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Novel geothermal driven CCHP systems integrating ejector transcritical CO2 and Rankine cycles: Thermodynamic modeling and parametric study
Energy Conversion and Management 205 (2020) 112396
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Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman
Novel geothermal driven CCHP systems integrating ejector transcritical CO2 and Rankine cycles: Thermodynamic modeling and parametric study
T
⁎
V. Zare , H. Rostamnejad Takleh Faculty of Mechanical Engineering, Urmia University of Technology, Urmia, Iran
A R T I C LE I N FO
A B S T R A C T
Keywords: CCHP Ejector Transcritical CO2 Geothermal energy Exergy analysis
With recent technical advances, Combined Cooling, Heating and Power (CCHP) system offers a practical and economical alternative to conventional energy conversion systems. Benefits of these systems are more marked with employing efficient thermodynamic cycle and using renewable energy resources. In this regard in the present work, a novel geothermal driven CCHP system, in which the ejector transcritical CO2 cycle is integrated with conventional Rankine cycle, is proposed (system a). The proposed system is then modified by replacing the gas cooler with an internal heat exchanger to make a more efficient CCHP system (system b). To investigate the first and second law performance of the proposed systems, thermodynamic models are developed and parametric analysis is carried out to examine the influences of design variables. The results indicated that, the gas cooler of conventional system can be replaced by an internal heat exchanger for a wide range of practical operating conditions and this replacement can improve the exergy efficiency, net output power and output cooling respectively by 30.9%, 49.1% and 75.8% at the expense of 39.1% reduction in heating output. Also a comparison is made between the two proposed systems in this work with similar systems (based on ejector transcritical CO2 cycle) proposed previously by other researchers and superiority of proposed systems in this paper is revealed and discussed.
1. Introduction The environmental problems and resource shortages of fossil fuels brings the necessity of employing high-efficiency energy conversion systems and renewable energy sources. Combined Cooling, Heating and Power (CCHP) concept is a promising technology to improve energy conversion efficiency by which the power, cooling and heating are produced simultaneously in a single unit. Such systems have many environmental and economic advantages over the separate production of power, heat and cold [1]. In conventional CCHP systems fossil fuels are usually utilized as the primary energy source [2]. In these systems prime movers such as steam turbines [3], gas turbines [4], stirling engines [5], fuel cells [6] and internal combustion engines [7] are employed. However, recently researchers have paid more attention on renewable energy-based CCHP systems. More recently, some research works are devoted to investigate Rankine or Kalina cycle-based micro CCHP systems using energy and exergy analyses [8,9]. Parikhani et al. [10] proposed a new CCHP system based on a modified version of Kalina cycle and analyzed its feasibility based on thermodynamic and thermoeconomic analyses. They reported an exergy efficiency and overall unit product cost of
⁎
27.68% and 198.3 $/GJ for the system. Novel integrated cogeneration and trigeneration configurations based on carbon dioxide vapor compression systems are proposed and analysed by Mohammadi and Powell [11], who investigated the systems for a 1000 kW capacity and evaporator temperatures of −35 °C to −45 °C. Chang et al. [12] proposed and evaluated two micro CCHP systems based on Rankine and vapor compression cycles and a high temperature PEMFC. They obtained the values of 49.7% and 47.4% for exergy efficiency of their proposed systems. In scientific literature, very few research works are available on geothermal driven CCHP systems. Mosaffa and Garousi [13] proposed a novel geothermal-based CCHP system consisting of an Organic Rankine Cycle (ORC), an ejector refrigeration cycle and a domestic water heater. They considered four different working fluids and found that the best performance of the system is achieved with R123. Another geothermalbased CCHP system is proposed by Mohammadi and Mehrpooya [14], who reported an exergy efficiency of 38.10% at a typical operating condition for a geothermal water temperature of 230 °C. Their proposed system employs the Kalina cycle as the Power Generation Unit (PGU) and it is also capable of producing some fresh water by reverse osmosis. Zare [15] made a comparison between two geothermal driven CCHP
Corresponding author. E-mail address: [email protected] (V. Zare).
https://doi.org/10.1016/j.enconman.2019.112396 Received 28 August 2019; Received in revised form 15 November 2019; Accepted 8 December 2019 0196-8904/ © 2019 Elsevier Ltd. All rights reserved.
Energy Conversion and Management 205 (2020) 112396
V. Zare and H. Rostamnejad Takleh
Nomenclature C Ė h ṁ P Q̇ s T Ẇ
e IHE is H ms n net out PGU pu T TRCC tot
Velocity (m/s) Exergy rate (kW) Enthalpy (kJ/kg) Mass flow rate (kg/s) Pressure (MPa) Heat transfer rate (kW) Specific entropy (kJ/kg.K) Temperature (°C) Power (kW)
Subscripts and abbreviations 0 b c d D CCHP ej EV
Environmental state boiler Condenser Diffuser Destruction Combined cooling, heating and power Ejector Expansion valve
evaporator Internal heat exchanger Isentropic process heater Mixing section Nozzle section Net Outlet Power Generation Unit Pump Turbine transcritical CO2 Total
Greek symbols
ηen ηex ηpu ηT µ
First law efficiency Second law efficiency Pump isentropic efficiency Turbine isentropic efficiency Entrainment ratio
considered as a favored choice with lots of proven advantages. CO2 as a working fluid in energy conversion systems has unique characteristics, such as low viscosity, high heat transfer coefficient, non-toxicity and inflammability. Also it has zero ozone depletion potential and negligible global warming potential beside a very low cost. Despite these favorable specifications of CO2, just a few studies can be found in open literature in the context of CCHP systems employing CO2 as the working fluid, however, a large number of studies are conducted on investigation of the ejector-expansion transcritical CO2 refrigeration cycle [18–21]. The first proposal of a CCHP system based on TRCC cycle was made by Wang et al. [22] in 2012. In their proposed system a TRCC cycle is integrated with a Brayton power cycle for trigeneration of cooling, heating and power. The cycle is supposed to be driven by the solar energy with a temperature of 230 °C. They found that, under a typical operating condition, thermal and exergy efficiencies of the system are
systems with ORC and Kalina cycles as the PGUs and LiBr/H2O absorption chiller for cooling production. His results showed that, the system with Kalina cycle generates 12.2% more power than the other system under the optimized operating conditions. A novel CCHP system driven by geothermal energy based on absorption chiller with ammonia-water mixture and an LNG power generation sub-system is developed by Ghaebi et al. [16]. They found that, their system (For a geothermal water temperature of 398 K and mass flow rate of 10 kg/s) yields net power output of 405.1 kW and thermal and exergy efficiencies of 85.92%, 18.52%, respectively. The increasing interest for geothermal driven CCHP systems can be maintained via introducing novel systems with higher efficiency and lower environmental impacts. Also, since the cold production is more complicated than the heat generation in CCHP systems, more attention should be paid on efficient cooling production [17]. Regarding these aspects, transcritical CO2 (TRCC) cycle equipped with an ejector can be
Fig. 1. Schematic diagram of the proposed CCHP system based on TRCC cycle by Wang et al. [22]. 2
Energy Conversion and Management 205 (2020) 112396
V. Zare and H. Rostamnejad Takleh
of proposed systems and conducting a parametric analysis to show their feasibility 4. Making a comparison between the proposed systems in this paper with the two similar previously proposed systems
53.0 and 28.8%. Xu et al. [23] modified the proposed system by Wang et al. [22] by replacing the turbine with an extraction one in order to provide relatively high pressure stream as the primary flow for the ejector, as a result of which the cooling output can be increased. They showed that, their proposed system yields energy utilization factor and exergy efficiency of 70.13 and 25.67% at a typical operating condition. In recent two proposed CCHP systems based on TRCC cycle, a technical mistake about the power generation is concerned. In the first system proposed by Wang et al. [22], the net output power ̇ = WTurbine ̇ ̇ − WCompressor (Wnet ) is close to zero in a wide range of operating conditions and for some operating conditions is even negative (as illustrated in Figs. 3–6 of Ref. [22]) which means that the system cannot generate output power and needs input power for operation. Also, in the second proposed system by Xu et al. [23] it is seen that the net output power of the system is negative in almost all the range of investigated operating conditions (as can be observed from Figs. 4–7 in Ref. [23] the consumed power by compressor is more than the produced power by the turbine). Considering the above discussion, in the present work a novel CCHP system based on ejector transcritical CO2 cycle is proposed and analyzed. The proposed system is an integration of TRCC and Rankine cycles which is capable of producing satisfactory amount of net output power beside the cooling and heating production. Also, a modified configuration of the proposed system in this work is suggested in which the gas cooler of TRCC cycle is replaced by an Internal Heat Exchanger (IHE) to enhance the system performance. For the two proposed systems, comprehensive energy and exergy analyses are conducted and a parametric study is carried out to investigate the effects of key parameters on the system performance. Also, performance of the two systems is compared to evaluate the influence of an IHE employment. The contribution, novelty and objectives of the present paper could be summarized as follows:
2. System description Before presenting the proposed systems in this work, it would be appropriate to present a brief discussion on the two previously proposed systems in the literature. Fig. 1 illustrates the CCHP system based on TRCC cycle proposed by Wang et al. [22]. As can be seen, the system is an integration of ejector-expansion TRCC refrigeration cycle and a Brayton power cycle. In this CCHP system the cooling, heating and power outputs are respectively produced by evaporator, heater and turbine. However, as mention before and reported by the authors, in this system the consumed power by the compressor is almost equal to (or even more than) the generated power by turbine. Therefore the system cannot be regarded as a CCHP system in practical applications. The modified CCHP system based on TRCC cycle proposed by Xu et al. [23] is illustrated in Fig. 2. As can be seen in this system the turbine is replaced by an extraction one by which a relatively high pressure stream is provided for the ejector as its primary flow to increase the cooling production. However, less attention is paid to the power production as the generated power by the turbine is less than the consumed power by compressor. Therefore, this system cannot be considered as a CCHP one in practical applications. In the present paper to solve the net power generation problem in two above-mentioned systems, the Brayton sub-system is proposed to be replaced by a Rankine one (as illustrated in Fig. 3), so that the compressor is replaced by a pump. To do so, a condenser is required after the ejector to convert the two-phase flow exiting from the ejector to a saturated liquid. Referring to Fig. 1, the high temperature supercritical CO2 (stream 1) enters the turbine where it is expanded to a lower pressure and generates power. The leaving CO2 from turbine (stream 2) enters to a heater to provide heating effect for the heat user, after which it would reject heat in the gas cooler (stream 3). To drive the ejector, exiting stream from the gas cooler (stream 4) enters to the ejector as the primary flow, so that the secondary vapor from the
1. Proposal of a novel geothermal driven CCHP system as an integration of TRCC and Rankine cycles 2. Replacing the gas cooler of TRCC cycle by an IHE to enhance the proposed CCHP system performance 3. A comprehensive evaluation of the first and second law performance
Fig. 2. Schematic diagram of the proposed CCHP system based on TRCC cycle by Xu et al. [23]. 3
Energy Conversion and Management 205 (2020) 112396
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Fig. 3. Schematic diagram of the proposed CCHP system based on TRCC cycle in this paper (system a).
be saturated. 6. Appropriate values for isentropic efficiencies of turbine and pump are assumed as given in Table 2. 7. In ejector modeling, the effects of mixing and frictional losses through the ejector are taken into account by appropriate values for nozzle efficiency, diffuser efficiency and mixing efficiency as listed in Table 2. 8. In ejector modeling, constant-pressure mixing model is considered.
evaporator (stream 13) is entrained. The exiting CO2 from ejector, which is a two-phase stream, (stream 7) flows to a condenser to reject heat and to be condensed to saturated liquid. Then stream 8, flows into a separator where it is divided into two streams, the first one (stream 11) flows to the evaporator after being expanded in the valve and the second one (stream 9) goes to the boiler after being pumped. To improve the performance of proposed system in this work, one more modification is made regarding the working conditions and temperature levels of entering and exiting streams from the gas cooler. In this configuration (as illustrated in Fig. 4) the gas cooler is proposed to be replaced by an Internal Heat Exchanger (IHE). In this system the rejected heat from the gas cooler, as a wasted one, is recovered by IHE as a result of which the working fluid enters the boiler (stream 10a) with a higher temperature. The higher temperature of the entering working fluid to the boiler results in a higher efficiency of the cycle. To model the proposed CCHP systems some assumptions are made as follows [22,23]:
3. Thermodynamic modeling Thermodynamic modeling of the proposed CCHP systems in this work is developed using the Engineering Equation Solver (EES) software. Each component of the systems is treated as a control volume for which the mass and energy conservations is applied. In modeling the considered CCHP systems, more attention is paid to ejector model, since it is a key component of the systems and plays important role to provide cooling. The modeling procedure of the ejector is described below.
1. Systems are working under steady state conditions. 2. There is no heat loss from the system components to the environment and no pressure drops in the pipelines. 3. Changes in kinetic and potential exergies and energies are ignored. 4. The isenthalpic process is considered for the throttling valve. 5. The streams at the evaporator and condenser outlets are assumed to
3.1. Ejector modeling Schematics of the ejector structure is illustrated in Fig. 5 which is comprised of four sections: suction, mixing, constant-area and diffuser sections. It employs a converging–diverging nozzle and converts the 4
Energy Conversion and Management 205 (2020) 112396
V. Zare and H. Rostamnejad Takleh
Fig. 4. Schematic diagram of the proposed CCHP system based on TRCC with an IHE (system b) and its T-s diagram.
constant-pressure model is preferred to the constant-area model for its simplicity and accuracy [24,25], so in this paper the constant-pressure method is selected. In order to simulate the ejector, as a component of the CCHP systems under consideration, presented procedure by Li et al. [26] is applied. The ejector modeling is started by determining thermodynamic states of the primary and secondary stream of the ejector. These state points, for considered system in this work as illustrated in Fig. 3, can be
mechanical energy of pressure type to kinetic energy, as a result of which a low-pressure region is created and a flow stream is entrained by suction. The occurring processes in the ejector include: expanding the high-pressure primary flow through the nozzle, mixing with the lowpressure secondary flow in mixing section and diffusing via the diffuser at the ejector outlet. Regarding the mixing process in the mixing section, two methods can be used in ejector modeling: 1) constant-pressure mixing and 2) constant-area mixing. In previous research works,
5
Energy Conversion and Management 205 (2020) 112396
V. Zare and H. Rostamnejad Takleh
Fig. 5. Ejector schematics and operation principle.
The exergy balance equation for the components of the systems, is expressed as follows to evaluate the exergy destruction/loss within the component:
determined noting that the gas cooler and evaporator temperature and pressure are known. Then, for entrainment ratio, an initial value is guessed while known values are considered for nozzle efficiency and diffuser efficiency. The relations used for modeling the ejector are given below, respectively for motive nozzle, mixing section and diffuser [25,27]: For motive nozzle:
ηn =
h4 − h5 , h4 − h5, is
h5, is = h (P5, s4 ),
C5 =
2(h4 − h5)
EḊ =
C ηm ⎜⎛ 5 ⎟⎞, 1 ⎝ + μ⎠
h6 =
C2 h4 + μh13 − 6 1+μ 2
(1)
(2)
For the diffuser:
h 7 = h6 +
C62 , 2
ηd =
ηth =
̇ ̇ ̇ + Qevaporator + Qheater Wnet ̇ Qin
(8)
ηex =
̇ + Ecooling ̇ ̇ Wnet + Eheating ̇ Ein
(9)
h7, is − h6 Ấ
(3)
h 7 − h6 Ấ
The obtained values for h 7 and h 7 from Eq. (3) will be compared with each other. If these two values would be close together with an acceptable accuracy (0.1 in this paper) the iterative calculations would be finished. Fig. 6 shows the flowchart of ejector modeling procedure. 3.2. Energy and exergy analyses To perform the energy analysis on a thermodynamic system, the mass conservation and first law principal would be applied to the system components. For a control volume accompanying a steady state process, the conservation of mass and energy can be expressed as:
∑ ṁ in = ∑ ṁ out
(4)
̇ − Wcv ̇ = Qcv
(5)
∑ ṁ out (hout ) − ∑ ṁ in (hin)
In addition to the energy analysis, the exergy analysis based on the second law principal is an important technique to assess the performance of energy systems. It reveals the location, type and true magnitude of inefficiencies in a thermodynamic cycle and provides a better understanding of the processes. Thus, it plays a key role in providing guidelines for designing efficient systems. Exergy, as the maximum attainable work from a system/fluid stream, can be divided into four parts, two of which (kinetic and potential terms) are usually negligible. Also, in systems with no chemical reactions and composition changes, chemical exergy term is not considered as it would be canceled out in exergy balances. Thus, for the CCHP systems in this work, exergy rate of the fluid streams can be expressed as [28]:
̇ = ṁ [(h − h 0) − T0 (s − s0)] E ̇ = Eph
(7)
̇ denotes for the overall input and output ̇ and ∑ Eout in which, ∑ Ein exergies to that component. To simulate the systems’ performance, the relations given in Table 1, along with the equations for ejector modeling as explained above, are introduced into the computer program. For performance evaluation of the proposed CCHP systems from the viewpoint of first and second laws, thermal and exergy efficiency is defined as follows [23]:
For mixing section with a proper guess value for the entrainment ratio:
C6 =
̇ ∑ Eiṅ − ∑ Eout
(6)
Fig. 6. Ejector modeling flowchart. 6
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Table 1 Energy- and exergy-related equations applied to analyse the system performance (system b). Component
Energy equation
Exergy balance equation
Condenser
ṁ 7 (h7 − h8) = ṁ 17 (h18 − h17)
Evaporator
c
ṁ 12 (h13 − h12) = ṁ 19 (h19 − h20)
̇ − E17 ̇ ) EḊ = (E7̇ − E8̇ ) − (E18 e ̇ − E13 ̇ ) − (E20 ̇ − E19 ̇ ) EḊ = (E12
Boiler
ṁ 1 (h1 − h10a) = ṁ 15 (h15 − h16)
̇ − E16 ̇ ) − (E1̇ − E10 ̇ a) EḊ = (E15
Turbine
ẆT = ṁ 1 (h1 − h2) ηis, T = (h1 − h2)/(h1 − h2, is )
T EḊ = (E1̇ − E2̇ ) − ẆT
Ejector
B
ej
̇ + E4̇ ) − E7̇ EḊ = (E13 pu ̇ − E9̇ ) EḊ = Ẇ pu − (E10
Pump
Ẇ pu = ṁ 10 (h10 − h9)
Expansion valve
h11 = h12
̇ − E12 ̇ EḊ = E11
Heater
ṁ 2 (h2 − h3) = ṁ 21 (h22 − h21)
̇ − E21 ̇ ) EḊ = (E2̇ − E3̇ ) − (E22
IHE
ṁ 3 (h3 − h4 ) = ṁ 10 (h10a − h10)
EḊ
ηis, pu = (h10 − h9, is )/(h10 − h9) EV
̇ indicates net output power and input energy ̇ and Qin In Eq. (8), Wnet to the system, respectively as: ̇ = WTurbine ̇ Wnet − ẆPump
(10)
̇ = Qboiler ̇ Qin = ṁ 15 (h15 − h16)
(11)
(12)
̇ ̇ − E21 ̇ Eheating = E22
(13)
IHE
̇ a − E10 ̇ ) = (E3̇ − E4̇ ) − (E10
geothermal heat sources, another design criterion for designing such systems is considered, that is designing and calculating the power, heating and cooling outputs for a given value of heat source mass flow rate. In the present paper this criterion is considered and the proposed systems performances are investigated for a unit mass flow rate of geothermal brine with a given temperature. For a base case operating condition (as given in Table 2), the properties of thermodynamic state points, according to stream numbers given in Fig. 4, are outlined in Table 3. Also, the values of performance parameters are given in Table 4. ̇ , Qė and The figures given in Table 4 reveals that, the values of Wnet ηex for system b are respectively, 49.1%, 75.8% and 30.9% higher than the corresponding values for system a. However, the values of Q̇H and ηth for system b are respectively 39.1% and 9.2% lower than those for system a. As the first important design variable is considered the boiler (turbine inlet) pressure, effects of which are illustrated in Fig. 8(a–e). As an important and general outcome in comparing performance of the two proposed systems, Fig. 8(a–e) reveals that employing the IHE (system b) ̇ , Qė and instead of a gas cooler (system a) would result in a higher Wnet ηex at the expense of decreasing Q̇H and ηth . Referring to Fig. 8(b–d) it can be seen that, for all the values of turbine inlet pressure, system b has significantly higher exergy efficiency and power and cooling outputs than system a, while its heating output is lower as presented by Fig. 8(e). Also, Fig. 8(a) shows that, system b has lower values of thermal efficiency than system a at higher turbine inlet pressures than 11 MPa. The better performance of system ̇ , Qė and ηex can be attributed to b compared to system a, in terms of Wnet the employment of an IHE instead of the gas cooler, as a result of which the rejected heat is recovered via the IHE. This internal heat recovery
̇ ̇ and Eheating indicates the output exergy asAlso in Eq. (9), Ecooling sociated with the cooling and heating effects, respectively as [29]: ̇ ̇ − E20 ̇ Ecooling = E19
H
4. Results and discussion Before presenting results of thermodynamic modeling, the accuracy of developed models is validated by comparing the results of present model with those reported in literature previously. For validation purposes, developed thermodynamic model in this work is applied to the proposed CCHP system by Wang et al. [22] and the calculated values for energy and exergy efficiencies are compared. Such a comparison is illustrated in Fig. 7 for varying turbine inlet pressures, while the other operating conditions of the system are set as those given in Ref. [22]. Referring to Fig. 7, an acceptable agreement is detected between the results of present work with those reported previously. In addition to the model validation of the overall cycle using theoretical data, our developed model for ejector is validated based on the experimental results reported by Huang et al. [30]. The comparison between the results of theoretical modeling of the ejector in this work with experimental data can be found in a recently published paper by the authors [31]. The parametric study plays an important role in performance evaluation of novel thermodynamic systems, by which the influence of design variables is investigated on the system performance. For the two proposed systems in this work, four important parameters can be considered as design/operating variables. These parameters are: turbine inlet (or boiler) pressure (Pin, T = P) 1 , turbine outlet pressure (Pout , T = P2) , evaporator temperature (Te = T13) and heater outlet temperature (Tout , H = Tin, IHE = T3) . Also, energy and exergy efficiencies (ηen & ηex ) are considered as the main objective functions. However, the values of net output power (Ẇ net ) as well as cooling (Q̇ e ) and heating outputs (Q̇ H ) are also considered and discussed as the products of proposed systems. For systems’ modeling, the given values in Table 2 are assumed as the input data and basic assumptions. In parametric study, when the influence of a design variable is to be studied the other variables are kept constant as those given in Table 2. The co- or tri-generation systems may be designed to produce a given value of power or heating or cooling outputs [32]. In the scientific literature, particularly for
Fig. 7. Comparison of results of this work with those reported by Wang et al. [22]. 7
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Table 2 Input data and basic assumptions [22,23]. Parameter
Value
Environmental pressure (MPa) Environmental temperature (°C) Turbine inlet pressure (MPa) Turbine outlet pressure (MPa) Turbine isentropic efficiency (%) Heater outlet temperature (°C) (system b) Heater outlet temperature (°C) (system a) Evaporator temperature (°C) Geothermal brine temperature (°C) Geothermal brine mass flow rate (kg/s) Pump isentropic efficiency (%) Nozzle isentropic efficiency (%) Mixing isentropic efficiency (%) Diffuser isentropic efficiency (%) Pinch point temperature difference (°C) Ejector back pressure (MPa)
0.101 15 15 7.6 85 120 70 5 230 1 70 90 88 85 10 5.5
Table 4 Performance parameters of the proposed systems at the base case operating conditions. Parameter
System a
System b
Turbine power (kW) Pump power (kW) Net output power (kW) Cooling output (kW) Heating output (kW) Thermal efficiency (%) Exergy efficiency (%)
27.18 10.01 17.18 12.54 69.13 43.84 25.27
40.53 14.92 25.61 22.05 42.11 39.82 32.14
result of increasing temperature and enthalpy of the ejector primary flow) which brings about a higher refrigerant flow rate across the evaporator. However, the cooling output with increasing Pin, T for system a is almost constant because the gas cooler outlet (ejector primary flow) temperature is constant in this configuration. Decrease of heating output with increasing Pin, T as illustrated in Fig. 8(e) is associated with the reduction of turbine outlet temperature. Decreasing Q̇H with increasing Pin, T is the main reason for the reduction of thermal efficiency as depicted by Fig. 8(a). Another finding which can be drawn from Fig. 8(b) and (c) is that, the better performance of system b than the system a is more pronounced at higher turbine inlet pressures. As can be seen, for example at Pin, T = 9 MPa the exergy efficiency of system b is 23.0% higher than system a (13.9% compared to 11.3%), while for Pin, T = 17 MPa the exergy efficiency of system b is 28.7% higher than system a (35.0% compared to 27.2%). As another key design variable for the proposed CCHP systems is the turbine outlet pressure, the effect of which is presented in Fig. 9(a–e). The figure indicates that for all the practical values of turbine outlet ̇ , Qė and ηex than system a, pressure, system b has significantly higher Wnet while its heating output and thermal efficiency is lower. Referring to Fig. 9(b), it is seen that exergy efficiency of system b is averagely 30% higher than that for system a. Fig. 9(c) indicates that, an increase of turbine back pressure results in a decrease of net output power due to the reduction of enthalpy drop across the turbine. The reduction of output power is the main responsible of decreasing exergy efficiency with increasing Pout , T as presented by Fig. 9(b). From Fig. 9(d) it can be found that, as Pout , T increases the cooling
results in a higher temperature of CO2 entering the boiler. Thus, the average temperature at which heat is absorbed by the working fluid in the boiler is increased and the performance would be improved. In other words, as the temperature of CO2 at boiler inlet increases, more vapor would be generated via the boiler and the entering mass flow rate to the turbine increases, as a result of which the net output power is increased. However, the lower heating output of system b than the system a (as illustrated in Fig. 8(e)) is the main responsible for lower thermal efficiency of system b. The lower heating output in system b is due to the fact that the temperature of CO2 at the heater outlet should be higher to justify the employment of an IHE instead of a gas cooler. In evaluating the effects of turbine inlet pressure, Fig. 8(a–e) iṅ , Qė and dicates that as turbine inlet pressure increases, the values of Wnet ηex are increased, while the values of Q̇H and ηth decrease. It is clear that, increasing turbine inlet pressure results in a higher pressure ratio across the turbine which brings about a higher enthalpy drop and work output ̇ and ηex are increased. From of the turbine, as a result of which Wnet Fig. 8(b) it can be observed for system b that, an increase of Pin, T from 9 MPa to 17 MPa results in a significant improvement of exergy efficiency by 151.8% (from 13.9% to 35.0%). The increase of cooling output with increasing Pin, T for system b (as presented by Fig. 8(d)) is due to the increase of entrainment ratio (as a Table 3 Thermodynamic properties of the state points for system b. State
Fluid
T (°C)
P (MPa)
h (kJ kg−1)
s (kJ kg−1 K−1)
ṁ (kg s−1)
1 2 3 4 5 6 7 8 9 10 10a 11 12 13 14 15 16 17 18 19 20 21 22
CO2 CO2 CO2 CO2 CO2 CO2 CO2 CO2 CO2 CO2 CO2 CO2 CO2 CO2 CO2 Geothermal brine Geothermal brine Water Water Water Water Water Water
220.00 159.50 120.00 42.68 5.00 5.00 22.74 18.27 18.27 32.68 72.51 18.27 5.00 5.00 5.00 230.00 180.00 15.00 20.00 15.00 10.00 15.00 60.00
15.000 7.600 7.600 7.600 3.969 3.969 5.500 5.500 5.500 15.000 15.000 5.500 3.969 3.969 3.969 — — 0.101 0.101 0.101 0.101 0.101 0.101
133.10 88.12 41.37 −81.55 −100.10 −93.47 −81.27 −256.70 −256.70 −240.10 −117.20 −256.70 −256.70 −79.29 −79.28 990.20 764.80 63.01 83.93 63.01 42.09 63.01 251.20
−0.5571 −0.5386 −0.6519 −1.0050 −0.9976 −0.9736 −0.9684 −1.5700 −1.5700 −1.5540 −1.5700 −1.5700 −1.5600 −0.9227 −0.9226 2.6070 2.1360 0.2242 0.2962 0.2242 0.1510 0.2242 0.8311
0.9008 0.9008 0.9008 0.9008 1.0250 1.0250 1.0250 1.0250 0.9008 0.9008 0.9008 0.1243 0.1243 0.1243 0.1243 1 1 8.5960 8.5960 1.0540 1.0540 0.2237 0.2237
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Fig. 8. (a–e) Effects of turbine inlet pressure on the performance of proposed CCHP systems.
depicted by Fig. 9(a). The influence of evaporator temperature on performance of proposed CCHP systems is illustrated in Fig. 10(a–e). For all the evaporator ̇ , Qė and ηex for temperatures, Fig. 10(a–e) indicates higher values of Wnet system b compared to system a, while system a have higher thermal efficiency and heating output than the system b. Also, it is obvious from Fig. 10(b), (c) and (e) that the evaporator temperature does not affect ̇ and Q̇H significantly, as expected. However, Fig. 10(d) and (a) ηex , Wnet show that Qė and hence ηth is increased with increasing evaporator
output is also increased. This is due to the fact that the pressure of primary flow of the ejector increases which results in a higher entrainment of the secondary flow and higher cooling production. Referring to Fig. 9(e) it can be observed that increasing turbine outlet pressure results in an increase of heating output. This is due to the increase of turbine outlet (heater inlet) temperate for higher turbine outlet pressures. Increasing the cooling and heating output with turbine inlet pressure is the main reason for increasing thermal efficiency as
9
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Fig. 9. (a–e). Effects of turbine outlet pressure on the performance of proposed CCHP systems.
of heating output of the CCHP system. It can be observed that if the temperature is increased from 110 to 130 °C, the heating output would decrease by around 35.4% from 50.8 kW to 32.8 kW. Such a significant decrease of Q̇H justifies the reduction of thermal efficiecny with increasing heater outlet temperature as presented in Fig. 11(a). However, ̇ , Qė and ηex are enthe other performance parmeters including Wnet hanced slightly as the heater outlet temperature increases as shown in Fig. 11(a) and (b). This is why an increase of heater outlet temperature brings about a higher temperature of working fluid entering the boiler
temperature. With increasing the evaporator temperature its pressure is also increased which results in an increase of the entrainment ratio of ejector, as a result of which the output cooling increases. Another important design variable in proposed system b in this work is the heater outlet temperature by which the value of heating output can be specified. Effects of this parameter on the system performance are represented by Fig. 11(a–c). As expected and shown in Fig. 11(c), increasing the heater outlet temperature results in a decrease
10
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Fig. 10. (a–e) Effects of evaporator temperature on the performance of proposed CCHP systems.
as a result of which the mass flow rate of CO2 entering the turbine increases resulting in a higher power generation by the turbine. Results of exergy destruction assessment in component level for two proposed systems are illustrated in Fig. 12, at the base case operating conditions as given in Table 3. Referring to Fig. 12, the boiler and heater have the largest contribution in total exergy destruction in both system a and system b. However, exergy destruction in these components in system b is significantly lower than the corresponding value in
system a. The replacement of gas cooler with IHE in system b leads to a higher temperature of entering CO2 to the boiler which results in a lower exergy destruction due to a lower temperature difference between the hot and cold streams. Also, the outlet temperature of heater in system b is higher than that in system a which leads to a lower exergy destruction. From Fig. 12 it can be observed that, total exegy destruction in system b is lower than that in system a by 10.4%.
11
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Fig. 11. (a–c). Effects of heater outlet temperature on the performance of system b.
Fig. 12. Exergy destruction rates within the components of proposed CCHP systems.
12
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Table 5a Performance comparison of proposed CCHP systems in this paper with proposed system by Wang et al. [22].
Proposed system by Wang et al. System a System b
Tin, T (°C)
Pin, T (MPa)
Te (°C)
ṁ 1 (kg/s)
̇ (kW) Wnet
Qė (kW)
Q̇H (kW)
ηth (%)
ηex (%)
220 220 220
15 15 15
5 5 5
0.561 0.561 0.561
0.109 15.95 15.95
7.966 11.64 13.73
63.545 64.20 26.23
53.0 43.84 39.82
28.8 25.27 32.14
Table 5b Performance comparison of proposed CCHP systems in this paper with proposed system by Xu et al. [23].
Proposed system by Wang et al. System a System b
Tin, T (°C)
Pin, T (MPa)
Te (°C)
ṁ 1 (kg/s)
̇ (kW) Wnet
Qė (kW)
Q̇H (kW)
ηth (%)
ηex (%)
220 220 220
15 15 15
5 5 5
1.4 1.4 1.4
−12.94 28.26 28.26
74.51 29.05 31.54
166 192.2 97.43
N.A. 45.95 44.82
25.67 20.14 25.44
5. With variation of decision variables, thermal and exergy efficiencies show different trend due to the dominance of heating output on thermal efficiency. 6. As turbine inlet pressure increases, the net output power and exergy efficiency are also increased, while the heating output and thermal efficiency is decreased.
4.1. Performance comparison with similar CCHP systems For performance improvement of the overall CCHP system, it is deemed essential to employ an efficient thermodynamic cycle. In this regard, performance of the proposed CCHP systems in this work is compared with similar systems (based on ejector transcritical CO2 cycle) proposed previously by other researchers, for the same operating conditions. Results of such a comparison is outlined in Tables 5a and 5b. Referring to these Tables the most important advantage of proposed systems in this work is the capability of producing considerable values of output power, while the previously proposed systems have very low or even no net output power. Beside this advantage, exergy efficiency of the proposed system b in this work is higher than (or comparable with) ̇ for the previously proposed systems. The given negative value for Wnet proposed system by Xu et al. (as outlined in Table 5b) indicates that, the system needs input power from electricity grid to provide heating and cooling effects.
CRediT authorship contribution statement V. Zare: Conceptualization, Methodology, Project administration, Writing - original draft, Writing - review & editing. H. Rostamnejad Takleh: Software, Validation, Investigation. Declaration of Competing Interest The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper.
5. Conclusions References In the present work, transcritical CO2 cycle with an ejector is integrated with conventional Rankine cycle to make a novel CCHP system which is driven by medium-temperature geothermal energy. This system is then modified by replacing the gas cooler with an internal heat exchanger for performance improvement. Performance of the two systems is thermodynamically modeled and compared with each other and with similar previously proposed CCHP systems. The quantitative results of comparison proved the superiority of systems proposed in this work over the previously proposed ones. The influence examination of design parameters on the proposed systems’ performance yielded important insights to their operation as follows:
[1] Boyaghchi FA, Heidarnejad P. Thermoeconomic assessment and multi objective optimization of a solar micro CCHP based on Organic Rankine Cycle for domestic application. Energy Convers Manage 2015;97:224–34. [2] Liu M, Shi Y, Fang F. Combined cooling, heating and power systems: a survey. Renew Sustain Energy Rev 2014;35:1–22. [3] Ebrahimi M, Keshavarz A, Jamali A. Energy and exergy analyses of a micro-steam CCHP cycle for a residential building. Energy Build 2012;45:202–10. [4] Wang Z, Han W, Zhang N, Liu M, Jin H. Proposal and assessment of a new CCHP system integrating gas turbine and heat-driven cooling/power cogeneration. Energy Convers Manage 2017;144:1–9. [5] Chahartaghi M, Sheykhi M. Energy, environmental and economic evaluations of a CCHP system driven by Stirling engine with helium and hydrogen as working gases. Energy 2019;174:1251–66. [6] Gholamian E, Zare V, Mousavi SM. Integration of biomass gasification with a solid oxide fuel cell in a combined cooling, heating and power system: a thermodynamic and environmental analysis. Int J Hydrogen Energy 2016;41:20396–406. [7] Rey G, Ulloa C, Cacabelos A, Barragáns B. Performance analysis, model development and validation with experimental data of an ICE-based micro-CCHP system. Appl Therm Eng 2015;76:233–44. [8] Chang H, Wan Z, Zheng Y, Chen X, Shu S, Tu Z, et al. Energy analysis of a hybrid PEMFC–solar energy residential micro-CCHP system combined with an organic Rankine cycle and vapor compression cycle. Energy Convers Manage 2017;142:374–84. [9] Chang H, Wan Z, Zheng Y, Chen X, Shu S, Tu Z, et al. Energy-and exergy-based working fluid selection and performance analysis of a high-temperature PEMFCbased micro combined cooling heating and power system. Appl Energy 2017;204:446–58. [10] Parikhani T, Azariyan H, Behrad R, Ghaebi H, Jannatkhah J. Thermodynamic and thermoeconomic analysis of a novel ammonia-water mixture combined cooling, heating, and power (CCHP) cycle. Renew Energy 2020;145:1158–75. [11] Mohammadi K, Powell K. Thermodynamic and economic analysis of different cogeneration and trigeneration systems based on carbon dioxide vapor compression refrigeration systems. Appl Therm Eng 2020;164:114503. [12] Chang H, Duan C, Xu X, Pei H, Shu S, Tu Z. Technical performance analysis of a micro-combined cooling, heating and power system based on solar energy and high temperature PEMFC. Int J Hydrogen Energy 2019;44:21080–9.
1. The gas cooler can be replaced by an internal heat exchanger for a wide range of practical operating conditions. 2. System b (in which the gas cooler is replaced by and internal heat exchanger) has significantly higher values of net output power and cooling and exergy efficiency than the system a. At a typical opeṙ , Qė and ηex for system b are reating condition, the values of Wnet spectively, 49.1%, 75.8% and 30.9% higher than the corresponding values for system a. 3. System b yields lower values of heating output and thermal efficiency as compared to system a. At a typical operating condition, the values of Q̇H and ηth for system b are respectively 39.1% and 9.2% lower than those for system a. 4. The most important advantage of proposed systems in this work as compared to the similar previously proposed systems, is the capability of producing considerable values of output power, while the previously proposed systems have very low (or even no) net output power. 13
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[13] Mosaffa AH, Farshi LG. Thermodynamic and economic assessments of a novel CCHP cycle utilizing low-temperature heat sources for domestic applications. Renew Energy 2018;120:134–50. [14] Mohammadi A, Mehrpooya M. Energy and exergy analyses of a combined desalination and CCHP system driven by geothermal energy Organic Rankine cycle. Appl Therm Eng 2017;116:685–94. [15] Zare V. A comparative thermodynamic analysis of two tri-generation systems utilizing low-grade geothermal energy. Energy Convers Manage 2016;118:264–74. [16] Ghaebi H, Parikhani T, Rostamzadeh H. A novel trigeneration system using geothermal heat source and liquefied natural gas cold energy recovery: energy, exergy and exergoeconomic analysis. Renew Energy 2018;119:513–27. [17] Takleh HR, Zare V. Employing thermoelectric generator and booster compressor for performance improvement of a geothermal driven combined power and ejectorrefrigeration cycle. Energy Convers Manage 2019;186:120–30. [18] Nemati A, Mohseni R, Yari M. A comprehensive comparison between CO2 and Ethane as a refrigerant in a two-stage ejector-expansion transcritical refrigeration cycle integrated with an organic Rankine cycle (ORC). J Supercrit Fluids 2018;133:494–502. [19] Yari M, Zare V, Mahmoudi SMS. Parametric study and optimization of an ejectorexpansion TRCC cycle integrated with a water purification system. Proc Inst Mech Eng Part A J Power Energy 2013;227:383–98. [20] Zhu Y, Li C, Zhang F, Jiang P. Comprehensive experimental study on a transcritical CO2 ejector-expansion refrigeration system. Energy Convers Manage 2017;151:98–106. [21] Liu X, Fu R, Wang Z, Lin L, Sun Z, Li X. Thermodynamic analysis of transcritical CO2 refrigeration cycle integrated with thermoelectric subcooler and ejector. Energy Convers Manage 2019;188:354–65. [22] Wang J, Zhao P, Niu X, Dai Y. Parametric analysis of a new combined cooling,
[23] [24] [25]
[26] [27]
[28] [29]
[30] [31]
[32]
14
heating and power system with transcritical CO2 driven by solar energy. Appl Energy 2012;94:58–64. Xu XX, Liu C, Fu X, Gao H, Li Y. Energy and exergy analyses of a modified combined cooling, heating, and power system using supercritical CO2. Energy 2015:414–22. Wang F, Li DY, Zhou Y. Analysis for the ejector used as expansion valve in vapor compression refrigeration cycle. Appl Therm Eng 2016;96:576–82. Ahmadi P, Dincer I, Rosen MA. Thermodynamic modeling and multi-objective evolutionary-based optimization of a new multigeneration energy system. Energy Convers Manage 2013;76:282–300. Wang J, Dai Y, Gao L. Parametric analysis and optimization for a combined power and refrigeration cycle. Appl Energy 2008;85:1071–85. Dai Y, Wang J, Gao L. Exergy analysis, parametric analysis and optimization for a novel combined power and ejector refrigeration cycle. Appl Therm Eng 2009;29:1983–90. Zare V. A comparative exergoeconomic analysis of different ORC configurations for binary geothermal power plants. Energy Convers Manage 2015;105:127–38. Balafkandeh S, Zare V, Gholamian E. Multi-objective optimization of a tri-generation system based on biomass gasification/digestion combined with S-CO2 cycle and absorption chiller. Energy Convers Manage 2019;200:112057. Huang BJ, Chang JM, Wang CP, Petrenko VA. A 1-D analysis of ejector performance. Int J Refrig 1999;22:354–64. Takleh HR, Zare V. Performance improvement of ejector expansion refrigeration cycles employing a booster compressor using different refrigerants: thermodynamic analysis and optimization. Int J Refrig 2019;101:56–70. Chang H, Xu X, Shen J, Shu S, Tu Z. Performance analysis of a micro-combined heating and power system with PEM fuel cell as a prime mover for a typical household in North China. Int J Hydrogen Energy 2019;44:24965–76.
Contents lists available at ScienceDirect
Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman
Novel geothermal driven CCHP systems integrating ejector transcritical CO2 and Rankine cycles: Thermodynamic modeling and parametric study
T
⁎
V. Zare , H. Rostamnejad Takleh Faculty of Mechanical Engineering, Urmia University of Technology, Urmia, Iran
A R T I C LE I N FO
A B S T R A C T
Keywords: CCHP Ejector Transcritical CO2 Geothermal energy Exergy analysis
With recent technical advances, Combined Cooling, Heating and Power (CCHP) system offers a practical and economical alternative to conventional energy conversion systems. Benefits of these systems are more marked with employing efficient thermodynamic cycle and using renewable energy resources. In this regard in the present work, a novel geothermal driven CCHP system, in which the ejector transcritical CO2 cycle is integrated with conventional Rankine cycle, is proposed (system a). The proposed system is then modified by replacing the gas cooler with an internal heat exchanger to make a more efficient CCHP system (system b). To investigate the first and second law performance of the proposed systems, thermodynamic models are developed and parametric analysis is carried out to examine the influences of design variables. The results indicated that, the gas cooler of conventional system can be replaced by an internal heat exchanger for a wide range of practical operating conditions and this replacement can improve the exergy efficiency, net output power and output cooling respectively by 30.9%, 49.1% and 75.8% at the expense of 39.1% reduction in heating output. Also a comparison is made between the two proposed systems in this work with similar systems (based on ejector transcritical CO2 cycle) proposed previously by other researchers and superiority of proposed systems in this paper is revealed and discussed.
1. Introduction The environmental problems and resource shortages of fossil fuels brings the necessity of employing high-efficiency energy conversion systems and renewable energy sources. Combined Cooling, Heating and Power (CCHP) concept is a promising technology to improve energy conversion efficiency by which the power, cooling and heating are produced simultaneously in a single unit. Such systems have many environmental and economic advantages over the separate production of power, heat and cold [1]. In conventional CCHP systems fossil fuels are usually utilized as the primary energy source [2]. In these systems prime movers such as steam turbines [3], gas turbines [4], stirling engines [5], fuel cells [6] and internal combustion engines [7] are employed. However, recently researchers have paid more attention on renewable energy-based CCHP systems. More recently, some research works are devoted to investigate Rankine or Kalina cycle-based micro CCHP systems using energy and exergy analyses [8,9]. Parikhani et al. [10] proposed a new CCHP system based on a modified version of Kalina cycle and analyzed its feasibility based on thermodynamic and thermoeconomic analyses. They reported an exergy efficiency and overall unit product cost of
⁎
27.68% and 198.3 $/GJ for the system. Novel integrated cogeneration and trigeneration configurations based on carbon dioxide vapor compression systems are proposed and analysed by Mohammadi and Powell [11], who investigated the systems for a 1000 kW capacity and evaporator temperatures of −35 °C to −45 °C. Chang et al. [12] proposed and evaluated two micro CCHP systems based on Rankine and vapor compression cycles and a high temperature PEMFC. They obtained the values of 49.7% and 47.4% for exergy efficiency of their proposed systems. In scientific literature, very few research works are available on geothermal driven CCHP systems. Mosaffa and Garousi [13] proposed a novel geothermal-based CCHP system consisting of an Organic Rankine Cycle (ORC), an ejector refrigeration cycle and a domestic water heater. They considered four different working fluids and found that the best performance of the system is achieved with R123. Another geothermalbased CCHP system is proposed by Mohammadi and Mehrpooya [14], who reported an exergy efficiency of 38.10% at a typical operating condition for a geothermal water temperature of 230 °C. Their proposed system employs the Kalina cycle as the Power Generation Unit (PGU) and it is also capable of producing some fresh water by reverse osmosis. Zare [15] made a comparison between two geothermal driven CCHP
Corresponding author. E-mail address: [email protected] (V. Zare).
https://doi.org/10.1016/j.enconman.2019.112396 Received 28 August 2019; Received in revised form 15 November 2019; Accepted 8 December 2019 0196-8904/ © 2019 Elsevier Ltd. All rights reserved.
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Nomenclature C Ė h ṁ P Q̇ s T Ẇ
e IHE is H ms n net out PGU pu T TRCC tot
Velocity (m/s) Exergy rate (kW) Enthalpy (kJ/kg) Mass flow rate (kg/s) Pressure (MPa) Heat transfer rate (kW) Specific entropy (kJ/kg.K) Temperature (°C) Power (kW)
Subscripts and abbreviations 0 b c d D CCHP ej EV
Environmental state boiler Condenser Diffuser Destruction Combined cooling, heating and power Ejector Expansion valve
evaporator Internal heat exchanger Isentropic process heater Mixing section Nozzle section Net Outlet Power Generation Unit Pump Turbine transcritical CO2 Total
Greek symbols
ηen ηex ηpu ηT µ
First law efficiency Second law efficiency Pump isentropic efficiency Turbine isentropic efficiency Entrainment ratio
considered as a favored choice with lots of proven advantages. CO2 as a working fluid in energy conversion systems has unique characteristics, such as low viscosity, high heat transfer coefficient, non-toxicity and inflammability. Also it has zero ozone depletion potential and negligible global warming potential beside a very low cost. Despite these favorable specifications of CO2, just a few studies can be found in open literature in the context of CCHP systems employing CO2 as the working fluid, however, a large number of studies are conducted on investigation of the ejector-expansion transcritical CO2 refrigeration cycle [18–21]. The first proposal of a CCHP system based on TRCC cycle was made by Wang et al. [22] in 2012. In their proposed system a TRCC cycle is integrated with a Brayton power cycle for trigeneration of cooling, heating and power. The cycle is supposed to be driven by the solar energy with a temperature of 230 °C. They found that, under a typical operating condition, thermal and exergy efficiencies of the system are
systems with ORC and Kalina cycles as the PGUs and LiBr/H2O absorption chiller for cooling production. His results showed that, the system with Kalina cycle generates 12.2% more power than the other system under the optimized operating conditions. A novel CCHP system driven by geothermal energy based on absorption chiller with ammonia-water mixture and an LNG power generation sub-system is developed by Ghaebi et al. [16]. They found that, their system (For a geothermal water temperature of 398 K and mass flow rate of 10 kg/s) yields net power output of 405.1 kW and thermal and exergy efficiencies of 85.92%, 18.52%, respectively. The increasing interest for geothermal driven CCHP systems can be maintained via introducing novel systems with higher efficiency and lower environmental impacts. Also, since the cold production is more complicated than the heat generation in CCHP systems, more attention should be paid on efficient cooling production [17]. Regarding these aspects, transcritical CO2 (TRCC) cycle equipped with an ejector can be
Fig. 1. Schematic diagram of the proposed CCHP system based on TRCC cycle by Wang et al. [22]. 2
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of proposed systems and conducting a parametric analysis to show their feasibility 4. Making a comparison between the proposed systems in this paper with the two similar previously proposed systems
53.0 and 28.8%. Xu et al. [23] modified the proposed system by Wang et al. [22] by replacing the turbine with an extraction one in order to provide relatively high pressure stream as the primary flow for the ejector, as a result of which the cooling output can be increased. They showed that, their proposed system yields energy utilization factor and exergy efficiency of 70.13 and 25.67% at a typical operating condition. In recent two proposed CCHP systems based on TRCC cycle, a technical mistake about the power generation is concerned. In the first system proposed by Wang et al. [22], the net output power ̇ = WTurbine ̇ ̇ − WCompressor (Wnet ) is close to zero in a wide range of operating conditions and for some operating conditions is even negative (as illustrated in Figs. 3–6 of Ref. [22]) which means that the system cannot generate output power and needs input power for operation. Also, in the second proposed system by Xu et al. [23] it is seen that the net output power of the system is negative in almost all the range of investigated operating conditions (as can be observed from Figs. 4–7 in Ref. [23] the consumed power by compressor is more than the produced power by the turbine). Considering the above discussion, in the present work a novel CCHP system based on ejector transcritical CO2 cycle is proposed and analyzed. The proposed system is an integration of TRCC and Rankine cycles which is capable of producing satisfactory amount of net output power beside the cooling and heating production. Also, a modified configuration of the proposed system in this work is suggested in which the gas cooler of TRCC cycle is replaced by an Internal Heat Exchanger (IHE) to enhance the system performance. For the two proposed systems, comprehensive energy and exergy analyses are conducted and a parametric study is carried out to investigate the effects of key parameters on the system performance. Also, performance of the two systems is compared to evaluate the influence of an IHE employment. The contribution, novelty and objectives of the present paper could be summarized as follows:
2. System description Before presenting the proposed systems in this work, it would be appropriate to present a brief discussion on the two previously proposed systems in the literature. Fig. 1 illustrates the CCHP system based on TRCC cycle proposed by Wang et al. [22]. As can be seen, the system is an integration of ejector-expansion TRCC refrigeration cycle and a Brayton power cycle. In this CCHP system the cooling, heating and power outputs are respectively produced by evaporator, heater and turbine. However, as mention before and reported by the authors, in this system the consumed power by the compressor is almost equal to (or even more than) the generated power by turbine. Therefore the system cannot be regarded as a CCHP system in practical applications. The modified CCHP system based on TRCC cycle proposed by Xu et al. [23] is illustrated in Fig. 2. As can be seen in this system the turbine is replaced by an extraction one by which a relatively high pressure stream is provided for the ejector as its primary flow to increase the cooling production. However, less attention is paid to the power production as the generated power by the turbine is less than the consumed power by compressor. Therefore, this system cannot be considered as a CCHP one in practical applications. In the present paper to solve the net power generation problem in two above-mentioned systems, the Brayton sub-system is proposed to be replaced by a Rankine one (as illustrated in Fig. 3), so that the compressor is replaced by a pump. To do so, a condenser is required after the ejector to convert the two-phase flow exiting from the ejector to a saturated liquid. Referring to Fig. 1, the high temperature supercritical CO2 (stream 1) enters the turbine where it is expanded to a lower pressure and generates power. The leaving CO2 from turbine (stream 2) enters to a heater to provide heating effect for the heat user, after which it would reject heat in the gas cooler (stream 3). To drive the ejector, exiting stream from the gas cooler (stream 4) enters to the ejector as the primary flow, so that the secondary vapor from the
1. Proposal of a novel geothermal driven CCHP system as an integration of TRCC and Rankine cycles 2. Replacing the gas cooler of TRCC cycle by an IHE to enhance the proposed CCHP system performance 3. A comprehensive evaluation of the first and second law performance
Fig. 2. Schematic diagram of the proposed CCHP system based on TRCC cycle by Xu et al. [23]. 3
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Fig. 3. Schematic diagram of the proposed CCHP system based on TRCC cycle in this paper (system a).
be saturated. 6. Appropriate values for isentropic efficiencies of turbine and pump are assumed as given in Table 2. 7. In ejector modeling, the effects of mixing and frictional losses through the ejector are taken into account by appropriate values for nozzle efficiency, diffuser efficiency and mixing efficiency as listed in Table 2. 8. In ejector modeling, constant-pressure mixing model is considered.
evaporator (stream 13) is entrained. The exiting CO2 from ejector, which is a two-phase stream, (stream 7) flows to a condenser to reject heat and to be condensed to saturated liquid. Then stream 8, flows into a separator where it is divided into two streams, the first one (stream 11) flows to the evaporator after being expanded in the valve and the second one (stream 9) goes to the boiler after being pumped. To improve the performance of proposed system in this work, one more modification is made regarding the working conditions and temperature levels of entering and exiting streams from the gas cooler. In this configuration (as illustrated in Fig. 4) the gas cooler is proposed to be replaced by an Internal Heat Exchanger (IHE). In this system the rejected heat from the gas cooler, as a wasted one, is recovered by IHE as a result of which the working fluid enters the boiler (stream 10a) with a higher temperature. The higher temperature of the entering working fluid to the boiler results in a higher efficiency of the cycle. To model the proposed CCHP systems some assumptions are made as follows [22,23]:
3. Thermodynamic modeling Thermodynamic modeling of the proposed CCHP systems in this work is developed using the Engineering Equation Solver (EES) software. Each component of the systems is treated as a control volume for which the mass and energy conservations is applied. In modeling the considered CCHP systems, more attention is paid to ejector model, since it is a key component of the systems and plays important role to provide cooling. The modeling procedure of the ejector is described below.
1. Systems are working under steady state conditions. 2. There is no heat loss from the system components to the environment and no pressure drops in the pipelines. 3. Changes in kinetic and potential exergies and energies are ignored. 4. The isenthalpic process is considered for the throttling valve. 5. The streams at the evaporator and condenser outlets are assumed to
3.1. Ejector modeling Schematics of the ejector structure is illustrated in Fig. 5 which is comprised of four sections: suction, mixing, constant-area and diffuser sections. It employs a converging–diverging nozzle and converts the 4
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Fig. 4. Schematic diagram of the proposed CCHP system based on TRCC with an IHE (system b) and its T-s diagram.
constant-pressure model is preferred to the constant-area model for its simplicity and accuracy [24,25], so in this paper the constant-pressure method is selected. In order to simulate the ejector, as a component of the CCHP systems under consideration, presented procedure by Li et al. [26] is applied. The ejector modeling is started by determining thermodynamic states of the primary and secondary stream of the ejector. These state points, for considered system in this work as illustrated in Fig. 3, can be
mechanical energy of pressure type to kinetic energy, as a result of which a low-pressure region is created and a flow stream is entrained by suction. The occurring processes in the ejector include: expanding the high-pressure primary flow through the nozzle, mixing with the lowpressure secondary flow in mixing section and diffusing via the diffuser at the ejector outlet. Regarding the mixing process in the mixing section, two methods can be used in ejector modeling: 1) constant-pressure mixing and 2) constant-area mixing. In previous research works,
5
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Fig. 5. Ejector schematics and operation principle.
The exergy balance equation for the components of the systems, is expressed as follows to evaluate the exergy destruction/loss within the component:
determined noting that the gas cooler and evaporator temperature and pressure are known. Then, for entrainment ratio, an initial value is guessed while known values are considered for nozzle efficiency and diffuser efficiency. The relations used for modeling the ejector are given below, respectively for motive nozzle, mixing section and diffuser [25,27]: For motive nozzle:
ηn =
h4 − h5 , h4 − h5, is
h5, is = h (P5, s4 ),
C5 =
2(h4 − h5)
EḊ =
C ηm ⎜⎛ 5 ⎟⎞, 1 ⎝ + μ⎠
h6 =
C2 h4 + μh13 − 6 1+μ 2
(1)
(2)
For the diffuser:
h 7 = h6 +
C62 , 2
ηd =
ηth =
̇ ̇ ̇ + Qevaporator + Qheater Wnet ̇ Qin
(8)
ηex =
̇ + Ecooling ̇ ̇ Wnet + Eheating ̇ Ein
(9)
h7, is − h6 Ấ
(3)
h 7 − h6 Ấ
The obtained values for h 7 and h 7 from Eq. (3) will be compared with each other. If these two values would be close together with an acceptable accuracy (0.1 in this paper) the iterative calculations would be finished. Fig. 6 shows the flowchart of ejector modeling procedure. 3.2. Energy and exergy analyses To perform the energy analysis on a thermodynamic system, the mass conservation and first law principal would be applied to the system components. For a control volume accompanying a steady state process, the conservation of mass and energy can be expressed as:
∑ ṁ in = ∑ ṁ out
(4)
̇ − Wcv ̇ = Qcv
(5)
∑ ṁ out (hout ) − ∑ ṁ in (hin)
In addition to the energy analysis, the exergy analysis based on the second law principal is an important technique to assess the performance of energy systems. It reveals the location, type and true magnitude of inefficiencies in a thermodynamic cycle and provides a better understanding of the processes. Thus, it plays a key role in providing guidelines for designing efficient systems. Exergy, as the maximum attainable work from a system/fluid stream, can be divided into four parts, two of which (kinetic and potential terms) are usually negligible. Also, in systems with no chemical reactions and composition changes, chemical exergy term is not considered as it would be canceled out in exergy balances. Thus, for the CCHP systems in this work, exergy rate of the fluid streams can be expressed as [28]:
̇ = ṁ [(h − h 0) − T0 (s − s0)] E ̇ = Eph
(7)
̇ denotes for the overall input and output ̇ and ∑ Eout in which, ∑ Ein exergies to that component. To simulate the systems’ performance, the relations given in Table 1, along with the equations for ejector modeling as explained above, are introduced into the computer program. For performance evaluation of the proposed CCHP systems from the viewpoint of first and second laws, thermal and exergy efficiency is defined as follows [23]:
For mixing section with a proper guess value for the entrainment ratio:
C6 =
̇ ∑ Eiṅ − ∑ Eout
(6)
Fig. 6. Ejector modeling flowchart. 6
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Table 1 Energy- and exergy-related equations applied to analyse the system performance (system b). Component
Energy equation
Exergy balance equation
Condenser
ṁ 7 (h7 − h8) = ṁ 17 (h18 − h17)
Evaporator
c
ṁ 12 (h13 − h12) = ṁ 19 (h19 − h20)
̇ − E17 ̇ ) EḊ = (E7̇ − E8̇ ) − (E18 e ̇ − E13 ̇ ) − (E20 ̇ − E19 ̇ ) EḊ = (E12
Boiler
ṁ 1 (h1 − h10a) = ṁ 15 (h15 − h16)
̇ − E16 ̇ ) − (E1̇ − E10 ̇ a) EḊ = (E15
Turbine
ẆT = ṁ 1 (h1 − h2) ηis, T = (h1 − h2)/(h1 − h2, is )
T EḊ = (E1̇ − E2̇ ) − ẆT
Ejector
B
ej
̇ + E4̇ ) − E7̇ EḊ = (E13 pu ̇ − E9̇ ) EḊ = Ẇ pu − (E10
Pump
Ẇ pu = ṁ 10 (h10 − h9)
Expansion valve
h11 = h12
̇ − E12 ̇ EḊ = E11
Heater
ṁ 2 (h2 − h3) = ṁ 21 (h22 − h21)
̇ − E21 ̇ ) EḊ = (E2̇ − E3̇ ) − (E22
IHE
ṁ 3 (h3 − h4 ) = ṁ 10 (h10a − h10)
EḊ
ηis, pu = (h10 − h9, is )/(h10 − h9) EV
̇ indicates net output power and input energy ̇ and Qin In Eq. (8), Wnet to the system, respectively as: ̇ = WTurbine ̇ Wnet − ẆPump
(10)
̇ = Qboiler ̇ Qin = ṁ 15 (h15 − h16)
(11)
(12)
̇ ̇ − E21 ̇ Eheating = E22
(13)
IHE
̇ a − E10 ̇ ) = (E3̇ − E4̇ ) − (E10
geothermal heat sources, another design criterion for designing such systems is considered, that is designing and calculating the power, heating and cooling outputs for a given value of heat source mass flow rate. In the present paper this criterion is considered and the proposed systems performances are investigated for a unit mass flow rate of geothermal brine with a given temperature. For a base case operating condition (as given in Table 2), the properties of thermodynamic state points, according to stream numbers given in Fig. 4, are outlined in Table 3. Also, the values of performance parameters are given in Table 4. ̇ , Qė and The figures given in Table 4 reveals that, the values of Wnet ηex for system b are respectively, 49.1%, 75.8% and 30.9% higher than the corresponding values for system a. However, the values of Q̇H and ηth for system b are respectively 39.1% and 9.2% lower than those for system a. As the first important design variable is considered the boiler (turbine inlet) pressure, effects of which are illustrated in Fig. 8(a–e). As an important and general outcome in comparing performance of the two proposed systems, Fig. 8(a–e) reveals that employing the IHE (system b) ̇ , Qė and instead of a gas cooler (system a) would result in a higher Wnet ηex at the expense of decreasing Q̇H and ηth . Referring to Fig. 8(b–d) it can be seen that, for all the values of turbine inlet pressure, system b has significantly higher exergy efficiency and power and cooling outputs than system a, while its heating output is lower as presented by Fig. 8(e). Also, Fig. 8(a) shows that, system b has lower values of thermal efficiency than system a at higher turbine inlet pressures than 11 MPa. The better performance of system ̇ , Qė and ηex can be attributed to b compared to system a, in terms of Wnet the employment of an IHE instead of the gas cooler, as a result of which the rejected heat is recovered via the IHE. This internal heat recovery
̇ ̇ and Eheating indicates the output exergy asAlso in Eq. (9), Ecooling sociated with the cooling and heating effects, respectively as [29]: ̇ ̇ − E20 ̇ Ecooling = E19
H
4. Results and discussion Before presenting results of thermodynamic modeling, the accuracy of developed models is validated by comparing the results of present model with those reported in literature previously. For validation purposes, developed thermodynamic model in this work is applied to the proposed CCHP system by Wang et al. [22] and the calculated values for energy and exergy efficiencies are compared. Such a comparison is illustrated in Fig. 7 for varying turbine inlet pressures, while the other operating conditions of the system are set as those given in Ref. [22]. Referring to Fig. 7, an acceptable agreement is detected between the results of present work with those reported previously. In addition to the model validation of the overall cycle using theoretical data, our developed model for ejector is validated based on the experimental results reported by Huang et al. [30]. The comparison between the results of theoretical modeling of the ejector in this work with experimental data can be found in a recently published paper by the authors [31]. The parametric study plays an important role in performance evaluation of novel thermodynamic systems, by which the influence of design variables is investigated on the system performance. For the two proposed systems in this work, four important parameters can be considered as design/operating variables. These parameters are: turbine inlet (or boiler) pressure (Pin, T = P) 1 , turbine outlet pressure (Pout , T = P2) , evaporator temperature (Te = T13) and heater outlet temperature (Tout , H = Tin, IHE = T3) . Also, energy and exergy efficiencies (ηen & ηex ) are considered as the main objective functions. However, the values of net output power (Ẇ net ) as well as cooling (Q̇ e ) and heating outputs (Q̇ H ) are also considered and discussed as the products of proposed systems. For systems’ modeling, the given values in Table 2 are assumed as the input data and basic assumptions. In parametric study, when the influence of a design variable is to be studied the other variables are kept constant as those given in Table 2. The co- or tri-generation systems may be designed to produce a given value of power or heating or cooling outputs [32]. In the scientific literature, particularly for
Fig. 7. Comparison of results of this work with those reported by Wang et al. [22]. 7
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Table 2 Input data and basic assumptions [22,23]. Parameter
Value
Environmental pressure (MPa) Environmental temperature (°C) Turbine inlet pressure (MPa) Turbine outlet pressure (MPa) Turbine isentropic efficiency (%) Heater outlet temperature (°C) (system b) Heater outlet temperature (°C) (system a) Evaporator temperature (°C) Geothermal brine temperature (°C) Geothermal brine mass flow rate (kg/s) Pump isentropic efficiency (%) Nozzle isentropic efficiency (%) Mixing isentropic efficiency (%) Diffuser isentropic efficiency (%) Pinch point temperature difference (°C) Ejector back pressure (MPa)
0.101 15 15 7.6 85 120 70 5 230 1 70 90 88 85 10 5.5
Table 4 Performance parameters of the proposed systems at the base case operating conditions. Parameter
System a
System b
Turbine power (kW) Pump power (kW) Net output power (kW) Cooling output (kW) Heating output (kW) Thermal efficiency (%) Exergy efficiency (%)
27.18 10.01 17.18 12.54 69.13 43.84 25.27
40.53 14.92 25.61 22.05 42.11 39.82 32.14
result of increasing temperature and enthalpy of the ejector primary flow) which brings about a higher refrigerant flow rate across the evaporator. However, the cooling output with increasing Pin, T for system a is almost constant because the gas cooler outlet (ejector primary flow) temperature is constant in this configuration. Decrease of heating output with increasing Pin, T as illustrated in Fig. 8(e) is associated with the reduction of turbine outlet temperature. Decreasing Q̇H with increasing Pin, T is the main reason for the reduction of thermal efficiency as depicted by Fig. 8(a). Another finding which can be drawn from Fig. 8(b) and (c) is that, the better performance of system b than the system a is more pronounced at higher turbine inlet pressures. As can be seen, for example at Pin, T = 9 MPa the exergy efficiency of system b is 23.0% higher than system a (13.9% compared to 11.3%), while for Pin, T = 17 MPa the exergy efficiency of system b is 28.7% higher than system a (35.0% compared to 27.2%). As another key design variable for the proposed CCHP systems is the turbine outlet pressure, the effect of which is presented in Fig. 9(a–e). The figure indicates that for all the practical values of turbine outlet ̇ , Qė and ηex than system a, pressure, system b has significantly higher Wnet while its heating output and thermal efficiency is lower. Referring to Fig. 9(b), it is seen that exergy efficiency of system b is averagely 30% higher than that for system a. Fig. 9(c) indicates that, an increase of turbine back pressure results in a decrease of net output power due to the reduction of enthalpy drop across the turbine. The reduction of output power is the main responsible of decreasing exergy efficiency with increasing Pout , T as presented by Fig. 9(b). From Fig. 9(d) it can be found that, as Pout , T increases the cooling
results in a higher temperature of CO2 entering the boiler. Thus, the average temperature at which heat is absorbed by the working fluid in the boiler is increased and the performance would be improved. In other words, as the temperature of CO2 at boiler inlet increases, more vapor would be generated via the boiler and the entering mass flow rate to the turbine increases, as a result of which the net output power is increased. However, the lower heating output of system b than the system a (as illustrated in Fig. 8(e)) is the main responsible for lower thermal efficiency of system b. The lower heating output in system b is due to the fact that the temperature of CO2 at the heater outlet should be higher to justify the employment of an IHE instead of a gas cooler. In evaluating the effects of turbine inlet pressure, Fig. 8(a–e) iṅ , Qė and dicates that as turbine inlet pressure increases, the values of Wnet ηex are increased, while the values of Q̇H and ηth decrease. It is clear that, increasing turbine inlet pressure results in a higher pressure ratio across the turbine which brings about a higher enthalpy drop and work output ̇ and ηex are increased. From of the turbine, as a result of which Wnet Fig. 8(b) it can be observed for system b that, an increase of Pin, T from 9 MPa to 17 MPa results in a significant improvement of exergy efficiency by 151.8% (from 13.9% to 35.0%). The increase of cooling output with increasing Pin, T for system b (as presented by Fig. 8(d)) is due to the increase of entrainment ratio (as a Table 3 Thermodynamic properties of the state points for system b. State
Fluid
T (°C)
P (MPa)
h (kJ kg−1)
s (kJ kg−1 K−1)
ṁ (kg s−1)
1 2 3 4 5 6 7 8 9 10 10a 11 12 13 14 15 16 17 18 19 20 21 22
CO2 CO2 CO2 CO2 CO2 CO2 CO2 CO2 CO2 CO2 CO2 CO2 CO2 CO2 CO2 Geothermal brine Geothermal brine Water Water Water Water Water Water
220.00 159.50 120.00 42.68 5.00 5.00 22.74 18.27 18.27 32.68 72.51 18.27 5.00 5.00 5.00 230.00 180.00 15.00 20.00 15.00 10.00 15.00 60.00
15.000 7.600 7.600 7.600 3.969 3.969 5.500 5.500 5.500 15.000 15.000 5.500 3.969 3.969 3.969 — — 0.101 0.101 0.101 0.101 0.101 0.101
133.10 88.12 41.37 −81.55 −100.10 −93.47 −81.27 −256.70 −256.70 −240.10 −117.20 −256.70 −256.70 −79.29 −79.28 990.20 764.80 63.01 83.93 63.01 42.09 63.01 251.20
−0.5571 −0.5386 −0.6519 −1.0050 −0.9976 −0.9736 −0.9684 −1.5700 −1.5700 −1.5540 −1.5700 −1.5700 −1.5600 −0.9227 −0.9226 2.6070 2.1360 0.2242 0.2962 0.2242 0.1510 0.2242 0.8311
0.9008 0.9008 0.9008 0.9008 1.0250 1.0250 1.0250 1.0250 0.9008 0.9008 0.9008 0.1243 0.1243 0.1243 0.1243 1 1 8.5960 8.5960 1.0540 1.0540 0.2237 0.2237
8
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Fig. 8. (a–e) Effects of turbine inlet pressure on the performance of proposed CCHP systems.
depicted by Fig. 9(a). The influence of evaporator temperature on performance of proposed CCHP systems is illustrated in Fig. 10(a–e). For all the evaporator ̇ , Qė and ηex for temperatures, Fig. 10(a–e) indicates higher values of Wnet system b compared to system a, while system a have higher thermal efficiency and heating output than the system b. Also, it is obvious from Fig. 10(b), (c) and (e) that the evaporator temperature does not affect ̇ and Q̇H significantly, as expected. However, Fig. 10(d) and (a) ηex , Wnet show that Qė and hence ηth is increased with increasing evaporator
output is also increased. This is due to the fact that the pressure of primary flow of the ejector increases which results in a higher entrainment of the secondary flow and higher cooling production. Referring to Fig. 9(e) it can be observed that increasing turbine outlet pressure results in an increase of heating output. This is due to the increase of turbine outlet (heater inlet) temperate for higher turbine outlet pressures. Increasing the cooling and heating output with turbine inlet pressure is the main reason for increasing thermal efficiency as
9
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Fig. 9. (a–e). Effects of turbine outlet pressure on the performance of proposed CCHP systems.
of heating output of the CCHP system. It can be observed that if the temperature is increased from 110 to 130 °C, the heating output would decrease by around 35.4% from 50.8 kW to 32.8 kW. Such a significant decrease of Q̇H justifies the reduction of thermal efficiecny with increasing heater outlet temperature as presented in Fig. 11(a). However, ̇ , Qė and ηex are enthe other performance parmeters including Wnet hanced slightly as the heater outlet temperature increases as shown in Fig. 11(a) and (b). This is why an increase of heater outlet temperature brings about a higher temperature of working fluid entering the boiler
temperature. With increasing the evaporator temperature its pressure is also increased which results in an increase of the entrainment ratio of ejector, as a result of which the output cooling increases. Another important design variable in proposed system b in this work is the heater outlet temperature by which the value of heating output can be specified. Effects of this parameter on the system performance are represented by Fig. 11(a–c). As expected and shown in Fig. 11(c), increasing the heater outlet temperature results in a decrease
10
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Fig. 10. (a–e) Effects of evaporator temperature on the performance of proposed CCHP systems.
as a result of which the mass flow rate of CO2 entering the turbine increases resulting in a higher power generation by the turbine. Results of exergy destruction assessment in component level for two proposed systems are illustrated in Fig. 12, at the base case operating conditions as given in Table 3. Referring to Fig. 12, the boiler and heater have the largest contribution in total exergy destruction in both system a and system b. However, exergy destruction in these components in system b is significantly lower than the corresponding value in
system a. The replacement of gas cooler with IHE in system b leads to a higher temperature of entering CO2 to the boiler which results in a lower exergy destruction due to a lower temperature difference between the hot and cold streams. Also, the outlet temperature of heater in system b is higher than that in system a which leads to a lower exergy destruction. From Fig. 12 it can be observed that, total exegy destruction in system b is lower than that in system a by 10.4%.
11
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Fig. 11. (a–c). Effects of heater outlet temperature on the performance of system b.
Fig. 12. Exergy destruction rates within the components of proposed CCHP systems.
12
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Table 5a Performance comparison of proposed CCHP systems in this paper with proposed system by Wang et al. [22].
Proposed system by Wang et al. System a System b
Tin, T (°C)
Pin, T (MPa)
Te (°C)
ṁ 1 (kg/s)
̇ (kW) Wnet
Qė (kW)
Q̇H (kW)
ηth (%)
ηex (%)
220 220 220
15 15 15
5 5 5
0.561 0.561 0.561
0.109 15.95 15.95
7.966 11.64 13.73
63.545 64.20 26.23
53.0 43.84 39.82
28.8 25.27 32.14
Table 5b Performance comparison of proposed CCHP systems in this paper with proposed system by Xu et al. [23].
Proposed system by Wang et al. System a System b
Tin, T (°C)
Pin, T (MPa)
Te (°C)
ṁ 1 (kg/s)
̇ (kW) Wnet
Qė (kW)
Q̇H (kW)
ηth (%)
ηex (%)
220 220 220
15 15 15
5 5 5
1.4 1.4 1.4
−12.94 28.26 28.26
74.51 29.05 31.54
166 192.2 97.43
N.A. 45.95 44.82
25.67 20.14 25.44
5. With variation of decision variables, thermal and exergy efficiencies show different trend due to the dominance of heating output on thermal efficiency. 6. As turbine inlet pressure increases, the net output power and exergy efficiency are also increased, while the heating output and thermal efficiency is decreased.
4.1. Performance comparison with similar CCHP systems For performance improvement of the overall CCHP system, it is deemed essential to employ an efficient thermodynamic cycle. In this regard, performance of the proposed CCHP systems in this work is compared with similar systems (based on ejector transcritical CO2 cycle) proposed previously by other researchers, for the same operating conditions. Results of such a comparison is outlined in Tables 5a and 5b. Referring to these Tables the most important advantage of proposed systems in this work is the capability of producing considerable values of output power, while the previously proposed systems have very low or even no net output power. Beside this advantage, exergy efficiency of the proposed system b in this work is higher than (or comparable with) ̇ for the previously proposed systems. The given negative value for Wnet proposed system by Xu et al. (as outlined in Table 5b) indicates that, the system needs input power from electricity grid to provide heating and cooling effects.
CRediT authorship contribution statement V. Zare: Conceptualization, Methodology, Project administration, Writing - original draft, Writing - review & editing. H. Rostamnejad Takleh: Software, Validation, Investigation. Declaration of Competing Interest The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper.
5. Conclusions References In the present work, transcritical CO2 cycle with an ejector is integrated with conventional Rankine cycle to make a novel CCHP system which is driven by medium-temperature geothermal energy. This system is then modified by replacing the gas cooler with an internal heat exchanger for performance improvement. Performance of the two systems is thermodynamically modeled and compared with each other and with similar previously proposed CCHP systems. The quantitative results of comparison proved the superiority of systems proposed in this work over the previously proposed ones. The influence examination of design parameters on the proposed systems’ performance yielded important insights to their operation as follows:
[1] Boyaghchi FA, Heidarnejad P. Thermoeconomic assessment and multi objective optimization of a solar micro CCHP based on Organic Rankine Cycle for domestic application. Energy Convers Manage 2015;97:224–34. [2] Liu M, Shi Y, Fang F. Combined cooling, heating and power systems: a survey. Renew Sustain Energy Rev 2014;35:1–22. [3] Ebrahimi M, Keshavarz A, Jamali A. Energy and exergy analyses of a micro-steam CCHP cycle for a residential building. Energy Build 2012;45:202–10. [4] Wang Z, Han W, Zhang N, Liu M, Jin H. Proposal and assessment of a new CCHP system integrating gas turbine and heat-driven cooling/power cogeneration. Energy Convers Manage 2017;144:1–9. [5] Chahartaghi M, Sheykhi M. Energy, environmental and economic evaluations of a CCHP system driven by Stirling engine with helium and hydrogen as working gases. Energy 2019;174:1251–66. [6] Gholamian E, Zare V, Mousavi SM. Integration of biomass gasification with a solid oxide fuel cell in a combined cooling, heating and power system: a thermodynamic and environmental analysis. Int J Hydrogen Energy 2016;41:20396–406. [7] Rey G, Ulloa C, Cacabelos A, Barragáns B. Performance analysis, model development and validation with experimental data of an ICE-based micro-CCHP system. Appl Therm Eng 2015;76:233–44. [8] Chang H, Wan Z, Zheng Y, Chen X, Shu S, Tu Z, et al. Energy analysis of a hybrid PEMFC–solar energy residential micro-CCHP system combined with an organic Rankine cycle and vapor compression cycle. Energy Convers Manage 2017;142:374–84. [9] Chang H, Wan Z, Zheng Y, Chen X, Shu S, Tu Z, et al. Energy-and exergy-based working fluid selection and performance analysis of a high-temperature PEMFCbased micro combined cooling heating and power system. Appl Energy 2017;204:446–58. [10] Parikhani T, Azariyan H, Behrad R, Ghaebi H, Jannatkhah J. Thermodynamic and thermoeconomic analysis of a novel ammonia-water mixture combined cooling, heating, and power (CCHP) cycle. Renew Energy 2020;145:1158–75. [11] Mohammadi K, Powell K. Thermodynamic and economic analysis of different cogeneration and trigeneration systems based on carbon dioxide vapor compression refrigeration systems. Appl Therm Eng 2020;164:114503. [12] Chang H, Duan C, Xu X, Pei H, Shu S, Tu Z. Technical performance analysis of a micro-combined cooling, heating and power system based on solar energy and high temperature PEMFC. Int J Hydrogen Energy 2019;44:21080–9.
1. The gas cooler can be replaced by an internal heat exchanger for a wide range of practical operating conditions. 2. System b (in which the gas cooler is replaced by and internal heat exchanger) has significantly higher values of net output power and cooling and exergy efficiency than the system a. At a typical opeṙ , Qė and ηex for system b are reating condition, the values of Wnet spectively, 49.1%, 75.8% and 30.9% higher than the corresponding values for system a. 3. System b yields lower values of heating output and thermal efficiency as compared to system a. At a typical operating condition, the values of Q̇H and ηth for system b are respectively 39.1% and 9.2% lower than those for system a. 4. The most important advantage of proposed systems in this work as compared to the similar previously proposed systems, is the capability of producing considerable values of output power, while the previously proposed systems have very low (or even no) net output power. 13
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V. Zare and H. Rostamnejad Takleh
[13] Mosaffa AH, Farshi LG. Thermodynamic and economic assessments of a novel CCHP cycle utilizing low-temperature heat sources for domestic applications. Renew Energy 2018;120:134–50. [14] Mohammadi A, Mehrpooya M. Energy and exergy analyses of a combined desalination and CCHP system driven by geothermal energy Organic Rankine cycle. Appl Therm Eng 2017;116:685–94. [15] Zare V. A comparative thermodynamic analysis of two tri-generation systems utilizing low-grade geothermal energy. Energy Convers Manage 2016;118:264–74. [16] Ghaebi H, Parikhani T, Rostamzadeh H. A novel trigeneration system using geothermal heat source and liquefied natural gas cold energy recovery: energy, exergy and exergoeconomic analysis. Renew Energy 2018;119:513–27. [17] Takleh HR, Zare V. Employing thermoelectric generator and booster compressor for performance improvement of a geothermal driven combined power and ejectorrefrigeration cycle. Energy Convers Manage 2019;186:120–30. [18] Nemati A, Mohseni R, Yari M. A comprehensive comparison between CO2 and Ethane as a refrigerant in a two-stage ejector-expansion transcritical refrigeration cycle integrated with an organic Rankine cycle (ORC). J Supercrit Fluids 2018;133:494–502. [19] Yari M, Zare V, Mahmoudi SMS. Parametric study and optimization of an ejectorexpansion TRCC cycle integrated with a water purification system. Proc Inst Mech Eng Part A J Power Energy 2013;227:383–98. [20] Zhu Y, Li C, Zhang F, Jiang P. Comprehensive experimental study on a transcritical CO2 ejector-expansion refrigeration system. Energy Convers Manage 2017;151:98–106. [21] Liu X, Fu R, Wang Z, Lin L, Sun Z, Li X. Thermodynamic analysis of transcritical CO2 refrigeration cycle integrated with thermoelectric subcooler and ejector. Energy Convers Manage 2019;188:354–65. [22] Wang J, Zhao P, Niu X, Dai Y. Parametric analysis of a new combined cooling,
[23] [24] [25]
[26] [27]
[28] [29]
[30] [31]
[32]
14
heating and power system with transcritical CO2 driven by solar energy. Appl Energy 2012;94:58–64. Xu XX, Liu C, Fu X, Gao H, Li Y. Energy and exergy analyses of a modified combined cooling, heating, and power system using supercritical CO2. Energy 2015:414–22. Wang F, Li DY, Zhou Y. Analysis for the ejector used as expansion valve in vapor compression refrigeration cycle. Appl Therm Eng 2016;96:576–82. Ahmadi P, Dincer I, Rosen MA. Thermodynamic modeling and multi-objective evolutionary-based optimization of a new multigeneration energy system. Energy Convers Manage 2013;76:282–300. Wang J, Dai Y, Gao L. Parametric analysis and optimization for a combined power and refrigeration cycle. Appl Energy 2008;85:1071–85. Dai Y, Wang J, Gao L. Exergy analysis, parametric analysis and optimization for a novel combined power and ejector refrigeration cycle. Appl Therm Eng 2009;29:1983–90. Zare V. A comparative exergoeconomic analysis of different ORC configurations for binary geothermal power plants. Energy Convers Manage 2015;105:127–38. Balafkandeh S, Zare V, Gholamian E. Multi-objective optimization of a tri-generation system based on biomass gasification/digestion combined with S-CO2 cycle and absorption chiller. Energy Convers Manage 2019;200:112057. Huang BJ, Chang JM, Wang CP, Petrenko VA. A 1-D analysis of ejector performance. Int J Refrig 1999;22:354–64. Takleh HR, Zare V. Performance improvement of ejector expansion refrigeration cycles employing a booster compressor using different refrigerants: thermodynamic analysis and optimization. Int J Refrig 2019;101:56–70. Chang H, Xu X, Shen J, Shu S, Tu Z. Performance analysis of a micro-combined heating and power system with PEM fuel cell as a prime mover for a typical household in North China. Int J Hydrogen Energy 2019;44:24965–76.