °C. The deadband of 10 °C allows one to limit the number of activation of GB to heat TK2.
During the cooling season, the heating of TK1 by the solar loop is performed to keep its top temperature between a level allowing to run the sorption chiller: for the absorption unit the temperature range is 80–90 °C, while for the adsorption chiller it is 60–70 °C.
The activation of the absorption unit occurs once the temperature rises to 80 °C, and the deactivation is performed when the temperature drops to 75 °C. For the adsorption chiller, the required temperature level allowing to activate the unit is set to 58 °C, while the one that deactivates the chiller is fixed to 53 °C. It is worth notice that the selected temperatures for the activation and deactivation of both thermally driven chillers are compatible with the operation specification of small-scale devices provided by manufacturers [
33,
34].
During the operation of ACH, GB may supply auxiliary heat to drive the sorption chillers. In the case of absorption chiller, when the inlet temperature of the generator of ACH decreases to 77 °C, GB heats HW to 80 °C until the TK1 top temperature increases to 79 °C. In the scenario where the adsorption chiller is integrated into the system layout, when TK1 temperature drops to 55 °C, GB is turned on to heat HW to 60 °C until TK1 top temperature rises to 57 °C. In the operation of the space cooling system, the chilled water temperature is maintained at the level of 7 °C (by ACH and RHP), allowing the proper operation of the fan-coil system.
The control system turns off the sorption chiller ACH when the generator inlet temperature decreases below the allowed minimum threshold, then the electric chiller ECH is switched on. Furthermore, during the operation of the ACH, when the outlet temperature of the load side of the HS exceeds 12 °C, in case of a high space cooling demand of the user or a decrease of ACH chilling power, the auxiliary chilling equipment is activated. In the mentioned case, ACH is switched off and ECH is used to reduce HS outlet temperature to 10 °C.
The space conditioning system ensures a temperature inside the building rooms in a range of 20–22 °C and 24–26 °C [
11] during the space heating and cooling season, respectively. When the temperature inside the building is on a sufficient level, the FC system is turned off. Finally, cooling water CW, supplied by P3 to the dry cooler is used in order to dissipate the thermal energy rejected by the ECH condenser or ACH cooling circuit.
2.3. Model of the System
The system under investigation was modeled and simulated using Transient System Simulation (TRNSYS) software, which is a tool capable to simulate the transient operation of conventional, renewable, and new concept energy systems. The main features of this environment consist of a vast library of validated components, the possibility of implement user-defined components, and flexibility in developing the layout of the systems as well as their control and operation strategy.
In the frame of this paper is not possible to present each model of the adopted components for reasons of brevity, thus only the detailed description of the energy and economic analysis model is presented, since the other models are available in the TRNSYS software documentation containing the mathematical structure of components [
35] or are available in the literature. Nevertheless, some information about the adopted models (types) are reported below.
The flat-plate solar collector is modelled taking into account the thermal performance of a theoretical collector. In particular, the Hottel-Whillier steady-state model [
36] is used for evaluating the thermal performance, while the overall thermal loss coefficient of the collector per unit area is determined on the basis of Ref. [
37].
The model of the concentrator and the absorption chiller have been presented in Ref. [
30]. For the concentrator, the working medium temperature at the outlet of the receiver
To was calculated with the following equation:
where
Tin is the inlet temperature,
Q is the amount of heat transferred to the working fluid,
m is the mass flow and
cp is the specific heat.
Q is given by the formula:
where:
A is the receiver surface (aperture),
FR is the overall collector heat removal efficiency factor,
GT is the intensity of direct solar radiation at the receiver surface,
U is the overall thermal loss coefficient of the collector per area of the unit,
T is the average working medium temperature inside the receiver and
Ta is the ambient temperature.
FR coefficient is given by:
where
fp is the total efficiency of the receiver.
U coefficient was calculated taking into account the radiative and convective heat losses according the following equations:
where:
hr is the radiative loss coefficient,
hk is the convective loss coefficient [
38]
σ is the Stefan-Boltzmann constant,
εp is the receiver surface emissivity,
Tm is the average medium temperature inside the receiver,
Ts is effective sky temperature and
v is the wind velocity.
The model of the absorption chiller was based on mass and enthalpy balance equations from the point of view of water-LiBr mixture and only LiBr in the different point of heat exchangers of the device (generator, absorber, evaporator and condenser). For the LiBr in the weak and strong solution concentrations were assumed from Ref. [
34]. The thermal power of the
i-th
Qi heat exchangers was calculated as:
where
min and
mout represent the mass flow rate at the inlet and outlet,
hin and
hout represent the enthalpy at the inlet and outlet, respectively. The
COP of the absorption unit was calculated with the following equation:
where
Qeva and
Qgen is the thermal power of evaporator and generator, respectively.
The efficiencies
η of heat exchangers in all loops of the absorption chiller were given by the following formula:
where
UHE it the overall heat-transfer coefficient of heat exchanger,
AHE is the total heat transfer surface,
m is the flow rate of the given working fluid and
cp is its specific heat.
The adsorption chiller component was based on manufacturer data (Invensor LTC 10 E PLUS) [
34] and a model comprehensively described in a previous paper of the authors [
32]. The model uses user-supplied data files with cooling capacity and COP taken from manufacturer data as a function of different inlet temperatures of hot, chilled and cooling parts of the device and liquid mass flow rates.
The model for summer operation of the fan-coils is reported in Ref. [
15]. The model was based on correction factors taking into account for fluid mass flow rate, inlet fluid and air dry/wet temperature, and air flow rate.
The list of the adopted components has been provided in
Table 1.
Energy and Economic Model
In order to assess the performance of the novel hybrid system from the energy and economic point of view, a Reference System (RS) was adopter for the comparison with the proposed system (PS). RS consisted of an air-to-water reversible heat pump for space heating and cooling and a natural gas boiler used for DHW production. The analysis was performed assuming that both systems must provide an equal quantity of final energy to the user, in the form of space conditioning and DHW.
The consumption of primary energy of PS and RS was calculated at a boiler system efficiency of 0.85% [
15] and an averaged Polish electric network efficiency of 0.33 [
39]. In the calculation, the following energy flows were taken into account for RS:
energy consumed in the form of natural gas for the production of DHW;
electrical energy used by RHP operating with air as a heat source for space heating (seasonal COP = 2.5);
electrical energy used by RHP operating for space cooling (seasonal COP = 3.5);
while for PS:
energy consumed by GB for the integration of the heat needed to produce DHW and HW auxiliary heating for sorption chiller operation under low temperature available at the generator;
electrical energy used by RHP operating with the solar loop and air as a heat source for space heating;
electrical energy used by RHP operating as an auxiliary unit for space cooling.
Under these assumptions the following equations were adopted to evaluate the Primary Energy (PE) of both RS and PS and the Primary Energy Saving ratio (PESr) of PS:
where
PERS is the primary energy consumption of reference system,
Eth,GB,DHW,RS is the energy consumed for purposes of domestic hot water preparation for RS,
ηGB is the efficiency of the gas boiler,
Eth,heating is the energy consumed for heating purposes,
COPRHP,heating,RS is the coefficient of performance of the reversible heat pump in heating mode for RS,
ηel is electric grid efficiency,
Eth,cooling is the energy consumed for purposes of cooling,
COPRHP,cooling,RS is the coefficient of performance of the reversible heat pump in cooling mode for RS.
where
PEPS is the primary energy consumption of PS,
Eth,GB,DHW,PS is the energy consumed for purposes of domestic hot water preparation for PS,
Eth,GB,ACH,PS is the energy consumed for purposes of cooling with the use of sorption chiller for PS,
Eel,RHP,heating,PS is the electric energy for purposes of heating with the use of reversible heat pump for PS and
Eel,RHP,cooling,PS is the electric energy for purposes of cooling with the use of reversible heat pump for PS.
where
PESR is the primary energy saving ratio.
The economic parameters of PS were evaluated by calculating the investment costs of the mentioned system taken from manufacturers and operating costs of both RS and PS, according to the methodology available in literature [
40]. Parabolic dish concentrator with a double-axis tracking system costs was assumed to 118 €/m
2 and 825 €, respectively [
30]. Solar thermal collectors cost was 150 €/m
2 and for the absorption and adsorption chiller units was 300 and 500 €/kW, respectively [
39]. The cost of other components, like control system, pumps, etc., was included in the total costs of PS and the maintenance costs of PS and RS were ware assumed to be the same, allowing to neglect the effect of maintenance on the economic results.
To calculate the operating costs of both systems, the natural gas and electrical energy price were taken from the Eurostat data [
41], thus were set to 0.0425 and 0.1475 €/kWh, respectively. Costs of operation of both RS and PS systems (
Cop) and the savings (Δ
Cop) were calculated based on Equations (3) and (4).
where
Cop,RS is the cost of operation of the RS,
cng is the cost of natural gas,
Eel,RHP,heating,RS is the electric energy used for purposes of heating with the use of reversible heat pump, RS,
Eel,RHP,cooling,RS is the electric energy used for purposes of cooling with the use of reversible heat pump, RS and
cel are the costs of electric energy.
where
Cop,PS are the costs of operation, PS, and
Eel,auxiliaries,PS is the electric energy used for purposes of auxiliary elements, proposed system.
where Δ
Cop are the savings caused by implementation of the proposed system
The simple Pay Back period (SPB), defined as the ratio between the cost of the proposed system and the savings, was calculated to assess the economic results of the hybrid system.
2.4. Case Study
The case study adopted to investigate the system energy and economic performance consisted of a one floor single family household, with an attic, and a sloped roof (
Figure 3). The model of the building structure has been adopted in other papers of the authors [
30]. In the present case, the thermal load of the building were calculated within the developed model assuming Typical Meteorological Year (TMY) climatic conditions with the adoption of Meteonorm weather data of Cracow, Southern Poland. The building geometrical structure has been reported in
Figure 3. The ground floor area consisted of one room of 50 m
2 and two rooms of 25 m
2, while the attic had a useful area of 75 m
2. The floor height was 2.70 m and the slope of the roof was 30°. The building exposition with respect to the south-north direction has been set as shown in
Figure 3. The building envelope elements, as walls, roof, and floor, were modeled setting several layers for each component simulating the structure of a realistic building envelope. The information about building envelope components have been reported in
Table 2 in terms of thermal transmittance and structure.
The building hydronic air conditioning system consisted of one fan-coil for each zone of the building, operating from 15 November to 31 March and from 1 May to 15 October during the heating and cooling season, respectively. Regarding the daily operation, air heating was set for the whole day, while air cooling was assumed to work from 8:00 am to 10:00 pm. In the thermal model of the building, realistic thermal load and loads were implemented (
Table 3). The DHW demand was set to 60 L/person/day with a realistic profile during the day. The hourly space heating and cooling demand for the case study building is shown in
Figure 4.
The case study in terms of system configuration was completed assuming several design and operational parameters of all components. The main parameters have been listed in
Table 4.
In order to determine the values of such parameters, manufacturer data were used and an iterative sizing/design procedure was performed, allowing to achieve a satisfactory system configuration from the point of view of energy generation, dynamic operation, and capability to cover the thermal demand of the building users. It is worth noticing that for the solar collecting devices the size was selected taking into account their thermal energy production in order to meet a part of the user demand. Thus, the aperture area of SC and CONC are higher than the ones of the units present in the experimental installation. The selection of different areas was possible because the models of the components are validated, thus the reliability of the simulations is independent of the selected configuration of each solar device.