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Thermodynamic performance assessment of a small size CCHP (combined cooling heating and power) system with numerical models
Energy xxx (2013) 1e10
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Thermodynamic performance assessment of a small size CCHP (combined cooling heating and power) system with numerical models E. Jannelli a, M. Minutillo a, R. Cozzolino b, *, G. Falcucci a a b
University of Naples “Parthenope”, Centro Direzionale, Isola C4, 80143 Naples, Italy University of Niccolò Cusano, Via Don Carlo Gnocchi, Rome, Italy
a r t i c l e i n f o
a b s t r a c t
Article history: Received 9 July 2013 Received in revised form 30 October 2013 Accepted 25 November 2013 Available online xxx
The aim of the this work has been to evaluate the performance of a small-size CCHP (Combined Cooling Heating and Power) system based on the integration of 20 kW Lombardini diesel engine and a double effect water-LiBr absorption chiller. This integrated system has been designed to produce both hot water, by recovering heat from the engine cooling system, and chilled water, by recovering heat from the engine exhaust gasses (the exhaust gasses are sent to the fired-combustor of the absorption chiller). The analysis has been conducted by using numerical simulations: the engine and the absorption chiller have been modeled by means of 0e1D dimensional and thermochemical models, respectively, and the validation procedure has been performed by using the available operating data. The system performance has been calculated by introducing some performance parameters that have allowed: i) to estimate the efficiency of the primary energy conversion into useful energy EUF (energy utilization factor), ii) to consider the quality difference between cool/heat and work and among heats at different temperatures ExUF (exergy utilization factor); iii) to evaluate the primary energy saving with respect to the separate production of the same energy fluxes TPES (trigeneration primary energy saving). Ó 2013 Elsevier Ltd. All rights reserved.
Keywords: Polygeneration Diesel engine Absorption chiller Numerical models Performance coefficients
1. Introduction In recent years, the development of polygeneration systems has received an increasing attention in the area of small scale power systems for different applications ranging from residential to yachting utilities [1e5]. The residential sector is an important consumer of energy that could receive substantial benefits from the extensive utilization of trigeneration technologies. In fact, instead of satisfying electricity, heat and cooling demands separately, a combined production could provide considerable energy saving and an important emission reduction, as well. In the yachting sector the replacement of the traditional vapor compression units with the absorption machines may represent an interesting alternative that allows to reduce the on board energy consumption, thanks to the heat recovery on the exhaust gases and/or to the cooling circuits from internal combustion engine. The interest in small-size CCHP (Combined Cooling, Heating and Power) systems is high, due both to the energy saving obtained * Corresponding author. Tel.: þ39 0645678350. E-mail addresses: [email protected] (E. Jannelli), mariagiovanna. [email protected] (M. Minutillo), [email protected] (R. Cozzolino), [email protected] (G. Falcucci).
with the recovery of waste heat and to the ability in satisfying the energy demand of several types of stand-alone utilities. The research in designing, development and optimization of these technologies [6e8] and in the operational planning is considerably active (strategy concerning operational state of the equipment, energy flow rates, purchase/selling of electricity, etc.), according to energetic and economic issues [9e11]. In particular, the research is carried out by using both experimental and numerical approaches. In this paper a numerical analysis of a micro-scale trigeneration system, based on an internal combustion engine integrated with a water/LiBr double-effect absorption chiller, has been carried out by 0e1D dimensional and thermochemical models.
2. Background CCHP systems do not represent a new technology: large-scale systems from a few hundreds of kWs to MWs are intensively studied and illustrated in the technical literature. Some studies have examined the performance and the optimal operating strategies of trigeneration systems based on different prime movers such as internal combustion engine [12e14], gas turbine [15e18] and fuel cells [19,20].
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Please cite this article in press as: Jannelli E, et al., Thermodynamic performance assessment of a small size CCHP (combined cooling heating and power) system with numerical models, Energy (2013), http://dx.doi.org/10.1016/j.energy.2013.11.074
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Nomenclature ABS A/F a C CCHP COND COPSP EVP EUF ExUF fPC FSP FCCHP h hBB hchill-in hchill-out HTG LTG LHV mBB mc _ chill m _f m _F m
absorber air to fuel ratio vibe parameter () cylinder combined cooling heating and power condenser coefficient of performance of an equivalent compression electric refrigerator group evaporator energy utilization factor () exergy utilization factor () post-combustion factor () total fuel energy input required for the separate production (kW) total fuel energy input to the trigeneration system (kW) heat transfer coefficient (J/m2 K) enthalpy of the blow-by (J/kg) enthalpy of entering chilled water (kJ/kg) enthalpy of leaving chilled water (kJ/kg) high temperature generator low temperature generator lower heating value mass flow rate of the blow-by (kg/s) mass flow rate inside the cylinder (kg/s) chilled mass flow rate (kg/s) fuel consumption of the chiller fired-combustor (kg/s) fuel consumption of the engine (kg/s)
Balli et al. [12,13] presented a methodology and relations for thermodynamic and thermoeconomic evaluation performance of a trigeneration system with a rated output power of 6.5 MW gasdiesel engine integrated with an absorption chiller. The proposed methodology includes the definition of performance parameters and the cost analysis formulation based on a specific exergy cost method. The results of their study can be beneficial to change the components that have low thermodynamic efficiencies and large exergy consumptions, allowing to regulate the sale price of the products and to review the plant’s economic policy. Bruno et al. [15] studied the integration of four microturbines (in the range 30e100 kWe) with a double effect direct-fired absorption chiller. The exhaust gases were directed to the chiller with the option of additional natural gas post-combustion. Thus, the authors analyzed the effect of the post-combustion degree on the trigeneration system performance and determined the working conditions that, in the proposed integrated system, allowed to obtain the maximum efficiency. Thus, the authors demonstrated that a directly driven absorption chiller with a post-combustion system can introduce advantages with respect to the more conventional single effect hot water system in terms of higher COP (coefficient of performance) and flexibility. This is due to the decoupling between the electricity and the chilled water production. The analysis was carried out by using a numerical model, developed with the software package EES (Engineering Equation Solver). Ho et al. [16] studied a cogeneration system that consists of a single effect commercial absorption chiller integrated with a Capstone microturbine (30 kW). Their results pointed out that the electric efficiency was 21% and the overall system efficiency was 46%. Huicochea et al. [17] studied the performance of a CCHP system based on a double effect absorption chiller driven by the exhaust
MP pc PC PES PL Pe Qc QF Qf QHTG Qt Qw SB Tg Ti TPES u V
measurement point pressure inside the cylinder (Pa) post-combustion primary energy saving () plenum engine electric power (kW) chiller cooling power (kW) total fuel heat entering to the engine (J) fuel power entering to the fired-combustor (kW) thermal power to the HTG (kW) thermal power recovered by the engine cooling system (kW) wall heat loss of the engine (J) system boundary temperature of the engine exhaust gas (K) temperature of the engine wall (K) trigeneration primary energy saving () specific internal energy (J/kg) cylinder volume (m3)
Greek letters a Crank angle ( ) a0 Crank angle at the start of combustion ( ) Dac combustion duration ( ) 4 and 4 parameters based on the temperature at which the heat is available () hSP electric reference efficiency for the separate e production () hSP thermal reference efficiency for the separate t production ()
gas of a 30 kWe microturbine. The authors developed a thermodynamic simulator by using Matlab programming language. The model was based on the mass, species and energy conservation equations for each component of the absorption chiller. The modeling allowed to predict the trigeneration performance at different operating conditions such as the ambient temperature and the microturbine mass flow rates. In particular, the decreasing tendency of all performance parameters (i.e. COP and electric efficiency) with the increase of ambient temperature was shown. The results pointed out that the proposed system for the co-production of electric, cooling and heating powers based on the microturbine technology represents an attractive solution in the fields of the distributed generation. Chen et al. [18] investigated behavior and performance during the off-design operation of a small-scale gas turbine (1747 kWe) coupled with a double effect chiller and a heat exchanger. The study was carried out by using a mathematical model, written in Fortran code and based on energy and exergy rate balances. The results showed that in off-design conditions, the performance of the gas turbine changed dramatically, whereas the performance of the chiller changed slightly. The calculated efficiency of the gas turbine ranged from 27 at full load to 11% at partial load, while the COP increased slightly with the decreasing of the load level of the CCHP system. Thus, the performance deterioration of the system was due to the bad performance of the gas turbine under the off-design conditions. Even if there are no distinctions between large-scale and microscale trigeneration systems and absorption chillers, the designing criteria and the operating conditions are different. Thus further and detailed experimental and numerical studies have to be conducted in order to allow the development and the optimization of these technologies, according to their configurations and applications.
Please cite this article in press as: Jannelli E, et al., Thermodynamic performance assessment of a small size CCHP (combined cooling heating and power) system with numerical models, Energy (2013), http://dx.doi.org/10.1016/j.energy.2013.11.074
E. Jannelli et al. / Energy xxx (2013) 1e10
Wang et al. [21] and Lin et al. [22] analyzed the performance of micro-scale trigeneration systems that use as prime mover an internal combustion engine. Wang et al. [21] examined a trigeneration system consisting of an engine fueled by hydrogen (6.5 kWe) whose waste heats, rejected from exhausts and engine cooling system, are used for domestic purpose (hot and cooling water). The authors simulated the trigeneration system by using the function of the energy and mass balance of the computational software called ECLIPSE. Their study showed that the hydrogen is a very interesting fuel that allows to achieve equal or better performance to the conventional diesel fuel in terms of energetic performance and near zero carbon emissions. Thus, the authors pointed out the enormous potential fuel savings and massive reductions in greenhouse gas emissions per unit of useful energy outputs with cogeneration and trigeneration compared with that of single generation. Lin et al. [22] studied and realized a small-scale trigeneration system based on 9.5 kW diesel engine coupled with both a heat recovery system (for heating purposes) and an absorption refrigerator (for cooling energy requirements). The experimental results, obtained by using their experimental apparatus, pointed out that if the engine load is over 50%, the exhaust gases are hot enough to drive the absorption refrigerator allowing very low temperature (the lowest evaporator temperature at the point in freezer of the refrigerator is 24.9 C; the temperature at the point in food compartment is 8.2 C while the refrigerator is powered by the heat from engine exhaust gas at the engine load of 75%). In the technical literature a limited number of studies deals with behavior and performance of CCHP systems in the power ranges less than 20 kWe because, in the past, their high costs limited the diffusion of these systems and consequently their designing and development; only in recent years the interest in these small-scale systems is increased, mainly as a result of the policies for the promotion of the DG (distributed generation). Some countries have implemented a legislation that offers economic bonuses to cogeneration/trigeneration systems so that, their development is become more profitable for companies that want to invest in the decentralized generation [23]. 3. System description The CCHP system, investigated in this paper, consists of a small size diesel engine and a double effect water/LieBr absorption chiller. The integration between these units is realized by introducing the engine exhaust gases into the fired-combustor of the high temperature generator of the chiller. Furthermore in order to satisfy the heat requirements, according to the cooling loads, fresh air and additional fuel can be added too. Waste heat from the engine cooling system is recovered in order to produce hot water (thermal power production). The main advantages of using a double effect absorption chiller driven with a fired-combustor system, over the conventional single-effect hot-water driven system, are both a higher COP and the decoupling between electric and cooling powers (wider ranges of operating conditions are allowed). Since only few commercial chillers are available in a small capacity range, the choice of the engine size is constrained both by the electric power demand and the cooling power requirements. The CCHP system proposed in this study is based on the integration of commercial units: the engine is a 20 kWmech Lombardini diesel engine [24] and the absorption chiller is a 23 kWc doubleeffect water/LiBr chiller commercialized by Broad [25]. The absorption system can work not only as chiller but also as heater.
3
Fig. 1. Interconnection in the outputeinput units models.
However, in this analysis the operation of the chiller as heater was not considered. 4. System modeling The CCHP system has been studied by using original models developed by the authors. In particular, the engine and the absorption chiller have been modeled by means of 0-1D dimensional and thermochemical models, respectively, and the validation procedure has been performed by using the available operating data. Fig. 1 sketches the interconnection in the outputeinput units models. 4.1. Engine simulation The behavior and the performance of the Lombardini diesel engine, whose characteristics are summarized in Table 1, have been determined by means 0e1D dimensional model, developed by the authors using AVL BOOST code [26]. The model describes the phenomena that take place inside the cylinder and the engine ducts; the operating parameters, such as exhaust gas mass flow, pressure and temperature, are computed at different engine speeds (the transient conditions due to the variation of the engine speed, are not taken into account). The equation, that describes the engine operation, is reported as follows:
dðmc $uÞ dV dQF X dQw dmBB ¼ pc $ þ hBB $ da da da da da
(1)
in which V is the cylinder volume, QF is the total fuel heat input, Qw is the wall heat loss, mc is the mass inside the cylinder, pc is the pressure inside the cylinder, u is the specific internal energy, a is the
Table 1 The main engine characteristics [24]. Cylinder Displacement Bore Stroke Compression ratio Rating kW Max torque Minimum idling speed Vol. of air required for the correct combustion @3600 rpm
n. cm3 mm mm N(80/1269/CEE) iso 1585 Nm m3/min
3 1028 75 77.6 22.8:1 20 67 @2000 900 63
Please cite this article in press as: Jannelli E, et al., Thermodynamic performance assessment of a small size CCHP (combined cooling heating and power) system with numerical models, Energy (2013), http://dx.doi.org/10.1016/j.energy.2013.11.074
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From the total heat supplied to the cycle, which is determined by the amount of fuel in the cylinder and by the A/F (Air/Fuel) ratio, the model calculates the actual heat input per degree crank angle. The combustion process is solved with a 0-dimensional model by using the standard single-Vibe approach (A/F ¼ 22 has been chosen for full load operating condition at different regimes). The Vibe function is: ðmþ1Þ dx a ¼ $ðm þ 1Þ$ym $ea$y Dac da
Fig. 2. Flowsheet of the engine model. PL: plenum, MP: measurement point, SB: system boundary, C: cylinder.
crank angle and hBB and mBB represent the enthalpy and mass of the blow-by, respectively. In order to solve this equation, models for the combustion process and the wall heat transfer, as well as the gas properties as a function of pressure, temperature and gas composition, are required.
(2)
with dx ¼ dQF/QF and y ¼ (a a0)/Dac, where a0 is the crank angle of the start of the combustion, Dac is the combustion duration and a is the Vibe parameter, which is 6.9 for complete combustion [26,28]. The heat transfer through the engine walls is simulated by means of the Woschni approach [28], by imposing the surface (exchange) areas of piston, cylinder head and valve openings. The heat is calculated as:
Qw ¼ h A Tg Ti
(3)
Fig. 3. Schematic layout of the double effect Water/LiBr absorption chiller.
Please cite this article in press as: Jannelli E, et al., Thermodynamic performance assessment of a small size CCHP (combined cooling heating and power) system with numerical models, Energy (2013), http://dx.doi.org/10.1016/j.energy.2013.11.074
E. Jannelli et al. / Energy xxx (2013) 1e10 Table 2 The main absorption chiller nominal operating data [25]. Chilled power Temperature of chilled water Chilled water flow rate Maximum gas consumption COP
kW C m3/h m3/h e
23 7e14 2.9 2.2 1.1
in which the heat transfer coefficient h is related to the main thermodynamics parameters of the engine [28,29], Tg and Ti are the gas temperature and the wall temperature, respectively. The flux inside pipes and connections is solved with a 1D approach considering the distributed pressure losses (the roughness coefficient is assumed equal to 0.019 [28]). An average cell size of 20 mm is chosen for the discretization of the 1D elements, i.e. the pipes. Fig. 2 depicts the flowsheet of the engine model. For internal mixture preparation it is assumed that: the fuel added to the cylinder charge is immediately burned; the combustion products mix instantaneously with the rest of the cylinder charge and form a uniform mixture; the A/F ratio of the charge diminishes continuously from a high value at the start of the combustion to the final value at the end of the combustion.
4.2. Absorption chiller simulation The selected 23 kWc double effect water/LiBr absorption chiller is a commercial unit that consists mainly of: a low pressure generator,
5
a high pressure generator, an absorber, an evaporator, a condenser, a pump, throttling valves and a cooling tower. Fig. 3 depicts its schematic layout and in Table 2 the main operating data are summarized. Since the absorption chiller plays an important role in recovering the waste heat, special attention has been devoted to study its performance at design and off-design conditions. The model of the absorption chiller is a thermochemical model developed by using the CHEMCAD code [27]. It is based on the ELECNRTL property method and the electrolyte wizard option has been used to generate a series of reactions (the relevant reaction is the association/dissociation of lithium bromide). The model is based on the numerical approach proposed by Somers et al. [30]. The assumptions are as follows: the analysis has been made under steady-state conditions the refrigerant (water) at outlet state of the condenser is saturated liquid the refrigerant (water) at outlet state of the evaporator is saturated vapor the pressure losses in pipelines and all heat exchanger are negligible the heat losses in the heat exchangers are negligible the expanding process in the throttling valves is isenthalpic. In Fig. 4, the flowsheet of the model is reported. The scheme consists of five main sections that simulate the HTG (high temperature generator), the LTG (low temperature generator), the COND (condenser), the EVP (evaporator) and the ABS (absorber), respectively. The cooling tower is not simulated because, in this first analysis, the power consumption of the chiller is not considered and the
Fig. 4. Flowsheet of the absorption chiller model.
Please cite this article in press as: Jannelli E, et al., Thermodynamic performance assessment of a small size CCHP (combined cooling heating and power) system with numerical models, Energy (2013), http://dx.doi.org/10.1016/j.energy.2013.11.074
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temperature of the cooling water can be varied as each input data (i.e. for a yachting utility the cooling tower is not necessary and the chiller is simplified in its layout). The refrigerant rich solution water/LiBr (water is the refrigerant and LiBr is the sorbent), that enters in the state (IN), is pumped (PUMP1) up to the maximum pressure of cycle and then it is split in two streams (2) and (5). The streams (2) and (5) are pre-heated in a heat exchangers (HEX4) and (H-EX6) by subtracting heat to streams (6) and (13) that are the LiBr rich solution throttled back to the absorber (ABS). The HTG has been simulated by using two components: a directfired combustor (COMB), that allows the heating of the refrigerant solution, and a flash separator (HT-SEP), in which the refrigerant vapor is separated and sent to the LTG. Similarly, the LTG has been simulated by using a heat exchanger (LT- HEX) and a further flash separator (LT-SEP). The primary refrigerant vapor (7) is cooled in the LTG allowing the vaporization of secondary refrigerant vapor (15) that is sent to the COND (condenser). The refrigerant liquid (18), after leaving the mixer (MIXER1), is throttled and sent to the EVP (evaporator) where the cooling effect is obtained. Thus, the refrigerant liquid, that is in the vapor phase, is absorbed by the LiBr rich solution (ABS). The circuit of the water/LiBr solution is closed, thus, the chemical and thermodynamic conditions of the outlet stream (OUT) are set equal to those of the input stream (IN). 4.3. Models validation The engine model has been validated by using the nominal data declared by the manufacturer [24]. In particular, the engine parameters, such as the heat transfer coefficient, the air/fuel ratio, the injection advance and the combustion duration, have been tuned in order to fit the power, the torque and the specific fuel consumption trends, as provided by manufacturer. Fig. 5 highlights the good agreement between the numerical results and the nominal data obtained with the tuned engine parameters. Table 3 shows the main engine operating data calculated by the model. The electric power has been determined assuming the efficiency of the electric generator equal to 0.97. The absorption chiller model validation has been carried out by analyzing some performance and operating parameters such as the
Fig. 5. Engine performance in terms of mechanical power and fuel consumption: comparison between calculated and nominal data [24].
Table 3 The engine operating conditions calculated by the model. RPM
Pmech [kW]
Pel [kW]
Fuel [kg/h]
Air [kg/h]
Exhaust gases mass flow [kg/h]
Exhaust gases temperature [K]
Qw [kW]
1500 2000 2400 2800 3200 3600
7.5 11.5 14.2 16.25 17.9 18.2
7.3 11.2 13.8 15.8 17.4 17.7
2 2.9 3.4 4 4.5 4.9
43 61.9 72.2 86 107 117
45 64.8 75.6 90 111.6 122.4
620 665 700 715 715 718
12.0 16.5 18.5 19.8 21.5 21.4
COP (coefficient of performance), the sorbent solution and the refrigerant mass flow rates. The COP values (thermal COP) have been calculated by applying the following equation:
COP ¼
_ ðh Qc hchillout Þ m ¼ chill chillin _ f LHV Qf m
(4)
in which, the numerator represents the cooling power (or cooling capacity) at the evaporator and the denominator is the fuel power input to the fired-combustor (based on the diesel Low Calorific Value). The data used for the model validation have been derived by the study conducted by Yin et al. [31], in which the 16 kWc double effect water/LiBr absorption chiller, produced by the same manufacturer, was modeled and tested. Thus, because the 16 kWc and 23 kWc absorption chillers are very similar in terms of behavior and performance, the author used the mentioned study as the reference case for the validation (experimental data on the 23 kWc were not available). Fig. 6 depicts the COP vs. the cooling load, whose variation was obtained by decreasing the chilled-water return temperature from 14 C to 8 C, while the chilled and cooling water flows and the cooling water inlet temperature were maintained constant. The numerical COP, calculated by the model, is compared to the numerical and the experimental COP of the reference case [31]. It can be noted that the trend and the values are very close (the deviation with respect to the experimental data ranges from 2.5% to
Fig. 6. Chiller performance (COP) under various load conditions: the comparison between the numerical results and the reference data [31].
Please cite this article in press as: Jannelli E, et al., Thermodynamic performance assessment of a small size CCHP (combined cooling heating and power) system with numerical models, Energy (2013), http://dx.doi.org/10.1016/j.energy.2013.11.074
E. Jannelli et al. / Energy xxx (2013) 1e10
4.5%), thus the model can be considered reliable for the analyzed operating conditions. In Table 4, the state point results from the 23 kWc chiller model (state points are defined in Fig. 3) are summarized. Fig. 7 shows the flow rates of the dilute sorbent solution from the absorber to the HTG and LTG at different load conditions. It can be noted that the sorbent solution flow rate in the LTG decreases with the cooling load decreasing, while the sorbent solution flow rate in the HTG is relatively constant. This trend in the HTG depends on the control strategy of the absorption chiller that, by using a variable frequency solution pump, has to maintain the same flow rate of sorbent solution in the HTG. Thus, the ratio between the sorbent solution in the HTG and the total solution flow rate from the absorber (sorbent solution split ratio) varies between 0.5 at design load to 0.7 at minimum cooling load, as it is explained in Ref. [31]. Fig. 8 depicts the refrigerant flow rates vaporized in the evaporator at different load conditions. It is worth noting that the refrigerant flow rate is proportional to the cooling load and the HTG generates more refrigerant than the LTG. The trends of the refrigerant flow rates are in accordance with the results of the reference case, even if the values are higher because the size of the absorption chiller is bigger. 5. Results and discussions The developed models have been used both to define the operating range of the CCHP system in terms of electric, thermal and cooling powers and to forecast its performance. Some of the
Fig. 7. Sorbent solution flow rate under various cooling capacities: the comparison between the numerical results and the reference data [31].
output data of the engine model have been used as input data for the chiller model. In particular, in order to connect match the electric power to the temperature and mass flow rate of the exhausts gasses (input data for the chiller model) the following relationships have been determined:
Table 4 The state point results calculated by the model. Point General IN 1 2 5 8 9 14 16 18 19 20 WATERIN HTG 3 4 6 7 LTG 10 11 12 13 15 COND 17 25 WATEROUT EVP 21 22 CHIL-IN CHIL-OUT ABS 23 24 OUT
Mass flow (kg/s)
Temp. ( C)
_ Exhaust ¼ m
Press. (kPa)
LiBr (%)
0.28 0.28 0.14 0.14 0.132 0.14 0.134 0.008 0.014 0.266 0.266 0.829
35.92 35.92 35.92 35.95 42.0 35.98 42.0 40.47 40.47 42.01 42.01 30.0
0.799 91.99 91.99 91.99 7.59 7.59 7.59 7.59 7.59 7.59 0.799 100
57.79 57.79 57.79 57.79 61.12 57.79 61.12
0.14 0.14 0.132 0.008
138.73 154.89 154.89 154.89
91.99 91.99 91.99 91.99
57.79 57.79 61.12
0.14 0.008 0.14 0.134 0.006
75.87 97.31 85.25 85.24 85.24
7.59 91.99 7.59 7.59 7.59
57.79
0.006 0.829 0.829
40.47 38.76 42.10
7.59 200 200
0.014 0.014 0.805 0.805
3.68 3.68 14.0 7.0
0.799 0.799 78.45 78.45
0.28 0.829 0.28
40.54 30.0 36.26
0.799 200 0.799
H2O (%)
100
0 0 0 0 0 0 0 0.098 0.055 0 0 0
100
0 0.082 0 1
100 100 61.12 61.12
100 57.79 61.12 61.12
TExhaust ¼
0:15$Pe3 5:08$Pe2 þ ð60:53$Pe Þ 184 3600
0:14$Pe3 þ 4:84$Pe2 ð40:43$Pe Þ þ 439
(5)
(6)
5.1. The CCHP system operating range The integration between the sub-units of the CCHP system has been realized by recovering the waste heat from the engine exhaust
0 0 0.065 0 0
100 100 100
0 0 0
100 100 100 100
0.1 0.8 0 0
100
0.059 0 0
57.79 57.79
Vapor frac. ()
7
Fig. 8. Refrigerant flow rate under various cooling capacities: the comparison between the numerical results and the reference data [31].
Please cite this article in press as: Jannelli E, et al., Thermodynamic performance assessment of a small size CCHP (combined cooling heating and power) system with numerical models, Energy (2013), http://dx.doi.org/10.1016/j.energy.2013.11.074
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gasses. These gases are recirculated to the fired-combustor of the absorption chiller where, if the available thermal power does not allow to satisfy the energy demands of the HTG (high temperature generator), additional fuel is used (the post-combustion of the exhaust gasses is realized). Furthermore, in order to improve the energy content in the HTG and, consequently, to increase the cooling load, additional air can be used. The presence of the fired-combustor allows to decouple the cooling power from the electric power. Furthermore, in order to satisfy the thermal demands, the waste heat from the engine cooling system is recovered by producing hot water. The proposed and studied CCHP system is able to operate in the following operating ranges: electric power: 7e18 kWe thermal power (hot water production): 10e17 kWt cooling power: 4e23 kWc Fig. 9 depicts the CCHP system operating field in terms of electric power vs. cooling power. In the figure, it is possible to identify 3 operating areas: - “area 234” (without PC zone), in which the waste heat from the engine exhaust gases is used as thermal input in the HTG; - “area 1245” (PC zone), in which the engine exhaust gases are used as oxidant in the fired-combustor of the chiller in order to provide additional heat to the HTG (the minimum oxygen concentration of 5% is assured in the exhaust gasses) for increasing the cooling capacity; - “area 1567” (PC þ Air zone), in which fresh air is provided to the fired-combustor in order to increase the cooling capacity until to the nominal value (23 kWc).
fpc ¼
_f m _F m
(7)
_ f and m _ F are the fuel consumptions of the chiller firedwhere m combustor and of the engine respectively. The post-combustion factor equal to 0 indicates that the integrated system does not use additional fuel into the fired-combustor of the absorption chiller. Fig. 10 shows the post-combustion factor vs. the electric power according to the operating field illustrated in Fig. 9. It can be noted that: - if the engine exhaust gasses are used as oxidant in the chiller fired-combustor (PC curve), fpc remains relatively constant, because the additional fuel is much lower compared to the engine fuel flow; - if both fresh air and fuel are added to the engine exhaust gasses (PC þ Air curve), the fpc curve decreases with the increase of the electric power, because more heat has to be provided at low electric power, in order to produce 23 kWc. The parameters introduced for evaluating the performance of CCHP system are the EUF (Energy Utilization Factor) [32] and the ExUF (Exergy Utilization Factor). The first one takes into account the efficiency in the conversion of the primary energy into work and heat (or cool). However, this performance coefficient does not discriminate between the electric power and the thermal and/or cooling power, while ExUF considers the quality difference between heat and work and among heats at different temperatures (heat at high temperature in the HTG absorption chiller and heat at low temperature from the engine cooling system) [33]. These parameters are defined as follows:
EUFCCHP ¼
Pe þ Qt þ Qc _ f LHV _ F LHV þ m m
(8)
5.2. The CCHP system performance
ExUFCCHP ¼
Pe þ Qt 4 þ fQHTG _ f LHV _ F LHV þ m m
(9)
In the paper the analysis of the CCHP system behavior has been carried out by using some operating and performance parameters. In order to take into account the additional fuel used into the fired-combustor with respect to the engine fuel consumption, the parameter fpc (post-combustion factor) has been defined as:
where Pe is the engine electric power (the auxiliaries power consumption of the absorption chiller has not been considered), Qt is the thermal power recovered by the engine cooling system
Fig. 9. CCHP system operating field in terms of electric power and cooling power.
Fig. 10. Post-combustion factor as function of the engine the electric power.
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E. Jannelli et al. / Energy xxx (2013) 1e10
9
(Qt ¼ ε$Qw where ε is the thermal recovery efficiency of the engine cooling system), QHTG is the thermal power to the HTG (the thermal input to the absorption chiller), 4 and f are parameters whose value changes between 0 and 1. These parameters are calculated by applying the following equations:
4 ¼
f¼
1
1
Ta
! (10)
TxðecsÞ
Ta TxðHTGÞ
! (11)
where Ta is the ambient temperature (reference temperature), Tx(ecs) and Tx(HTG) are the temperatures at which the heats for cogeneration and trigeneration purposes, are available respectively. These temperatures are calculated as the logarithmic mean temperatures between the inlet and the outlet temperatures of the heating sources (the temperatures at the engine cooling system and the temperatures at the HTG). The calculated values of 4 and f are 0.17 and 0.4 respectively. In Fig. 11 the EUF and the electric efficiency are shown as function of electric power. It can be noted that when the contribution of the post-combustion increases, EUF increases; as a matter of fact the highest EUF values are obtained when the exhaust gases and the additional air are mixed and used as combustion air (PC þ Air curve) to the fired-combustor. This allows to provide the thermal energy required to the absorption chiller for producing 23 kWc. EUF, calculated at maximum electric power and without the post-combustion, is equal to 0.78. Furthermore, it can be noted that the trend of EUF is in accordance with the trend of the electric efficiency. Fig. 12 depicts the trend of ExUF as function of the electric power. It is worth noting that the highest ExUF values are achieved without the post-combustion of the exhaust gasses (without PC curve). This behavior depends on the exergy destruction that occurs when the fired-combustor is used to increase the cooling power availability. As a matter of fact, its lowest value (0.38) is obtained at 7 kWe and 23 kWc with a post-combustor factor equal to 0.45. Furthermore, in order to evaluate the fuel saving that can be achieved with the co-production of electric, thermal and cooling powers, a new generalized performance indicator, called TPES (Trigeneration Primary Energy Saving), has been calculated, as
Fig. 12. System performance: the ExUF (Exergy Utilization Factor) as a function of electric power.
proposed in Ref. [34]. In fact, instead to calculated the PES coefficient, according to the European Directive 2004/08/EC, the coefficient TPES takes into account the cooling power explicitly as well as the reference efficiencies for comparison with the separate energy production. This parameter is defined as:
TPES ¼
F SP F CCHP F CCHP ¼ 1 SP F SP SP Pe =h þ Qt =h þ Qc = hSP $COPSP e
e
t
(12) where FCCHP and FSP are the total fuel energy input to the trigeneration system and the total fuel energy input required for the separate production of the same energy vectors (work, heat and cool) respectively. Furthermore in the eq. (12), the coefficients hSP e , SP hSP are the separate production reference efficiencies that t , COP have been assumed equal to 0.4, 0.9 and 4 (COP of an equivalent compression electric refrigerator group) respectively. These values are the average efficiencies of technologies used for the separate production of the same energy fluxes, as suggested in Ref. [34]. Instead of calculating the TPES for a assigned case study, (it is necessary to know the electric, thermal and cooling energy demands) in this analysis it has been chosen to estimate the TPES in the main working points of the CCHP system operating field, (Fig. 9). Thus, the analysis has been carried out considering the power values and not the annual energy values. In Table 5 the calculated data are illustrated. It can be seen that the TPES ranges from 0.03 to 0.21. The negative values are obtained at low electric power (7 kW) and when the fuel consumption in the fired-combustor is high. Thus, the CCHP system has good performance in a wide operating range, but the better performance can be achieved when the Table 5 The TPES (trigeneration primary energy saving) values calculated in the operating points illustrated in Fig. 9.
Fig. 11. System performance: the EUF (Energy Utilization Factor) as a function of electric power.
Points
1
2
3
4
5
6
7
FCCHP [kW] Pe [kW] Qt [kW] Qc [kW] TPES
27.9 7.3 9.6 3.8 0.11
35.4 11.2 13.2 3.8 0.21
58.8 17.7 17.1 3.8 0.10
58.8 17.7 17.1 11.4 0.16
68.5 17.7 17.1 21.9 0.11
69.5 17.7 17.1 23 0.10
44.79 7.3 9.6 23 0.03
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cooling power demand is low so that the additional fuel consumption in the fired-combustor is not necessary or is low. 6. Conclusion In this paper the performance of a micro-scale CCHP system, that consists of a diesel engine and a double effect water/LiBr absorption chiller, has been evaluated by means of numerical simulations. The engine and the absorption chiller have been modeled by using 0-1D dimensional and thermochemical codes respectively. The models, validated by using the available operating data, have been integrated in order to evaluate the operating field of the CCHP system in terms of electric, thermal and cooling powers. The performance analysis has been carried by introducing some performance coefficients such as: EUF, ExUF and TPES. The numerical results have highlight that the CCHP system has good performance in a wide operating range. However the better performance are achieved when the cooling power demand is low so that the additional fuel consumption in the fired-combustor is not necessary (this means high values of ExUF and TPES) or is low. Finally, in analyzing the operation of this integrated system, some practical considerations must be taken into account, such as the maximum mass flow rate that can be treated in the firedcombustor according to its designing conditions. This constraint can involve a modification to the fired combustor increasing the costs of the system and thus its application as micro-scale energy conversion system in grid-connected or stand-alone utilities. The paper has highlight that the numerical models can be useful tools in designing and optimization of energy conversion systems allowing to investigate on their behavior in different operating points and thus to assist to the experimental activities. Acknowledgments The authors would like to thanks the laboratory of energy systems of University of Cassino e Southern Lazio for the possibility of using their licensed code AVL BOOTS. References [1] Wu DW, Wang RZ. Combined cooling, heating and power: a review. Prog Energy Combust Sci 2006;32:459e95. [2] Fabrizio E. Feasibility of polygeneration in energy supply systems for healthcare facilities under the Italian climate and boundary conditions. Energy Sustain Dev 2011;15:92e103. [3] Bingöl E, Kılkıs B, Eralp C. Exergy based performance analysis of high efficiency poly-generation systems for sustainable building applications. Energy Build 2011;43:3074e81. [4] Suamir IN, Tassou SA. Performance evaluation of integrated trigeneration and CO2 refrigeration systems. Appl Therm Eng 2013;50:1487e95. [5] Mancarella P, Chicco G. Assessment of the greenhouse gas emissions from cogeneration and trigeneration systems. Part II: analysis techniques and application cases. Energy 2008;33:418e30. [6] Ge YT, Tassou SA, Chaer I, Suguartha N. Performance evaluation of a trigeneration system with simulation and experiment. Appl Energy 2009;86: 2317e26.
[7] Godefroy J, Boukhanouf R, Riffat S. Design, testing and mathematical modelling of a small-scale CHP and cooling system (small CHP-ejector trigeneration). Appl Therm Eng 2007;27:68e77. [8] Piacentino A, Cardona F. An original multi-objective criterion for the design of small-scale polygeneration systems based on realistic operating conditions. Appl Therm Eng 2008;28:2391e404. [9] Lozano MA, Carvalho M, Serra LM. Operational strategy and marginal costs in simple trigeneration systems. Energy 2009;34:2001e8. [10] Schicktanz MD, Wapler J, Henning HM. Primary energy and economic analysis of combined heating, cooling and power systems. Energy 2011;36:575e85. [11] Roque Díaz P, Benito YR, Parise JAR. Thermoeconomic assessment of a multiengine, multi-heat-pump CCHP (combined cooling, heating and power generation) system: a case study. Energy 2010;35:3540e50. [12] Balli O, Aras H, Hepbasli A. Thermodynamic and thermoeconomic analyses of a trigeneration (TRIGEN) system with a gas-diesel engine: part I - methodology. Energy Convers Manag 2010;51:2252e9. [13] Balli O, Aras H, Hepbasli A. Thermodynamic and thermoeconomic analyses of a trigeneration (TRIGEN) system with a gasediesel engine: part II e an application. Energy Convers Manag 2010;51:2260e71. [14] Arcuri P, Florio G, Fragiacomo P. A mixed integer programming model for optimal design of trigeneration in a hospital complex. Energy 2007;32(8): 1430e47. [15] Carles Bruno J, Valero A, Coronas A. Performance analysis of combined microgas turbines and gas fired water/LiBr absorption chillers with postcombustion. Appl Therm Eng 2005;25:87e99. [16] Ho JC, Chua KJ, Chou SK. Performance study of a microturbine system for cogeneration application. Renew Energy 2004;29:1121e33. [17] Huicochea A, Rivera W, Gutierrez-Urueta G, Bruno JC, Coronas A. Thermodynamic analysis of a trigeneration system consisting of a micro gas turbine and a double effect absorption chiller. Appl Therm Eng 2011;31:3347e53. [18] Chen Q, Zheng J, Han W, Sui J, Jin H. Thermodynamic analysis of a combined cooling, heating and power system under part load condition. In: Proceedings of ECOS 2012 e the 25th international conference on efficiency, cost, optimization, simulation, and environmental impact of energy systems June 26e 29, 2012. Perugia, Italy. [19] Al-Sulaiman FA, Hamdullahpur F, Dincer I. Performance comparison of three trigeneration systems using organic rankine cycles. Energy 2011;36:5741e54. [20] Malico I, Carvalhinho AP, Tenreiro J. Design of a trigeneration system using a high-temperature fuel cell. Int J Energy Res 2009;33(2):144e51. [21] Wang Y, Huang Y, Chiremba E, Roskilly AP, Hewitt N, Ding Y, et al. An investigation of a household size trigeneration running with hydrogen. Appl Energy 2011;88:2176e82. [22] Lin L, Wang Y, Al-Shemmeri T, Ruxton T, Turner S, Zeng S, et al. An experimental investigation of a household size trigeneration. Appl Therm Eng 2007;27:576e85. [23] Martínez-Lera S, Ballester J. A novel method for the design of CHCP (combined heat, cooling and power) systems for buildings. Energy 2010;35:2972e84. [24] Lombardini Group. Available: www.lombardinigroup.it [accessed 29.06.13]. [25] BROAD Group. Available: www.broad.com [accessed 29.06.13]. [26] Software AVL List GMBH, BOOST v. 5.0, user manual: https://www.avl.com/ [accessed 07.10.13]. [27] Software CHEMCAD 6.5, user manual: http://www.chemstations.com/ Downloads/ [accessed 07.10.13]. [28] Heywood JB. Internal combustion engines fundamentals. McGrow-Hill; 1988. [29] Ferrari G. Motori a Combustione interna. Il Capitello; 1992. [30] Somers C, Mortazavi A, Hwang Y, Radermacher R, Rodgers P, Al-Hashimi S. Modeling water/lithium bromide absorption chillers in ASPEN Plus. Appl Energy 2011;88:4197e205. [31] Yin H, Qu M, Archer DH. Model based experimental performance analysis of a microscale LiBr-H2O steam-driven double-effect absorption chiller. Appl Therm Eng 2010;30:1741e50. [32] Horlock JH. Cogeneration: combined heat and power: thermodynamics and economics. Pergamon Press; 1987. [33] Feng X, Cai YN, Qian LL. A new performance criterion for cogeneration system. Energy Convers Manag 1998;39:1607e9. [34] Chicco G, Mancarella P. Trigeneration primary energy saving evaluation for energy planning and policy development. Energy Policy 2007;35:6132e44.
Please cite this article in press as: Jannelli E, et al., Thermodynamic performance assessment of a small size CCHP (combined cooling heating and power) system with numerical models, Energy (2013), http://dx.doi.org/10.1016/j.energy.2013.11.074
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Energy journal homepage: www.elsevier.com/locate/energy
Thermodynamic performance assessment of a small size CCHP (combined cooling heating and power) system with numerical models E. Jannelli a, M. Minutillo a, R. Cozzolino b, *, G. Falcucci a a b
University of Naples “Parthenope”, Centro Direzionale, Isola C4, 80143 Naples, Italy University of Niccolò Cusano, Via Don Carlo Gnocchi, Rome, Italy
a r t i c l e i n f o
a b s t r a c t
Article history: Received 9 July 2013 Received in revised form 30 October 2013 Accepted 25 November 2013 Available online xxx
The aim of the this work has been to evaluate the performance of a small-size CCHP (Combined Cooling Heating and Power) system based on the integration of 20 kW Lombardini diesel engine and a double effect water-LiBr absorption chiller. This integrated system has been designed to produce both hot water, by recovering heat from the engine cooling system, and chilled water, by recovering heat from the engine exhaust gasses (the exhaust gasses are sent to the fired-combustor of the absorption chiller). The analysis has been conducted by using numerical simulations: the engine and the absorption chiller have been modeled by means of 0e1D dimensional and thermochemical models, respectively, and the validation procedure has been performed by using the available operating data. The system performance has been calculated by introducing some performance parameters that have allowed: i) to estimate the efficiency of the primary energy conversion into useful energy EUF (energy utilization factor), ii) to consider the quality difference between cool/heat and work and among heats at different temperatures ExUF (exergy utilization factor); iii) to evaluate the primary energy saving with respect to the separate production of the same energy fluxes TPES (trigeneration primary energy saving). Ó 2013 Elsevier Ltd. All rights reserved.
Keywords: Polygeneration Diesel engine Absorption chiller Numerical models Performance coefficients
1. Introduction In recent years, the development of polygeneration systems has received an increasing attention in the area of small scale power systems for different applications ranging from residential to yachting utilities [1e5]. The residential sector is an important consumer of energy that could receive substantial benefits from the extensive utilization of trigeneration technologies. In fact, instead of satisfying electricity, heat and cooling demands separately, a combined production could provide considerable energy saving and an important emission reduction, as well. In the yachting sector the replacement of the traditional vapor compression units with the absorption machines may represent an interesting alternative that allows to reduce the on board energy consumption, thanks to the heat recovery on the exhaust gases and/or to the cooling circuits from internal combustion engine. The interest in small-size CCHP (Combined Cooling, Heating and Power) systems is high, due both to the energy saving obtained * Corresponding author. Tel.: þ39 0645678350. E-mail addresses: [email protected] (E. Jannelli), mariagiovanna. [email protected] (M. Minutillo), [email protected] (R. Cozzolino), [email protected] (G. Falcucci).
with the recovery of waste heat and to the ability in satisfying the energy demand of several types of stand-alone utilities. The research in designing, development and optimization of these technologies [6e8] and in the operational planning is considerably active (strategy concerning operational state of the equipment, energy flow rates, purchase/selling of electricity, etc.), according to energetic and economic issues [9e11]. In particular, the research is carried out by using both experimental and numerical approaches. In this paper a numerical analysis of a micro-scale trigeneration system, based on an internal combustion engine integrated with a water/LiBr double-effect absorption chiller, has been carried out by 0e1D dimensional and thermochemical models.
2. Background CCHP systems do not represent a new technology: large-scale systems from a few hundreds of kWs to MWs are intensively studied and illustrated in the technical literature. Some studies have examined the performance and the optimal operating strategies of trigeneration systems based on different prime movers such as internal combustion engine [12e14], gas turbine [15e18] and fuel cells [19,20].
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Nomenclature ABS A/F a C CCHP COND COPSP EVP EUF ExUF fPC FSP FCCHP h hBB hchill-in hchill-out HTG LTG LHV mBB mc _ chill m _f m _F m
absorber air to fuel ratio vibe parameter () cylinder combined cooling heating and power condenser coefficient of performance of an equivalent compression electric refrigerator group evaporator energy utilization factor () exergy utilization factor () post-combustion factor () total fuel energy input required for the separate production (kW) total fuel energy input to the trigeneration system (kW) heat transfer coefficient (J/m2 K) enthalpy of the blow-by (J/kg) enthalpy of entering chilled water (kJ/kg) enthalpy of leaving chilled water (kJ/kg) high temperature generator low temperature generator lower heating value mass flow rate of the blow-by (kg/s) mass flow rate inside the cylinder (kg/s) chilled mass flow rate (kg/s) fuel consumption of the chiller fired-combustor (kg/s) fuel consumption of the engine (kg/s)
Balli et al. [12,13] presented a methodology and relations for thermodynamic and thermoeconomic evaluation performance of a trigeneration system with a rated output power of 6.5 MW gasdiesel engine integrated with an absorption chiller. The proposed methodology includes the definition of performance parameters and the cost analysis formulation based on a specific exergy cost method. The results of their study can be beneficial to change the components that have low thermodynamic efficiencies and large exergy consumptions, allowing to regulate the sale price of the products and to review the plant’s economic policy. Bruno et al. [15] studied the integration of four microturbines (in the range 30e100 kWe) with a double effect direct-fired absorption chiller. The exhaust gases were directed to the chiller with the option of additional natural gas post-combustion. Thus, the authors analyzed the effect of the post-combustion degree on the trigeneration system performance and determined the working conditions that, in the proposed integrated system, allowed to obtain the maximum efficiency. Thus, the authors demonstrated that a directly driven absorption chiller with a post-combustion system can introduce advantages with respect to the more conventional single effect hot water system in terms of higher COP (coefficient of performance) and flexibility. This is due to the decoupling between the electricity and the chilled water production. The analysis was carried out by using a numerical model, developed with the software package EES (Engineering Equation Solver). Ho et al. [16] studied a cogeneration system that consists of a single effect commercial absorption chiller integrated with a Capstone microturbine (30 kW). Their results pointed out that the electric efficiency was 21% and the overall system efficiency was 46%. Huicochea et al. [17] studied the performance of a CCHP system based on a double effect absorption chiller driven by the exhaust
MP pc PC PES PL Pe Qc QF Qf QHTG Qt Qw SB Tg Ti TPES u V
measurement point pressure inside the cylinder (Pa) post-combustion primary energy saving () plenum engine electric power (kW) chiller cooling power (kW) total fuel heat entering to the engine (J) fuel power entering to the fired-combustor (kW) thermal power to the HTG (kW) thermal power recovered by the engine cooling system (kW) wall heat loss of the engine (J) system boundary temperature of the engine exhaust gas (K) temperature of the engine wall (K) trigeneration primary energy saving () specific internal energy (J/kg) cylinder volume (m3)
Greek letters a Crank angle ( ) a0 Crank angle at the start of combustion ( ) Dac combustion duration ( ) 4 and 4 parameters based on the temperature at which the heat is available () hSP electric reference efficiency for the separate e production () hSP thermal reference efficiency for the separate t production ()
gas of a 30 kWe microturbine. The authors developed a thermodynamic simulator by using Matlab programming language. The model was based on the mass, species and energy conservation equations for each component of the absorption chiller. The modeling allowed to predict the trigeneration performance at different operating conditions such as the ambient temperature and the microturbine mass flow rates. In particular, the decreasing tendency of all performance parameters (i.e. COP and electric efficiency) with the increase of ambient temperature was shown. The results pointed out that the proposed system for the co-production of electric, cooling and heating powers based on the microturbine technology represents an attractive solution in the fields of the distributed generation. Chen et al. [18] investigated behavior and performance during the off-design operation of a small-scale gas turbine (1747 kWe) coupled with a double effect chiller and a heat exchanger. The study was carried out by using a mathematical model, written in Fortran code and based on energy and exergy rate balances. The results showed that in off-design conditions, the performance of the gas turbine changed dramatically, whereas the performance of the chiller changed slightly. The calculated efficiency of the gas turbine ranged from 27 at full load to 11% at partial load, while the COP increased slightly with the decreasing of the load level of the CCHP system. Thus, the performance deterioration of the system was due to the bad performance of the gas turbine under the off-design conditions. Even if there are no distinctions between large-scale and microscale trigeneration systems and absorption chillers, the designing criteria and the operating conditions are different. Thus further and detailed experimental and numerical studies have to be conducted in order to allow the development and the optimization of these technologies, according to their configurations and applications.
Please cite this article in press as: Jannelli E, et al., Thermodynamic performance assessment of a small size CCHP (combined cooling heating and power) system with numerical models, Energy (2013), http://dx.doi.org/10.1016/j.energy.2013.11.074
E. Jannelli et al. / Energy xxx (2013) 1e10
Wang et al. [21] and Lin et al. [22] analyzed the performance of micro-scale trigeneration systems that use as prime mover an internal combustion engine. Wang et al. [21] examined a trigeneration system consisting of an engine fueled by hydrogen (6.5 kWe) whose waste heats, rejected from exhausts and engine cooling system, are used for domestic purpose (hot and cooling water). The authors simulated the trigeneration system by using the function of the energy and mass balance of the computational software called ECLIPSE. Their study showed that the hydrogen is a very interesting fuel that allows to achieve equal or better performance to the conventional diesel fuel in terms of energetic performance and near zero carbon emissions. Thus, the authors pointed out the enormous potential fuel savings and massive reductions in greenhouse gas emissions per unit of useful energy outputs with cogeneration and trigeneration compared with that of single generation. Lin et al. [22] studied and realized a small-scale trigeneration system based on 9.5 kW diesel engine coupled with both a heat recovery system (for heating purposes) and an absorption refrigerator (for cooling energy requirements). The experimental results, obtained by using their experimental apparatus, pointed out that if the engine load is over 50%, the exhaust gases are hot enough to drive the absorption refrigerator allowing very low temperature (the lowest evaporator temperature at the point in freezer of the refrigerator is 24.9 C; the temperature at the point in food compartment is 8.2 C while the refrigerator is powered by the heat from engine exhaust gas at the engine load of 75%). In the technical literature a limited number of studies deals with behavior and performance of CCHP systems in the power ranges less than 20 kWe because, in the past, their high costs limited the diffusion of these systems and consequently their designing and development; only in recent years the interest in these small-scale systems is increased, mainly as a result of the policies for the promotion of the DG (distributed generation). Some countries have implemented a legislation that offers economic bonuses to cogeneration/trigeneration systems so that, their development is become more profitable for companies that want to invest in the decentralized generation [23]. 3. System description The CCHP system, investigated in this paper, consists of a small size diesel engine and a double effect water/LieBr absorption chiller. The integration between these units is realized by introducing the engine exhaust gases into the fired-combustor of the high temperature generator of the chiller. Furthermore in order to satisfy the heat requirements, according to the cooling loads, fresh air and additional fuel can be added too. Waste heat from the engine cooling system is recovered in order to produce hot water (thermal power production). The main advantages of using a double effect absorption chiller driven with a fired-combustor system, over the conventional single-effect hot-water driven system, are both a higher COP and the decoupling between electric and cooling powers (wider ranges of operating conditions are allowed). Since only few commercial chillers are available in a small capacity range, the choice of the engine size is constrained both by the electric power demand and the cooling power requirements. The CCHP system proposed in this study is based on the integration of commercial units: the engine is a 20 kWmech Lombardini diesel engine [24] and the absorption chiller is a 23 kWc doubleeffect water/LiBr chiller commercialized by Broad [25]. The absorption system can work not only as chiller but also as heater.
3
Fig. 1. Interconnection in the outputeinput units models.
However, in this analysis the operation of the chiller as heater was not considered. 4. System modeling The CCHP system has been studied by using original models developed by the authors. In particular, the engine and the absorption chiller have been modeled by means of 0-1D dimensional and thermochemical models, respectively, and the validation procedure has been performed by using the available operating data. Fig. 1 sketches the interconnection in the outputeinput units models. 4.1. Engine simulation The behavior and the performance of the Lombardini diesel engine, whose characteristics are summarized in Table 1, have been determined by means 0e1D dimensional model, developed by the authors using AVL BOOST code [26]. The model describes the phenomena that take place inside the cylinder and the engine ducts; the operating parameters, such as exhaust gas mass flow, pressure and temperature, are computed at different engine speeds (the transient conditions due to the variation of the engine speed, are not taken into account). The equation, that describes the engine operation, is reported as follows:
dðmc $uÞ dV dQF X dQw dmBB ¼ pc $ þ hBB $ da da da da da
(1)
in which V is the cylinder volume, QF is the total fuel heat input, Qw is the wall heat loss, mc is the mass inside the cylinder, pc is the pressure inside the cylinder, u is the specific internal energy, a is the
Table 1 The main engine characteristics [24]. Cylinder Displacement Bore Stroke Compression ratio Rating kW Max torque Minimum idling speed Vol. of air required for the correct combustion @3600 rpm
n. cm3 mm mm N(80/1269/CEE) iso 1585 Nm m3/min
3 1028 75 77.6 22.8:1 20 67 @2000 900 63
Please cite this article in press as: Jannelli E, et al., Thermodynamic performance assessment of a small size CCHP (combined cooling heating and power) system with numerical models, Energy (2013), http://dx.doi.org/10.1016/j.energy.2013.11.074
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E. Jannelli et al. / Energy xxx (2013) 1e10
From the total heat supplied to the cycle, which is determined by the amount of fuel in the cylinder and by the A/F (Air/Fuel) ratio, the model calculates the actual heat input per degree crank angle. The combustion process is solved with a 0-dimensional model by using the standard single-Vibe approach (A/F ¼ 22 has been chosen for full load operating condition at different regimes). The Vibe function is: ðmþ1Þ dx a ¼ $ðm þ 1Þ$ym $ea$y Dac da
Fig. 2. Flowsheet of the engine model. PL: plenum, MP: measurement point, SB: system boundary, C: cylinder.
crank angle and hBB and mBB represent the enthalpy and mass of the blow-by, respectively. In order to solve this equation, models for the combustion process and the wall heat transfer, as well as the gas properties as a function of pressure, temperature and gas composition, are required.
(2)
with dx ¼ dQF/QF and y ¼ (a a0)/Dac, where a0 is the crank angle of the start of the combustion, Dac is the combustion duration and a is the Vibe parameter, which is 6.9 for complete combustion [26,28]. The heat transfer through the engine walls is simulated by means of the Woschni approach [28], by imposing the surface (exchange) areas of piston, cylinder head and valve openings. The heat is calculated as:
Qw ¼ h A Tg Ti
(3)
Fig. 3. Schematic layout of the double effect Water/LiBr absorption chiller.
Please cite this article in press as: Jannelli E, et al., Thermodynamic performance assessment of a small size CCHP (combined cooling heating and power) system with numerical models, Energy (2013), http://dx.doi.org/10.1016/j.energy.2013.11.074
E. Jannelli et al. / Energy xxx (2013) 1e10 Table 2 The main absorption chiller nominal operating data [25]. Chilled power Temperature of chilled water Chilled water flow rate Maximum gas consumption COP
kW C m3/h m3/h e
23 7e14 2.9 2.2 1.1
in which the heat transfer coefficient h is related to the main thermodynamics parameters of the engine [28,29], Tg and Ti are the gas temperature and the wall temperature, respectively. The flux inside pipes and connections is solved with a 1D approach considering the distributed pressure losses (the roughness coefficient is assumed equal to 0.019 [28]). An average cell size of 20 mm is chosen for the discretization of the 1D elements, i.e. the pipes. Fig. 2 depicts the flowsheet of the engine model. For internal mixture preparation it is assumed that: the fuel added to the cylinder charge is immediately burned; the combustion products mix instantaneously with the rest of the cylinder charge and form a uniform mixture; the A/F ratio of the charge diminishes continuously from a high value at the start of the combustion to the final value at the end of the combustion.
4.2. Absorption chiller simulation The selected 23 kWc double effect water/LiBr absorption chiller is a commercial unit that consists mainly of: a low pressure generator,
5
a high pressure generator, an absorber, an evaporator, a condenser, a pump, throttling valves and a cooling tower. Fig. 3 depicts its schematic layout and in Table 2 the main operating data are summarized. Since the absorption chiller plays an important role in recovering the waste heat, special attention has been devoted to study its performance at design and off-design conditions. The model of the absorption chiller is a thermochemical model developed by using the CHEMCAD code [27]. It is based on the ELECNRTL property method and the electrolyte wizard option has been used to generate a series of reactions (the relevant reaction is the association/dissociation of lithium bromide). The model is based on the numerical approach proposed by Somers et al. [30]. The assumptions are as follows: the analysis has been made under steady-state conditions the refrigerant (water) at outlet state of the condenser is saturated liquid the refrigerant (water) at outlet state of the evaporator is saturated vapor the pressure losses in pipelines and all heat exchanger are negligible the heat losses in the heat exchangers are negligible the expanding process in the throttling valves is isenthalpic. In Fig. 4, the flowsheet of the model is reported. The scheme consists of five main sections that simulate the HTG (high temperature generator), the LTG (low temperature generator), the COND (condenser), the EVP (evaporator) and the ABS (absorber), respectively. The cooling tower is not simulated because, in this first analysis, the power consumption of the chiller is not considered and the
Fig. 4. Flowsheet of the absorption chiller model.
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temperature of the cooling water can be varied as each input data (i.e. for a yachting utility the cooling tower is not necessary and the chiller is simplified in its layout). The refrigerant rich solution water/LiBr (water is the refrigerant and LiBr is the sorbent), that enters in the state (IN), is pumped (PUMP1) up to the maximum pressure of cycle and then it is split in two streams (2) and (5). The streams (2) and (5) are pre-heated in a heat exchangers (HEX4) and (H-EX6) by subtracting heat to streams (6) and (13) that are the LiBr rich solution throttled back to the absorber (ABS). The HTG has been simulated by using two components: a directfired combustor (COMB), that allows the heating of the refrigerant solution, and a flash separator (HT-SEP), in which the refrigerant vapor is separated and sent to the LTG. Similarly, the LTG has been simulated by using a heat exchanger (LT- HEX) and a further flash separator (LT-SEP). The primary refrigerant vapor (7) is cooled in the LTG allowing the vaporization of secondary refrigerant vapor (15) that is sent to the COND (condenser). The refrigerant liquid (18), after leaving the mixer (MIXER1), is throttled and sent to the EVP (evaporator) where the cooling effect is obtained. Thus, the refrigerant liquid, that is in the vapor phase, is absorbed by the LiBr rich solution (ABS). The circuit of the water/LiBr solution is closed, thus, the chemical and thermodynamic conditions of the outlet stream (OUT) are set equal to those of the input stream (IN). 4.3. Models validation The engine model has been validated by using the nominal data declared by the manufacturer [24]. In particular, the engine parameters, such as the heat transfer coefficient, the air/fuel ratio, the injection advance and the combustion duration, have been tuned in order to fit the power, the torque and the specific fuel consumption trends, as provided by manufacturer. Fig. 5 highlights the good agreement between the numerical results and the nominal data obtained with the tuned engine parameters. Table 3 shows the main engine operating data calculated by the model. The electric power has been determined assuming the efficiency of the electric generator equal to 0.97. The absorption chiller model validation has been carried out by analyzing some performance and operating parameters such as the
Fig. 5. Engine performance in terms of mechanical power and fuel consumption: comparison between calculated and nominal data [24].
Table 3 The engine operating conditions calculated by the model. RPM
Pmech [kW]
Pel [kW]
Fuel [kg/h]
Air [kg/h]
Exhaust gases mass flow [kg/h]
Exhaust gases temperature [K]
Qw [kW]
1500 2000 2400 2800 3200 3600
7.5 11.5 14.2 16.25 17.9 18.2
7.3 11.2 13.8 15.8 17.4 17.7
2 2.9 3.4 4 4.5 4.9
43 61.9 72.2 86 107 117
45 64.8 75.6 90 111.6 122.4
620 665 700 715 715 718
12.0 16.5 18.5 19.8 21.5 21.4
COP (coefficient of performance), the sorbent solution and the refrigerant mass flow rates. The COP values (thermal COP) have been calculated by applying the following equation:
COP ¼
_ ðh Qc hchillout Þ m ¼ chill chillin _ f LHV Qf m
(4)
in which, the numerator represents the cooling power (or cooling capacity) at the evaporator and the denominator is the fuel power input to the fired-combustor (based on the diesel Low Calorific Value). The data used for the model validation have been derived by the study conducted by Yin et al. [31], in which the 16 kWc double effect water/LiBr absorption chiller, produced by the same manufacturer, was modeled and tested. Thus, because the 16 kWc and 23 kWc absorption chillers are very similar in terms of behavior and performance, the author used the mentioned study as the reference case for the validation (experimental data on the 23 kWc were not available). Fig. 6 depicts the COP vs. the cooling load, whose variation was obtained by decreasing the chilled-water return temperature from 14 C to 8 C, while the chilled and cooling water flows and the cooling water inlet temperature were maintained constant. The numerical COP, calculated by the model, is compared to the numerical and the experimental COP of the reference case [31]. It can be noted that the trend and the values are very close (the deviation with respect to the experimental data ranges from 2.5% to
Fig. 6. Chiller performance (COP) under various load conditions: the comparison between the numerical results and the reference data [31].
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E. Jannelli et al. / Energy xxx (2013) 1e10
4.5%), thus the model can be considered reliable for the analyzed operating conditions. In Table 4, the state point results from the 23 kWc chiller model (state points are defined in Fig. 3) are summarized. Fig. 7 shows the flow rates of the dilute sorbent solution from the absorber to the HTG and LTG at different load conditions. It can be noted that the sorbent solution flow rate in the LTG decreases with the cooling load decreasing, while the sorbent solution flow rate in the HTG is relatively constant. This trend in the HTG depends on the control strategy of the absorption chiller that, by using a variable frequency solution pump, has to maintain the same flow rate of sorbent solution in the HTG. Thus, the ratio between the sorbent solution in the HTG and the total solution flow rate from the absorber (sorbent solution split ratio) varies between 0.5 at design load to 0.7 at minimum cooling load, as it is explained in Ref. [31]. Fig. 8 depicts the refrigerant flow rates vaporized in the evaporator at different load conditions. It is worth noting that the refrigerant flow rate is proportional to the cooling load and the HTG generates more refrigerant than the LTG. The trends of the refrigerant flow rates are in accordance with the results of the reference case, even if the values are higher because the size of the absorption chiller is bigger. 5. Results and discussions The developed models have been used both to define the operating range of the CCHP system in terms of electric, thermal and cooling powers and to forecast its performance. Some of the
Fig. 7. Sorbent solution flow rate under various cooling capacities: the comparison between the numerical results and the reference data [31].
output data of the engine model have been used as input data for the chiller model. In particular, in order to connect match the electric power to the temperature and mass flow rate of the exhausts gasses (input data for the chiller model) the following relationships have been determined:
Table 4 The state point results calculated by the model. Point General IN 1 2 5 8 9 14 16 18 19 20 WATERIN HTG 3 4 6 7 LTG 10 11 12 13 15 COND 17 25 WATEROUT EVP 21 22 CHIL-IN CHIL-OUT ABS 23 24 OUT
Mass flow (kg/s)
Temp. ( C)
_ Exhaust ¼ m
Press. (kPa)
LiBr (%)
0.28 0.28 0.14 0.14 0.132 0.14 0.134 0.008 0.014 0.266 0.266 0.829
35.92 35.92 35.92 35.95 42.0 35.98 42.0 40.47 40.47 42.01 42.01 30.0
0.799 91.99 91.99 91.99 7.59 7.59 7.59 7.59 7.59 7.59 0.799 100
57.79 57.79 57.79 57.79 61.12 57.79 61.12
0.14 0.14 0.132 0.008
138.73 154.89 154.89 154.89
91.99 91.99 91.99 91.99
57.79 57.79 61.12
0.14 0.008 0.14 0.134 0.006
75.87 97.31 85.25 85.24 85.24
7.59 91.99 7.59 7.59 7.59
57.79
0.006 0.829 0.829
40.47 38.76 42.10
7.59 200 200
0.014 0.014 0.805 0.805
3.68 3.68 14.0 7.0
0.799 0.799 78.45 78.45
0.28 0.829 0.28
40.54 30.0 36.26
0.799 200 0.799
H2O (%)
100
0 0 0 0 0 0 0 0.098 0.055 0 0 0
100
0 0.082 0 1
100 100 61.12 61.12
100 57.79 61.12 61.12
TExhaust ¼
0:15$Pe3 5:08$Pe2 þ ð60:53$Pe Þ 184 3600
0:14$Pe3 þ 4:84$Pe2 ð40:43$Pe Þ þ 439
(5)
(6)
5.1. The CCHP system operating range The integration between the sub-units of the CCHP system has been realized by recovering the waste heat from the engine exhaust
0 0 0.065 0 0
100 100 100
0 0 0
100 100 100 100
0.1 0.8 0 0
100
0.059 0 0
57.79 57.79
Vapor frac. ()
7
Fig. 8. Refrigerant flow rate under various cooling capacities: the comparison between the numerical results and the reference data [31].
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E. Jannelli et al. / Energy xxx (2013) 1e10
gasses. These gases are recirculated to the fired-combustor of the absorption chiller where, if the available thermal power does not allow to satisfy the energy demands of the HTG (high temperature generator), additional fuel is used (the post-combustion of the exhaust gasses is realized). Furthermore, in order to improve the energy content in the HTG and, consequently, to increase the cooling load, additional air can be used. The presence of the fired-combustor allows to decouple the cooling power from the electric power. Furthermore, in order to satisfy the thermal demands, the waste heat from the engine cooling system is recovered by producing hot water. The proposed and studied CCHP system is able to operate in the following operating ranges: electric power: 7e18 kWe thermal power (hot water production): 10e17 kWt cooling power: 4e23 kWc Fig. 9 depicts the CCHP system operating field in terms of electric power vs. cooling power. In the figure, it is possible to identify 3 operating areas: - “area 234” (without PC zone), in which the waste heat from the engine exhaust gases is used as thermal input in the HTG; - “area 1245” (PC zone), in which the engine exhaust gases are used as oxidant in the fired-combustor of the chiller in order to provide additional heat to the HTG (the minimum oxygen concentration of 5% is assured in the exhaust gasses) for increasing the cooling capacity; - “area 1567” (PC þ Air zone), in which fresh air is provided to the fired-combustor in order to increase the cooling capacity until to the nominal value (23 kWc).
fpc ¼
_f m _F m
(7)
_ f and m _ F are the fuel consumptions of the chiller firedwhere m combustor and of the engine respectively. The post-combustion factor equal to 0 indicates that the integrated system does not use additional fuel into the fired-combustor of the absorption chiller. Fig. 10 shows the post-combustion factor vs. the electric power according to the operating field illustrated in Fig. 9. It can be noted that: - if the engine exhaust gasses are used as oxidant in the chiller fired-combustor (PC curve), fpc remains relatively constant, because the additional fuel is much lower compared to the engine fuel flow; - if both fresh air and fuel are added to the engine exhaust gasses (PC þ Air curve), the fpc curve decreases with the increase of the electric power, because more heat has to be provided at low electric power, in order to produce 23 kWc. The parameters introduced for evaluating the performance of CCHP system are the EUF (Energy Utilization Factor) [32] and the ExUF (Exergy Utilization Factor). The first one takes into account the efficiency in the conversion of the primary energy into work and heat (or cool). However, this performance coefficient does not discriminate between the electric power and the thermal and/or cooling power, while ExUF considers the quality difference between heat and work and among heats at different temperatures (heat at high temperature in the HTG absorption chiller and heat at low temperature from the engine cooling system) [33]. These parameters are defined as follows:
EUFCCHP ¼
Pe þ Qt þ Qc _ f LHV _ F LHV þ m m
(8)
5.2. The CCHP system performance
ExUFCCHP ¼
Pe þ Qt 4 þ fQHTG _ f LHV _ F LHV þ m m
(9)
In the paper the analysis of the CCHP system behavior has been carried out by using some operating and performance parameters. In order to take into account the additional fuel used into the fired-combustor with respect to the engine fuel consumption, the parameter fpc (post-combustion factor) has been defined as:
where Pe is the engine electric power (the auxiliaries power consumption of the absorption chiller has not been considered), Qt is the thermal power recovered by the engine cooling system
Fig. 9. CCHP system operating field in terms of electric power and cooling power.
Fig. 10. Post-combustion factor as function of the engine the electric power.
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E. Jannelli et al. / Energy xxx (2013) 1e10
9
(Qt ¼ ε$Qw where ε is the thermal recovery efficiency of the engine cooling system), QHTG is the thermal power to the HTG (the thermal input to the absorption chiller), 4 and f are parameters whose value changes between 0 and 1. These parameters are calculated by applying the following equations:
4 ¼
f¼
1
1
Ta
! (10)
TxðecsÞ
Ta TxðHTGÞ
! (11)
where Ta is the ambient temperature (reference temperature), Tx(ecs) and Tx(HTG) are the temperatures at which the heats for cogeneration and trigeneration purposes, are available respectively. These temperatures are calculated as the logarithmic mean temperatures between the inlet and the outlet temperatures of the heating sources (the temperatures at the engine cooling system and the temperatures at the HTG). The calculated values of 4 and f are 0.17 and 0.4 respectively. In Fig. 11 the EUF and the electric efficiency are shown as function of electric power. It can be noted that when the contribution of the post-combustion increases, EUF increases; as a matter of fact the highest EUF values are obtained when the exhaust gases and the additional air are mixed and used as combustion air (PC þ Air curve) to the fired-combustor. This allows to provide the thermal energy required to the absorption chiller for producing 23 kWc. EUF, calculated at maximum electric power and without the post-combustion, is equal to 0.78. Furthermore, it can be noted that the trend of EUF is in accordance with the trend of the electric efficiency. Fig. 12 depicts the trend of ExUF as function of the electric power. It is worth noting that the highest ExUF values are achieved without the post-combustion of the exhaust gasses (without PC curve). This behavior depends on the exergy destruction that occurs when the fired-combustor is used to increase the cooling power availability. As a matter of fact, its lowest value (0.38) is obtained at 7 kWe and 23 kWc with a post-combustor factor equal to 0.45. Furthermore, in order to evaluate the fuel saving that can be achieved with the co-production of electric, thermal and cooling powers, a new generalized performance indicator, called TPES (Trigeneration Primary Energy Saving), has been calculated, as
Fig. 12. System performance: the ExUF (Exergy Utilization Factor) as a function of electric power.
proposed in Ref. [34]. In fact, instead to calculated the PES coefficient, according to the European Directive 2004/08/EC, the coefficient TPES takes into account the cooling power explicitly as well as the reference efficiencies for comparison with the separate energy production. This parameter is defined as:
TPES ¼
F SP F CCHP F CCHP ¼ 1 SP F SP SP Pe =h þ Qt =h þ Qc = hSP $COPSP e
e
t
(12) where FCCHP and FSP are the total fuel energy input to the trigeneration system and the total fuel energy input required for the separate production of the same energy vectors (work, heat and cool) respectively. Furthermore in the eq. (12), the coefficients hSP e , SP hSP are the separate production reference efficiencies that t , COP have been assumed equal to 0.4, 0.9 and 4 (COP of an equivalent compression electric refrigerator group) respectively. These values are the average efficiencies of technologies used for the separate production of the same energy fluxes, as suggested in Ref. [34]. Instead of calculating the TPES for a assigned case study, (it is necessary to know the electric, thermal and cooling energy demands) in this analysis it has been chosen to estimate the TPES in the main working points of the CCHP system operating field, (Fig. 9). Thus, the analysis has been carried out considering the power values and not the annual energy values. In Table 5 the calculated data are illustrated. It can be seen that the TPES ranges from 0.03 to 0.21. The negative values are obtained at low electric power (7 kW) and when the fuel consumption in the fired-combustor is high. Thus, the CCHP system has good performance in a wide operating range, but the better performance can be achieved when the Table 5 The TPES (trigeneration primary energy saving) values calculated in the operating points illustrated in Fig. 9.
Fig. 11. System performance: the EUF (Energy Utilization Factor) as a function of electric power.
Points
1
2
3
4
5
6
7
FCCHP [kW] Pe [kW] Qt [kW] Qc [kW] TPES
27.9 7.3 9.6 3.8 0.11
35.4 11.2 13.2 3.8 0.21
58.8 17.7 17.1 3.8 0.10
58.8 17.7 17.1 11.4 0.16
68.5 17.7 17.1 21.9 0.11
69.5 17.7 17.1 23 0.10
44.79 7.3 9.6 23 0.03
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E. Jannelli et al. / Energy xxx (2013) 1e10
cooling power demand is low so that the additional fuel consumption in the fired-combustor is not necessary or is low. 6. Conclusion In this paper the performance of a micro-scale CCHP system, that consists of a diesel engine and a double effect water/LiBr absorption chiller, has been evaluated by means of numerical simulations. The engine and the absorption chiller have been modeled by using 0-1D dimensional and thermochemical codes respectively. The models, validated by using the available operating data, have been integrated in order to evaluate the operating field of the CCHP system in terms of electric, thermal and cooling powers. The performance analysis has been carried by introducing some performance coefficients such as: EUF, ExUF and TPES. The numerical results have highlight that the CCHP system has good performance in a wide operating range. However the better performance are achieved when the cooling power demand is low so that the additional fuel consumption in the fired-combustor is not necessary (this means high values of ExUF and TPES) or is low. Finally, in analyzing the operation of this integrated system, some practical considerations must be taken into account, such as the maximum mass flow rate that can be treated in the firedcombustor according to its designing conditions. This constraint can involve a modification to the fired combustor increasing the costs of the system and thus its application as micro-scale energy conversion system in grid-connected or stand-alone utilities. The paper has highlight that the numerical models can be useful tools in designing and optimization of energy conversion systems allowing to investigate on their behavior in different operating points and thus to assist to the experimental activities. Acknowledgments The authors would like to thanks the laboratory of energy systems of University of Cassino e Southern Lazio for the possibility of using their licensed code AVL BOOTS. References [1] Wu DW, Wang RZ. Combined cooling, heating and power: a review. Prog Energy Combust Sci 2006;32:459e95. [2] Fabrizio E. Feasibility of polygeneration in energy supply systems for healthcare facilities under the Italian climate and boundary conditions. Energy Sustain Dev 2011;15:92e103. [3] Bingöl E, Kılkıs B, Eralp C. Exergy based performance analysis of high efficiency poly-generation systems for sustainable building applications. Energy Build 2011;43:3074e81. [4] Suamir IN, Tassou SA. Performance evaluation of integrated trigeneration and CO2 refrigeration systems. Appl Therm Eng 2013;50:1487e95. [5] Mancarella P, Chicco G. Assessment of the greenhouse gas emissions from cogeneration and trigeneration systems. Part II: analysis techniques and application cases. Energy 2008;33:418e30. [6] Ge YT, Tassou SA, Chaer I, Suguartha N. Performance evaluation of a trigeneration system with simulation and experiment. Appl Energy 2009;86: 2317e26.
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Please cite this article in press as: Jannelli E, et al., Thermodynamic performance assessment of a small size CCHP (combined cooling heating and power) system with numerical models, Energy (2013), http://dx.doi.org/10.1016/j.energy.2013.11.074