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Thermodynamic analysis of a novel combined cooling, heating and power system driven by solar energy
Accepted Manuscript Thermodynamic analysis of a novel combined cooling, heating and power system driven by solar energy Beneta Eisavi, Shahram Khalilarya, Ata Chitsaz, Marc A. Rosen PII: DOI: Reference:
S1359-4311(16)34525-2 https://doi.org/10.1016/j.applthermaleng.2017.10.132 ATE 11330
To appear in:
Applied Thermal Engineering
Received Date: Revised Date: Accepted Date:
31 December 2016 9 September 2017 22 October 2017
Please cite this article as: B. Eisavi, S. Khalilarya, A. Chitsaz, M.A. Rosen, Thermodynamic analysis of a novel combined cooling, heating and power system driven by solar energy, Applied Thermal Engineering (2017), doi: https://doi.org/10.1016/j.applthermaleng.2017.10.132
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Thermodynamic analysis of a novel combined cooling, heating and power system driven by solar energy Beneta Eisavi a, Shahram Khalilarya a, Ata Chitsazb,1, Marc A.Rosenc
a
Mechanical Engineering Department, Faculty of Engineering, Urmia University, Urmia, Iran b
c
Faculty of Mechanical Engineering, Urmia University, Urmia, Iran
Faculty of Engineering and Applied Science, University of Ontario Institute of Technology, 2000 Simcoe Street North, Oshawa, Ontario, L1H 7K4, Canada
Abstract A novel combined cooling, heating and power system (CCHP) driven by solar energy is proposed. The system applies an organic Rankine cycle, a double effect lithium bromide-water absorption refrigeration system, and heat exchangers to generate electricity, cooling and heating, respectively. Energy and exergy analyses are conducted to determine system efficiencies and losses. The cycle is compared with similar combined cycle that differs by using a cooling subsystem having a single effect absorption chiller. The effects of some key thermodynamic parameters on the performance of the cycle are examined. The results reveal that, for the same amount of input heat, the usage of a double effect absorption refrigeration system instead of a single effect absorption chiller can raise the amount of cooling power up to 48.5% and consequently can improve the performance of the system. At the same time, heating power rises by 20.5%, 1
Corresponding author: Tel: +989143636112
E–mail address: [email protected]
1
resulting in an increase in the cogeneration heat and power efficiency to 96.0%, while net electrical power production declines by 27%. In addition, it is determined that a high exergy destruction rate occurs in the solar collectors.
Keywords: Combined cooling, heating and power (CCHP); Solar energy; Organic Rankine Cycle; Double effect lithium bromide-water absorption refrigeration; Exergy 1. Introduction In recent years, increasing energy demands and the dependence of industry on fossil fuels and, in many instances, low energy efficiency systems contribute not only to the depletion of nonrenewable energy resources, but to various environmental problems, such as global warming and air pollution. Combined cooling, heating and power (CCHP), which simultaneously produces diverse types of energy (electricity, heating and cooling) from the same primary energy resource, can provide an alternative energy option which has high efficiency and helps address these challenges. A CCHP unit consists of a prime mover (e.g., steam turbine, microturbine, fuel cell, organic Rankine cycle), a heat recovery system (typically a heat exchanger that recovers waste heat), and thermally activated equipment (such as an absorption or adsorption chiller) [1]. Much research has been carried out on combined cycles. Xu et al. [2] proposed a combined cooling and power system, comprised of a Rankine cycle and an absorption chiller. Minciuc et al. [3] presented a method for analyzing CCHP systems. The performance of a cogeneration system based on a gas turbine has been examined [4]. Ho et al. [5] examined the total efficiency of a CCHP system under a diverse range of operating conditions. Hwang et al. [6] proposed a CCHP cycle based on a micro-turbine using an absorption chiller as a cooling
2
system, and explored the cycle’s energy savings and economy feasibility. Rosen et al. [7] performed exergy and energy analyses for a CCHP system. Kanoglu and Dincer [8] analyzed the performance for various configurations of CCHP systems which were designed to meet building energy demands. Most of combined cycles in that study included steam turbines, gas turbines, diesel engines and geothermal energy systems. In addition, several arrangements of CCHP systems have been examined based on first and second laws [9-14]. A brief review of micro-trigeneration system is conducted by Sonar et al [15]. In that study it is discussed that how integrated energy systems offer key solution to environmental problems and inefficient energy conversion systems. Basic elements of micro CCHP system along with their pros and cons were discussed. In addition application of integrated energy systems in energy sectors, alternative fuels in micro-trigeneration were investigated. The literature devoted to the study CCHP systems driven by solar energy are relatively rare. Buck and Friedmann [16] studied the performance of a CCHP based on a micro-turbine, with a cooling system comprised of a single effect or double effect absorption chiller. Better performance was observed for the CCHP system with a double effect absorption chiller. Medrano et al. [17] proposed a solar CCHP system using natural gas as an auxiliary fuel and analyzed the cycle based on the first and second laws of thermodynamics. Fumo et al. [18] performed energy and economic analyses of a solar CCHP system. Wang et al. [19] proposed a solar CCHP system and explored the effects of some key parameters on its performance. The authors also performed an optimization to achieve the maximum exergy efficiency. AlSulaiman et al. [20] presented a CCHP system driven by solar energy, and performed energy and exergy analyses to assess the effect of key parameters on the performance of the cycle. Thilak Raj et al. [21] carried out a review on the cogeneration technologies based on renewable energy sources, including biomass, solar energy, fuel cell and waste heat recovery
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in which the various researches around them were classified and discussed based on study, design, modeling approach, environmental and economic issues. The most common working fluid combinations used in absorption refrigeration systems are lithium bromide-water (LiBr-H2O) and ammonia-water (NH3-H2O). The NH3-H2O system is more complicated than the LiBr-H2O system, since it needs a rectifying column that assures no water vapors enter the evaporator. Moreover, the refrigerant used in LiBr-H2O systems is water which is everywhere available, inexpensive and no toxic, In addition, the LiBr-H2O system can operate at a low generator temperature with better coefficient of performance than NH3-H2O system and are commercially available for air conditioning applications [22, 23]. Absorption chillers exist in single-effect and multiple effect designs although in multipleeffect designs, the double-effect one is more conventional. The double effect absorption refrigeration system was introduced in 1956-1958 [24]. Since then, much literature has been devoted to the study of the performance of these refrigerators for various working fluids [2531]. For instance, Gomri and Hakimi [31] applied energy and exergy analyses to study the performance of a double effect water-lithium absorption chiller. Also, Garousi et al. [32] conducted exergy and energy analyses for several arrangements including series, parallel and reverse parallel of absorption systems. One of the major design choices in double effect technology is the choice of how to connect the solution circuits. The basic options are parallel or series flow. Performance analysis showed that although COP of the series flow configuration is lower than the parallel one but its capacity is higher. In addition, in series flow the solution path is simpler. Marinating the correct flow split in parallel flow is problematic without active controls. Thus particularly for small systems, series flow is attractive as a design compromise. [33] To the best of the authors’ knowledge, investigations of CCHP systems based on solar energy using absorption chillers for refrigeration are scarce but needed to better understand such
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systems and to increase the chances of their beneficial exploitation. Hence, in this paper, the following are considered:
A thermodynamic simulation and analysis is conducted of a CCHP system that uses solar energy as a heat source and applies an absorption chiller to produce cooling power.
The CCHP considered follows the system studied by Al-Sulaiman et al. [20], but with several differences, especially altering the cooling subsystem to accommodate a double effect absorption refrigeration system. In addition, another heat exchanger for producing thermal heat is considered.
The effect of this replacement is investigated thermodynamically. Furthermore, the proposed cycle is examined only for the solar mode,
The effects of several key thermodynamic parameters on the performance of the system are examined.
Exergy destruction rate of each device is evaluated.
2. System description A schematic diagram of the proposed CCHP system is given in Fig. 1. It consists of a solar collector field which provides the primary energy input of the combined cycle, an organic Rankine cycle that produces electricity, heat exchangers that generate steam from waste heat in heating process I and II, and series-flow double effect LiBr-H2O absorption chiller that provides cooling power. The operation of proposed system can be summarized as follows:
The heat transfer fluid in the solar collector domain which is Therminol-66 oil has a low relative pressure and its pressure is not sensitive to the increase in the temperature is selected according to [20] and some of its properties are brought in
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Table 1 [34]. It passes through parabolic trough solar collectors. The absorbed heat is transferred to the ORC working fluid through the ORC evaporator. The flow exiting the ORC evaporator is used to produce steam in thermal heat exchanger II.
In the ORC, at the exit of the high pressure generator (HPG), n-octane as an organic fluid at the saturation state is pumped to the turbine inlet pressure. After passing through the ORC evaporator, the fluid becomes superheated then is expanded in the ORC turbine to produce mechanical energy which is converted to electricity by the generator. After, the flow passes through thermal heat exchanger I, where steam is generated. Then, the organic fluid enters the high pressure generator (HPG) in order to drive the absorption chiller.
In the series-flow double effect LiBr-H2O absorption chiller, the weak solution leaving the absorber is pumped to the HPG, passing through the low temperature heat exchanger (LTHE) and the high temperature heat exchanger (HTHE). Then the heat transferred to the HPG is utilized to boil out refrigerant vapor (primary refrigerant vapor) from the weak solution entering the HPG. In the LPG, the primary refrigerant vapor is condensed due to the low temperature of strong solution entering the LPG after passing through HTHE and EV-4. The refrigerant latent heat is consumed in producing the secondary refrigerant vapor from the strong solution, making the solution stronger. The condensed primary refrigerant vapor -after passing through EV-2- and the secondary refrigerant vapor both flow into the condenser to reject heat. At the exit of the condenser, the total amount of liquid refrigerant is delivered to the evaporator via EV-1 to generate the cooling effect. In the absorber, the refrigerant vapor leaving the evaporator is mixed with the strong solution entering the absorber, which comes from LPG after passing through LTHE and EV-3, thereby rejecting its heat of absorption.
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3. Mathematical modeling In order to simplify the mathematical modeling of the proposed system based on the conservation principles of mass and energy, some assumptions are invoked. The system is taken to operate at steady state conditions, and pressure drops due to friction and heat losses across all the components of the CCHP are treated as negligible according to ref [20]. The ORC turbine and pumps have fixed isentropic efficiencies. It is worth nothing that as the main purpose is to investigate the effect of changing cooling subsystem with more efficient one the dynamic modeling was not conducted although it can be regarded in another study. Also, the series-flow double effect absorption refrigeration system, which uses LiBr-H2O as a working fluid, functions based on the following assumptions:
The refrigerant is assumed saturated exiting the condenser and the evaporator [31].
The weak solution leaving the absorber and the strong solution leaving the HPG and LPG are assumed to be saturated and at equilibrium conditions at their respective temperatures. [31]
The reference environment state for the system is taken to be
and
.
3.1. General analysis At steady state conditions, mass rate balances for the total mass and for the mixture of LiBr-H2O for the system components can be written as: (1) (2) where x is the concentration of lithium bromide in solution. Neglecting kinetic and potential energy, the first law of thermodynamics, which embodies the the law of conservation of energy, can be expressed on a rate basis as follows:
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(3) Here,
denotes the heat transfer rate and
the work transfer rate.
An energy balance provides no information on the degradation of energy or resources during a process and does not quantify the usefulness or quality of the various energy and material streams flowing through a system and exiting as products and wastes. Therefore, the exergy method of analysis which overcomes the limitations of the first law of thermodynamics is considered here in addition to energy analysis. Exergy is maximum useful work that can be obtained from the amount of energy or the flow of materials in a specified reference environment [35]. An exergy rate balance for a steady state system can be written as: (4)
Neglecting kinetic and potential exergy, the physical exergy can be expressed as follows: (5) According to the results achieved by Avanessian et al. [36], the error caused by neglecting the chemical exergy of the solution is very low and does not affect results; therefore, chemical exergy is neglected for all flows and components here. The simulation and analysis of the combined cycle involves writing appropriate mass, energy and exergy rate balances based on the stated assumptions for each of the components. This information is presented in Table 2. 3.2. Overall analysis The performance of the combined cycle is analyzed and the energy and exergy efficiencies are determined. The energy efficiency is a measure expressing the rate of useful output energy to the rate of input energy of the system, and is defined as:
8
(6) where
denotes the electrical efficiency,
the net electrical power output and
the
heat input rate to the combined cycle provided by solar energy. Also, (7)
The combined heating and power efficiency can be expressed as: (8)
where
, the total rate of heating produced by the CCHP, is calculated by: (9)
The combined cooling and power efficiency can be written as: (10)
where
, the cooling effect generated by the absorption system, can be expressed as
follows: (11) The electrical to heating power ratio can be defined as: (12) and he electrical to cooling power ratio as: (13) The electrical exergy efficiency can be expressed as: (14) the combined heating and power exergy efficiency as:
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(15)
and the combined cooling and power exergy efficiency as: (16)
Here
denotes the exergy rate input to the solar collector [20].
The equations used to model the solar field can be found in Refs. [20, 37]. 4. Validation In order to simulate the mathematical modeling, a computer program has been developed in Engineering Equation Solver (EES). The thermodynamic properties of LiBr-H2O which used are based on the new set of formulations presented by Patek and Klomfar [38], which are as internal functions in EES. In addition, n-octane thermodynamic properties were obtained using Functions defined in EES based on Fundamental Equation of State, as described by[39]. Since there was lack in experimental data in literature to use in validation of present model of the absorption chiller, the alternative approach was adopted, in which the obtained results of the present model are compared with those reported by Gomri [31]. There is a good agreement between the results from the two works, as seen in Table 3. The proposed system which is presented in Fig. 1 has been modeled and simulated based on the input data given in Table 4. 5. Results and discussion The performance of the CCHP system depends on the performance of each unit as well as the way the units are combined. Thermodynamic properties of the proposed cycle for the base case operating conditions are reported in Table 5.
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The exergy destruction rates of various components of the proposed cycle which is presented in Fig. 2 provide important information about the sources of irreversibilities. Note that in this figure only components with higher exergy destruction rates are shown. The highest exergy destruction rate is attributable to the solar collectors which is affected by heat transfer irreversibilities between the sun and the collector, between the collector and the ambient air and inside the collector [40] and is found to be 1280 kW. The second highest exergy destruction rate occurs in the ORC evaporator due to high temperature difference and is determined to be 650 kW. In Table 6, the results of the thermodynamic modeling and simulation for present work are compared with the corresponding results reported by [20]. Relative to a single effect absorption chiller, the double effect absorption chiller operates with a higher temperature heat source. Therefore, the inlet temperature of the ORC pump needs to be increased. This measure leads to a decline of 27% in the net electrical power production, due to decrease in differences between lower and upper temperature of ORC cycle, consequently, electrical energy and exergy efficiency decrease, while cooling production under the same input power is increased by 48.5% because of the fact that an absorption chiller with higher coefficient of performance is used and the heating power rises by 20.5% .In fact, increasing ORC pump inlet temperature, decreases input energy to the ORC cycle although it brings about decrease in heating process I power, it increases the amount of available energy to the heating process II. As a result, the cogeneration heat and power production also increases. On the other hand, because of high exergy destruction rate occurred in heat exchanger of heating process II combined heating and power exergy efficiency declines. It is also observed that combined cooling and power efficiency is declined in comparison to base case which is due to the fact that the amount of cooling power is smaller than power production in size. Therefore, its increase could not cover the decline in power production decreases.
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In the following subsections, the effects of several key parameters, such ORC evaporator pinch point temperature, ORC pump inlet temperature and turbine inlet pressure, are studied. 5.1. Effect of ORC evaporator pinch point temperature Although an evaporator with a lower pinch point temperature (Tpp) typically has higher performance, it incurs increased cost and design requirements. So, studying the effects of this change is an important factor, as demonstrated in Figs. 3-5. In Fig. 3 the effect of ORC evaporator pinch point temperature on the system outputs electrical to heating power ratio (Rh) and the electrical to cooling power ratio (Rc) are presented. It is observed that, with increasing Tpp, both net electrical power and thermal process I power output are reduced while thermal process II power output is increased. In fact, with increasing ORC evaporator pinch point temperature, the amount of input heat to the ORC evaporator is reduced and consequently the inlet temperature and inlet enthalpy of the ORC turbine are decreased. Referring to Fig. 3, heating process I power output decreases and heating process II power output increases from 4086 kW to 3313 kW and 898.4k W to 1745 kW, respectively. In addition, the net electrical power output decreases from 556.4 kW to 476.7 kW. Since in this study the fluid properties at points 5 and 1 control the cooling capacity, the only possible change for point 5 is pressure. Also, the temperature, pressure and mass flow rate of points 5 and 1 are identical. Therefore, the amount of cooling power output remains constant as the ORC evaporator pinch point temperature is varied. It is seen from Fig. 3 that both Rh and Rc decrease with increasing ORC evaporator pinch point temperature, and this is primarily due to the reduction of net electrical power. It is also observed that the electrical to heating power ratio decreases from 0.11 to 0.09 and similarly the electrical to cooling power ratio decreases from 2.36 to 2.03. Note that, as the effect of net
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heating process power output is more dominant than the effect of net electrical power, Rh changes only slightly. Fig. 4 shows the changes of electrical, combined cooling and power, and combined heating and power efficiencies as ORC evaporator pinch point temperature varies. According to this figure, the electrical efficiency declines from 9.6% to 8.3% as Tpp changes from 5 30
to
. The reason of this drop is the reduction in the rate of heat input to the ORC evaporator,
which leads to a decrease in the turbine inlet temperature and consequently a decrease in turbine power output. Also, as stated before, when the ORC evaporator pinch point temperature increases, the available heat rate at the turbine exit declines, leading to a decrease in the heating process I power output; however, the heating process II power output increases, due to the decrease of ORC evaporator inlet energy. Therefore, the combined heating and power efficiency slightly increases from 96.0% at 5
to 96.1% at 30
as ORC evaporator
pinch point temperature increases. In contrast, the efficiency of the combined cooling and power system is reduced from 13.7% to 12.3%, as Tpp changes from 5
to 30
, mainly
because of the decrease in the net electrical power production. Fig.4 also shows the effect of Tpp on exergy efficiencies. With increasing T pp, the electrical, combined cooling and power and combined heating and power exergy efficiencies of the proposed system are seen to decline. The electrical exergy efficiency and the combined cooling and power exergy efficiency are observed to decrease from 4.8% to 4.1% and from 4.9% to 4.2%, respectively. In addition, the exergy efficiency of the combined heating and power system due to the reduction of input heat to the ORC evaporator decreases from 13.4% to 12.6%. Fig. 5 shows the variation of exergy destruction rate of the cycle components as ORC evaporator pinch point temperature varies from 5
to 30
. It is observed that the highest
rate of exergy destruction rate is related to the solar collectors and is found to be 1280 kW.
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Due to the fixed operating temperature of the solar collectors, their exergy destruction rate is constant. Referring to Fig. 5, the exergy destruction rate of the ORC evaporator decreases from 694 kW to 625 kW as ORC evaporator pinch point temperature changes from 5
to 30
, due to the reduction of the exergy difference between the entering and exiting flows of the ORC evaporator. In addition, the exergy destruction rate of the heat exchangers for thermal processes I and II decrease from 390 kW to 264 kW and increase from 377 to 746 kW, respectively. As Tpp increases, the exergy destruction rate of the heat exchanger of heating process II changes significantly. Therefore, more care should be exercised in the selection and design of this component. It is also observed from Fig. 5 that the exergy destruction rate of the turbine decreases from 89 kW to 83 kW, with increasing Tpp from 5
to 30
which results
in lower available exergy. 5.2.Effect of ORC pump inlet temperature The effect of varying ORC pump inlet temperature (T1) on the performance of the cycle is described in Figs. 6-8. The temperature of ORC pump inlet temperature is selected to vary from 114
to 145
since in higher temperature of T1, crystallization phenomenon which is
probable in the weak solution entering the absorber is going to happen. [41] Fig. 6 shows the effect of ORC pump inlet temperature variation on system outputs, Rh and Rc. A variation in T1 has a direct impact on the amount of electrical power production due to the fact that with increasing T1, the difference between the minimum and maximum temperatures of the ORC decreases, which causes the electrical power production to decrease from 566 kW at 114
to 387 kW at 145
. Also, for the same reason, the thermal process I
power output is reduced from 3423 kW to 3298 kW. Referring to Fig. 6, an increase of ORC pump inlet temperature is seen to result in an ORC input heat rate decrease, which causes the thermal process II power output to increase from 1231 kW to 1883 kW. In addition, the ORC
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pump inlet temperature rise causes a reduction in the rate of heat transfer to the HPG. It is also observed from Fig.6 that Rc decreases from 2.4 to 1.6. This drop is primarily due to the significant reduction in electrical power production as the ORC pump inlet temperature increases from 114
to 145
. In addition, as the ORC pump inlet temperature rises, as
mentioned earlier, a significant reduction is observed in the electrical power output and an increase in the thermal process II power output, which causes Rh to decrease from 0.12 at 114 to 0.07 at 145
.
Fig. 7 depicts the effect of ORC pump inlet temperature on the electrical, combined cooling and power and combined heating and power efficiencies. It is observed from Fig. 7 that the electrical efficiency decreases from 9.8% to 6.7%, mainly because of net electrical power output reduction as the ORC pump inlet temperature increases from 114
to 145
.
Similarly, the net electrical power reduction along with thermal process II power output rise cause the combined cooling and power efficiency to decrease from 14% at 114 145
to 10.8% at
. But, the combined heating and power efficiency increases slightly, from 90.5% to
96.5%. The reasons for this phenomenon are the decrease and increase of net electrical power output and thermal process II power output. Fig. 7 also shows the effect of varying ORC pump inlet temperature on the electrical, combined heating and power and combined cooling and power exergy efficiencies. An increase of the ORC pump inlet temperature is observed to lead to an exergy efficiency reduction owing to the decrement of net electrical power. Based on reasons mentioned earlier, as the ORC pump inlet temperature increases from 114
to
145 , the electrical exergy efficiency decreases from 4.9 % to 3.4 %, while the combined cooling and power and combined heating and power exergy efficiencies decrease from 5% to 3.4% and from 12.6% to 12%, respectively.
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The impact of varying pump inlet temperature on the exergy destruction rate of components is presented in Fig. 8. The exergy destruction rate of the solar collectors is observed to be the highest of all components. With increasing T1 from 114
to 145
the
exergy destruction rate of the evaporator decreases, from 691 kW to 567 kW, owing to the lower available exergy in the ORC evaporator. Also, because of the increase in available energy of thermal process I, the exergy destruction rate of the heat exchanger assigned to thermal process I increases from 268 kW to 408 kW. In addition, the exergy destruction rate of the heat exchanger for thermal process II rises approximately by 286 kW (from 517 kW to 803 kW), due to increase of exergy difference. 5.4.Effect of turbine inlet pressure The analysis results demonstrate that the effect of varying turbine inlet pressure on system performance is insignificant. Therefore, it is not considered further here. Parametric analysis showed how system performance is affected by different key parameters. Critical components with higher exergy destruction rate are recognized and it is investigated that for having higher exergy efficiency for critical components including ORC evaporator, heating process II heat exchanger, heating process I heat exchanger and turbine; the cycle should work under specific conditions which lead to obtain higher exergy efficiency for each component and are presented as follows:
Parametric analysis showed that when cycle operates at lower pinch point temperature of ORC evaporator higher exergy efficiency is obtained for ORC evaporator and turbine while exergy efficiency for heat exchangers of heating process I and II is maximized at higher Tpp.
Moreover, it is deduced that when cycle operates in higher ORC pump inlet temperature, ORC evaporate, heat exchanger of heating process II and turbine operate
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with higher exergy efficiency whereas it happens for heat exchanger of heating process I when ORC pump inlet temperature is minimum. 6. Conclusions A novel combined cooling heating and power system which is driven by solar energy has been successfully assessed. The main foci of the study were determining the impact of changing the refrigeration subsystem to one with a higher coefficient of performance and analyzing based on first and second laws of thermodynamics the effect on the cycle performance of important working parameters, such as ORC evaporator pinch point temperature, ORC pump inlet temperature and turbine inlet pressure. The main results and the conclusions that can be drawn from them are as follows:
Relative to a single effect absorption chiller, a series-flow double effect absorption chiller increases the cooling power by 48.5% for the same input heat.
For the base case analysis, the electrical efficiency is 8.9%, the combined heating and power efficiency is 96.0% and the combined cooling and power efficiency is 12.9%. Correspondingly, the electrical, the combined heating and power and combined cooling and power exergy efficiencies respectively are determined to be 4.4%, 12.8% and 4.5%. The significant difference between the energy and exergy efficiencies is due to the high exergy destruction rate destroyed by solar collectors.
The main sources of exergy destruction rate are the solar collector and the ORC evaporator, implying that these two components should be carefully selected and designed.
The electrical to the cooling power ratio is sensitive to the changes in the ORC pump inlet temperature. Controlling this parameter can be useful factor improving the electrical power amount.
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The turbine inlet pressure has little impact on the performance of the CCHP. Therefore, the ORC can operate at lower pressures, reducing capital costs of the cycle.
Raising the ORC evaporator pinch point temperature has negative effect on the turbine and thermal processing I outputs, while it has the opposite effect on the thermal process II output. In general, increasing the ORC evaporator pinch point temperature reduces the energy and exergy efficiencies and, due to the reduction of thermal and electrical power outputs, the electrical to heating power and electrical to cooling power ratios.
An ORC pump inlet temperature increase has the greatest impact on the performance of the system, reducing the electrical and the heating power of thermal process I while increasing the output of thermal process II power. Also, raising the ORC pump inlet temperature, due to increase of ORC pump inlet temperature, reduces the energy and exergy efficiencies.
Nomenclature CCHP
combined cooling heating and power specific exergy (kJ/kg) specific enthalpy (kJ/kg)
HPG
high pressure generator
HTHE
high temperature heat exchanger
LPG
low pressure generator
LTHE
low temperature heat exchanger mass (kg)
ORC
organic Rankine cycle
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heat (kJ) heat rate (kW) electrical to cooling power ratio electrical to heating power ratio temperature (
)
work rate (kW) x
lithium bromide concentration (%)
Greek symbols efficiency
Subscripts 0
atmospheric condition combined cooling and power combined heating and power collector exit electrical evaporator exergy generator heating inlet inlet
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water
Superscripts .
rate physical
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[15] Sonar, D., Soni, S. L., & Sharma, D. (2014). Micro-trigeneration for energy sustainability: Technologies, tools and trends. Applied Thermal Engineering, 71(2), 790796. [16] Buck, R., & Friedmann, S. (2007). Solar-assisted small solar tower trigeneration systems. Journal of Solar Energy Engineering, 129(4), 349-354. [17] Medrano, M., Castell, A., Fontanals, G., Castellón, C., & Cabeza, L. F. (2008). Economics and climate change emissions analysis of a bioclimatic institutional building with trigeneration and solar support. Applied Thermal Engineering, 28(17), 2227-2235. [18] Fumo, N., Chamra, L. M., & Bortone, V. (2009, January). Potential of solar thermal energy for CCHP systems. In ASME 2009 3rd International Conference on Energy Sustainability collocated with the Heat Transfer and InterPACK09 Conferences (pp. 5763). American Society of Mechanical Engineers. [19] Wang, J., Dai, Y., Gao, L., & Ma, S. (2009). A new combined cooling, heating and power system driven by solar energy. Renewable Energy, 34(12), 2780-2788. [20] Al-Sulaiman, F. A., Dincer, I., & Hamdullahpur, F. (2011). Exergy modeling of a new solar driven trigeneration system. Solar Energy, 85(9), 2228-2243. [21] Raj, N. T., Iniyan, S., & Goic, R. (2011). A review of renewable energy based cogeneration technologies. Renewable and Sustainable Energy Reviews, 15(8), 36403648. [22] Seraj, M., & Siddiqui, M. A. (2013). Performance analysis of parallel flow single and double effect absorption cycles. International Journal of Innovative Research in Science, Engineering and Technology, 2, 1570-1576.
22
[23] Kalogirou, S. (2008). Recent patents in absorption cooling systems. Recent Patents on Mechanical Engineering, 1(1), 58-64. [24] Lawson, M. B., Lithgow, R. A., & Vliet, G. C. (1982). Water-lithium bromide doubleeffect absorption cooling cycle analysis. ASHRAE Trans.;(United States), 88(CONF820112-). [25] Xu, G. P., & Dai, Y. Q. (1997). Theoretical analysis and optimization of a double-effect parallel-flow-type absorption chiller. Applied Thermal Engineering, 17(2), 157-170. [26] Arun, M. B., Maiya, M. P., & Murthy, S. S. (2000). Equilibrium low pressure generator temperatures for double-effect series flow absorption refrigeration systems. Applied Thermal Engineering, 20(3), 227-242. [27] Arun, M. B., Maiya, M. P., & Murthy, S. S. (2001). Performance comparison of doubleeffect parallel-flow and series flow water–lithium bromide absorption systems. Applied Thermal Engineering, 21(12), 1273-1279. [28] Liu, Y. L., & Wang, R. Z. (2004). Performance prediction of a solar/gas driving double effect LiBr–H2O absorption system. Renewable Energy, 29(10), 1677-1695. [29] Torrella, E., Sánchez, D., Cabello, R., Larumbe, J. A., & Llopis, R. (2009). On-site realtime evaluation of an air-conditioning direct-fired double-effect absorption chiller. Applied Energy, 86(6), 968-975. [30] Ravikumar, T. S., Suganthi, L., & Samuel, A. A. (1998). Exergy analysis of solar assisted double effect absorption refrigeration system. Renewable Energy, 14(1), 55-59. [31] Gomri, R., & Hakimi, R. (2008). Second law analysis of double effect vapour absorption cooler system. Energy Conversion and Management, 49(11), 3343-3348.
23
[32] Garousi Farshi, L., Seyed Mahmoudi, S. M., Rosen, M. A., & Yari, M. (2012). A comparative study of the performance characteristics of double‐effect absorption refrigeration systems. International Journal of Energy Research, 36(2), 182-192. [33] Herold, K. E., Radermacher, R., & Klein, S. A. (2016). Absorption chillers and heat pumps. CRC press. [34] Therminol.therminol-66 heat transfer fluid, www.therminol.com;2010. [35] Dincer I, Rosen MA. (2013) Exergy: Energy, environment and sustainable development. 2nd edition Elsevier. [36] Avanessian, T., & Ameri, M. (2014). Energy, exergy, and economic analysis of single and double effect LiBr–H2O absorption chillers. Energy and Buildings, 73, 26-36. [37] Chen, X., Gong, G., Wan, Z., Zhang, C., & Tu, Z. (2016). Performance study of a dual power source residential CCHP system based on PEMFC and PTSC. Energy Conversion and Management, 119, 163-176. [38] Patek, J., & Klomfar, J. (2006). A computationally effective formulation of the thermodynamic properties of LiBr–H2O solutions from 273 to 500 K over full composition range. International Journal of Refrigeration, 29(4), 566-578. [39] Span, R., & Wagner, W. (2003). Equations of state for technical applications. II. Results for nonpolar fluids. International Journal of Thermophysics, 24(1), 41-109. [40] Tyagi, S. K., Wang, S., Singhal, M. K., Kaushik, S. C., & Park, S. R. (2007). Exergy analysis and parametric study of concentrating type solar collectors. International Journal of Thermal Sciences, 46(12), 1304-1310.
24
[41] Farshi, L. G., Mahmoudi, S. S., & Rosen, M. A. (2011). Analysis of crystallization risk in double effect absorption refrigeration systems. Applied Thermal Engineering, 31(10), 1712-1717. [42] Xu, D., Liu, Q., Lei, J., & Jin, H. (2015). Performance of a combined cooling heating and power system with mid-and-low temperature solar thermal energy and methanol decomposition integration. Energy Conversion and Management, 102, 17-25. [43] Karellas, S., & Braimakis, K. (2016). Energy–exergy analysis and economic investigation of a cogeneration and trigeneration ORC–VCC hybrid system utilizing biomass fuel and solar power. Energy Conversion and Management, 107, 103-113.
25
Table 1 Typical properties of Therminol-66 Source: [34] Substance name Composition Density at 25 C Kinematic viscosity at 40 C Normal boiling point Optimum use range Pseudocritical temperature Pseudocritical pressure
Therminol-66 Modified terphenyl 1005 kg/m3 29.6 mm2/s 359 C 0-345 C 569 C 24.3 bar
26
Table 2 Relations of thermodynamic rate balances for system components. Component
Energy rate balance
ORC turbine Heating process I
ORC pump ORC evaporator
Heating process II
Solar pump
HPG
LPG
Condenser Evaporator Absorber LTHE
HTHE Solution pump EV-1 EV-2 EV-3
27
Exergy rate balance
Table 3 Comparison of present modeling and simulation results of absorption chiller with those of Gomri [31]. State point Gomri Present
Gomri
work
Present
Gomri
work
Present work
Present work
1
35
35
0.12702
0.127
0
0
146.6
2
4
4
0.12702
0.127
0
0
146.6
3
4
4
0.12702
0.127
0
0
2508
4
35
35
1.73674
1.737
55.869
55.87
87.59
5
35
35
1.73674
1.737
55.869
55.87
87.62
6
62.49
62.66
1.73674
1.737
55.869
55.87
143.4
7
106.81
107.2
1.73674
1.737
55.869
55.87
236.4
8
130
130
1.67132
1.672
58.056
58.06
288.6
9
76.09
82.86
1.67132
1.672
58.056
58.06
192
10
76.09
82.86
1.67132
1.672
58.056
58.06
192
11
130
130
0.06542
0.06544
0
0
2741
12
82.46
82.46
0.06542
0.06544
0
0
345.2
13
35
35
0.06542
0.06544
0
0
345.2
14
80
80
0.06160
0.06161
0
0
2650
15
80
80
1.60972
1.61
60.278
60.28
195.8
16
48.67
48.5
1.60972
1.61
60.278
60.28
135.6
17
48.67
48.5
1.60972
1.61
60.278
60.28
135.6
28
Table 4 Conditions of simulation for the proposed CCHP. Term
Value
ORC turbine efficiency
80 % a
ORC pump efficiency
80 % a 2000 kPaa
Baseline turbine inlet pressure ORC pump inlet temperature
120
Electrical generator efficiency
95 % a
Electrical motor efficiency
95 % a
Solution pump efficiency
95 % b
Pinch point temperature of ORC evaporator
20
Effectiveness of solution heat exchangers
70 % b 120 o C b
HPG temperature Condensation temperature
35 o C b
Absorber temperature
35 o C b
Evaporation temperature
7 oC
Minimum heat transfer temperature in LPG
5 oC b
b
500 W/m2a
Solar radiation a [20] b [31]
29
Table 5 Thermodynamic properties of cycle states. State point 1
120
86.81
-
229.5
0.6659
217.4
7
2
121
2000
-
233.4
0.6679
240.3
7
3
283.6
2000
-
856
1.941
1942
7
4
223.4
86.81
-
774.8
1.982
1287
7
5
120
86.81
-
257.9
0.7381
265.5
7
6
120
44.72
0
2722
7.849
19.01
0.04918
7
78.58
44.72
0
329
1.058
0.8851
0.04918
8
35
5.627
0
329
1.097
0.3219
0.04918
9
73.58
5.627
0
2637
8.574
4.319
0.05027
10
35
5.627
0
146.6
0.505
0.05938
0.09945
11
7
1.002
0
146.6
0.5246
-0.5221
0.09945
12
7
1.002
0
2513
8.973
-15.6
0.09945
13
120
44.72
0.5563
263.2
0.7133
94.21
1.71
14
77.71
44.72
0.5563
173.8
0.4744
63.21
1.71
15
77.71
5.627
0.5563
173.8
0.4744
63.21
1.71
16
73.58
5.627
0.5731
170.6
0.4359
75
1.66
17
46.57
5.627
0.5731
116.7
0.2739
65.64
1.66
18
46.57
1.002
0.5731
116.7
0.2739
65.64
1.66
19
35
1.002
0.5408
81.15
0.2184
36.21
1.759
20
35
44.72
0.5408
81.18
0.2184
36.26
1.759
30
21
59.59
44.72
0.5408
132.1
0.3776
42.32
1.759
22
100.3
44.72
0.5408
218.9
0.6225
66.67
1.759
23
326.9
54.05
-
1602
1.606
5006
7.1
24
186.8
1.497
-
987.8
0.9701
1994
7.1
25
131.9
0.1997
-
788.8
0.7097
1132
7.1
26
131.9
0.1997
-
788.8
0.7097
1132
7.1
* * * *
*The data in the literature and EES can calculate only vapor pressure of therminol-66.
31
Table 6 Comparison of present results with those of [20]. Parameter COP
Present work
ref [20]
1.18
0.79
Net electrical power
511.4
Cooling power
235.4
158.5
Heating power
5031
4172.7
700
Electrical efficiency
8.9 %
12.5 %
Combined heating and power efficiency
96.0 %
88 %
Combined cooling and power efficiency
12.9 %
15.5 %
Electrical exergy efficiency
4.4
6.2 %
Combined heating and power exergy efficiency
12.8 %
17.6 %
Combined cooling and power exergy efficiency
4.5 %
6.4 %
32
Fig. 1. Schematic of CCHP system.
33
Fig. 2. Exergy destruction rate of components of the proposed cycle (P 3=2000 kPa, T1=120 , Tpp=20
34
).
Fig. 3. Effect of ORC evaporator pinch point temperature (T pp) on cycle outputs, Rc and Rh at P3=2000 kPa, T1=120
35
.
Fig. 4. Effect of ORC evaporator pinch point temperature (T pp) on efficiencies at P3=2000 kPa, T1=120
36
.
Fig. 5. Effect of ORC evaporator pinch point temperature (T pp) on exergy destruction rate of components at P3=2000 kPa, T1=120
37
.
Fig. 6. Effect of ORC pump inlet temperature (T1) on cycle outputs, Rh and Rc at P3=2000 kPa, Tpp=20
38
.
Fig. 7. Effect of ORC pump inlet temperature (T1) on efficiencies at P3=2000 kPa, Tpp=20
39
Fig. 8. Effect of ORC pump inlet temperature (T1) on exergy destruction rate of components at P3=2000 kPa, Tpp=20
40
.
S1359-4311(16)34525-2 https://doi.org/10.1016/j.applthermaleng.2017.10.132 ATE 11330
To appear in:
Applied Thermal Engineering
Received Date: Revised Date: Accepted Date:
31 December 2016 9 September 2017 22 October 2017
Please cite this article as: B. Eisavi, S. Khalilarya, A. Chitsaz, M.A. Rosen, Thermodynamic analysis of a novel combined cooling, heating and power system driven by solar energy, Applied Thermal Engineering (2017), doi: https://doi.org/10.1016/j.applthermaleng.2017.10.132
This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
Thermodynamic analysis of a novel combined cooling, heating and power system driven by solar energy Beneta Eisavi a, Shahram Khalilarya a, Ata Chitsazb,1, Marc A.Rosenc
a
Mechanical Engineering Department, Faculty of Engineering, Urmia University, Urmia, Iran b
c
Faculty of Mechanical Engineering, Urmia University, Urmia, Iran
Faculty of Engineering and Applied Science, University of Ontario Institute of Technology, 2000 Simcoe Street North, Oshawa, Ontario, L1H 7K4, Canada
Abstract A novel combined cooling, heating and power system (CCHP) driven by solar energy is proposed. The system applies an organic Rankine cycle, a double effect lithium bromide-water absorption refrigeration system, and heat exchangers to generate electricity, cooling and heating, respectively. Energy and exergy analyses are conducted to determine system efficiencies and losses. The cycle is compared with similar combined cycle that differs by using a cooling subsystem having a single effect absorption chiller. The effects of some key thermodynamic parameters on the performance of the cycle are examined. The results reveal that, for the same amount of input heat, the usage of a double effect absorption refrigeration system instead of a single effect absorption chiller can raise the amount of cooling power up to 48.5% and consequently can improve the performance of the system. At the same time, heating power rises by 20.5%, 1
Corresponding author: Tel: +989143636112
E–mail address: [email protected]
1
resulting in an increase in the cogeneration heat and power efficiency to 96.0%, while net electrical power production declines by 27%. In addition, it is determined that a high exergy destruction rate occurs in the solar collectors.
Keywords: Combined cooling, heating and power (CCHP); Solar energy; Organic Rankine Cycle; Double effect lithium bromide-water absorption refrigeration; Exergy 1. Introduction In recent years, increasing energy demands and the dependence of industry on fossil fuels and, in many instances, low energy efficiency systems contribute not only to the depletion of nonrenewable energy resources, but to various environmental problems, such as global warming and air pollution. Combined cooling, heating and power (CCHP), which simultaneously produces diverse types of energy (electricity, heating and cooling) from the same primary energy resource, can provide an alternative energy option which has high efficiency and helps address these challenges. A CCHP unit consists of a prime mover (e.g., steam turbine, microturbine, fuel cell, organic Rankine cycle), a heat recovery system (typically a heat exchanger that recovers waste heat), and thermally activated equipment (such as an absorption or adsorption chiller) [1]. Much research has been carried out on combined cycles. Xu et al. [2] proposed a combined cooling and power system, comprised of a Rankine cycle and an absorption chiller. Minciuc et al. [3] presented a method for analyzing CCHP systems. The performance of a cogeneration system based on a gas turbine has been examined [4]. Ho et al. [5] examined the total efficiency of a CCHP system under a diverse range of operating conditions. Hwang et al. [6] proposed a CCHP cycle based on a micro-turbine using an absorption chiller as a cooling
2
system, and explored the cycle’s energy savings and economy feasibility. Rosen et al. [7] performed exergy and energy analyses for a CCHP system. Kanoglu and Dincer [8] analyzed the performance for various configurations of CCHP systems which were designed to meet building energy demands. Most of combined cycles in that study included steam turbines, gas turbines, diesel engines and geothermal energy systems. In addition, several arrangements of CCHP systems have been examined based on first and second laws [9-14]. A brief review of micro-trigeneration system is conducted by Sonar et al [15]. In that study it is discussed that how integrated energy systems offer key solution to environmental problems and inefficient energy conversion systems. Basic elements of micro CCHP system along with their pros and cons were discussed. In addition application of integrated energy systems in energy sectors, alternative fuels in micro-trigeneration were investigated. The literature devoted to the study CCHP systems driven by solar energy are relatively rare. Buck and Friedmann [16] studied the performance of a CCHP based on a micro-turbine, with a cooling system comprised of a single effect or double effect absorption chiller. Better performance was observed for the CCHP system with a double effect absorption chiller. Medrano et al. [17] proposed a solar CCHP system using natural gas as an auxiliary fuel and analyzed the cycle based on the first and second laws of thermodynamics. Fumo et al. [18] performed energy and economic analyses of a solar CCHP system. Wang et al. [19] proposed a solar CCHP system and explored the effects of some key parameters on its performance. The authors also performed an optimization to achieve the maximum exergy efficiency. AlSulaiman et al. [20] presented a CCHP system driven by solar energy, and performed energy and exergy analyses to assess the effect of key parameters on the performance of the cycle. Thilak Raj et al. [21] carried out a review on the cogeneration technologies based on renewable energy sources, including biomass, solar energy, fuel cell and waste heat recovery
3
in which the various researches around them were classified and discussed based on study, design, modeling approach, environmental and economic issues. The most common working fluid combinations used in absorption refrigeration systems are lithium bromide-water (LiBr-H2O) and ammonia-water (NH3-H2O). The NH3-H2O system is more complicated than the LiBr-H2O system, since it needs a rectifying column that assures no water vapors enter the evaporator. Moreover, the refrigerant used in LiBr-H2O systems is water which is everywhere available, inexpensive and no toxic, In addition, the LiBr-H2O system can operate at a low generator temperature with better coefficient of performance than NH3-H2O system and are commercially available for air conditioning applications [22, 23]. Absorption chillers exist in single-effect and multiple effect designs although in multipleeffect designs, the double-effect one is more conventional. The double effect absorption refrigeration system was introduced in 1956-1958 [24]. Since then, much literature has been devoted to the study of the performance of these refrigerators for various working fluids [2531]. For instance, Gomri and Hakimi [31] applied energy and exergy analyses to study the performance of a double effect water-lithium absorption chiller. Also, Garousi et al. [32] conducted exergy and energy analyses for several arrangements including series, parallel and reverse parallel of absorption systems. One of the major design choices in double effect technology is the choice of how to connect the solution circuits. The basic options are parallel or series flow. Performance analysis showed that although COP of the series flow configuration is lower than the parallel one but its capacity is higher. In addition, in series flow the solution path is simpler. Marinating the correct flow split in parallel flow is problematic without active controls. Thus particularly for small systems, series flow is attractive as a design compromise. [33] To the best of the authors’ knowledge, investigations of CCHP systems based on solar energy using absorption chillers for refrigeration are scarce but needed to better understand such
4
systems and to increase the chances of their beneficial exploitation. Hence, in this paper, the following are considered:
A thermodynamic simulation and analysis is conducted of a CCHP system that uses solar energy as a heat source and applies an absorption chiller to produce cooling power.
The CCHP considered follows the system studied by Al-Sulaiman et al. [20], but with several differences, especially altering the cooling subsystem to accommodate a double effect absorption refrigeration system. In addition, another heat exchanger for producing thermal heat is considered.
The effect of this replacement is investigated thermodynamically. Furthermore, the proposed cycle is examined only for the solar mode,
The effects of several key thermodynamic parameters on the performance of the system are examined.
Exergy destruction rate of each device is evaluated.
2. System description A schematic diagram of the proposed CCHP system is given in Fig. 1. It consists of a solar collector field which provides the primary energy input of the combined cycle, an organic Rankine cycle that produces electricity, heat exchangers that generate steam from waste heat in heating process I and II, and series-flow double effect LiBr-H2O absorption chiller that provides cooling power. The operation of proposed system can be summarized as follows:
The heat transfer fluid in the solar collector domain which is Therminol-66 oil has a low relative pressure and its pressure is not sensitive to the increase in the temperature is selected according to [20] and some of its properties are brought in
5
Table 1 [34]. It passes through parabolic trough solar collectors. The absorbed heat is transferred to the ORC working fluid through the ORC evaporator. The flow exiting the ORC evaporator is used to produce steam in thermal heat exchanger II.
In the ORC, at the exit of the high pressure generator (HPG), n-octane as an organic fluid at the saturation state is pumped to the turbine inlet pressure. After passing through the ORC evaporator, the fluid becomes superheated then is expanded in the ORC turbine to produce mechanical energy which is converted to electricity by the generator. After, the flow passes through thermal heat exchanger I, where steam is generated. Then, the organic fluid enters the high pressure generator (HPG) in order to drive the absorption chiller.
In the series-flow double effect LiBr-H2O absorption chiller, the weak solution leaving the absorber is pumped to the HPG, passing through the low temperature heat exchanger (LTHE) and the high temperature heat exchanger (HTHE). Then the heat transferred to the HPG is utilized to boil out refrigerant vapor (primary refrigerant vapor) from the weak solution entering the HPG. In the LPG, the primary refrigerant vapor is condensed due to the low temperature of strong solution entering the LPG after passing through HTHE and EV-4. The refrigerant latent heat is consumed in producing the secondary refrigerant vapor from the strong solution, making the solution stronger. The condensed primary refrigerant vapor -after passing through EV-2- and the secondary refrigerant vapor both flow into the condenser to reject heat. At the exit of the condenser, the total amount of liquid refrigerant is delivered to the evaporator via EV-1 to generate the cooling effect. In the absorber, the refrigerant vapor leaving the evaporator is mixed with the strong solution entering the absorber, which comes from LPG after passing through LTHE and EV-3, thereby rejecting its heat of absorption.
6
3. Mathematical modeling In order to simplify the mathematical modeling of the proposed system based on the conservation principles of mass and energy, some assumptions are invoked. The system is taken to operate at steady state conditions, and pressure drops due to friction and heat losses across all the components of the CCHP are treated as negligible according to ref [20]. The ORC turbine and pumps have fixed isentropic efficiencies. It is worth nothing that as the main purpose is to investigate the effect of changing cooling subsystem with more efficient one the dynamic modeling was not conducted although it can be regarded in another study. Also, the series-flow double effect absorption refrigeration system, which uses LiBr-H2O as a working fluid, functions based on the following assumptions:
The refrigerant is assumed saturated exiting the condenser and the evaporator [31].
The weak solution leaving the absorber and the strong solution leaving the HPG and LPG are assumed to be saturated and at equilibrium conditions at their respective temperatures. [31]
The reference environment state for the system is taken to be
and
.
3.1. General analysis At steady state conditions, mass rate balances for the total mass and for the mixture of LiBr-H2O for the system components can be written as: (1) (2) where x is the concentration of lithium bromide in solution. Neglecting kinetic and potential energy, the first law of thermodynamics, which embodies the the law of conservation of energy, can be expressed on a rate basis as follows:
7
(3) Here,
denotes the heat transfer rate and
the work transfer rate.
An energy balance provides no information on the degradation of energy or resources during a process and does not quantify the usefulness or quality of the various energy and material streams flowing through a system and exiting as products and wastes. Therefore, the exergy method of analysis which overcomes the limitations of the first law of thermodynamics is considered here in addition to energy analysis. Exergy is maximum useful work that can be obtained from the amount of energy or the flow of materials in a specified reference environment [35]. An exergy rate balance for a steady state system can be written as: (4)
Neglecting kinetic and potential exergy, the physical exergy can be expressed as follows: (5) According to the results achieved by Avanessian et al. [36], the error caused by neglecting the chemical exergy of the solution is very low and does not affect results; therefore, chemical exergy is neglected for all flows and components here. The simulation and analysis of the combined cycle involves writing appropriate mass, energy and exergy rate balances based on the stated assumptions for each of the components. This information is presented in Table 2. 3.2. Overall analysis The performance of the combined cycle is analyzed and the energy and exergy efficiencies are determined. The energy efficiency is a measure expressing the rate of useful output energy to the rate of input energy of the system, and is defined as:
8
(6) where
denotes the electrical efficiency,
the net electrical power output and
the
heat input rate to the combined cycle provided by solar energy. Also, (7)
The combined heating and power efficiency can be expressed as: (8)
where
, the total rate of heating produced by the CCHP, is calculated by: (9)
The combined cooling and power efficiency can be written as: (10)
where
, the cooling effect generated by the absorption system, can be expressed as
follows: (11) The electrical to heating power ratio can be defined as: (12) and he electrical to cooling power ratio as: (13) The electrical exergy efficiency can be expressed as: (14) the combined heating and power exergy efficiency as:
9
(15)
and the combined cooling and power exergy efficiency as: (16)
Here
denotes the exergy rate input to the solar collector [20].
The equations used to model the solar field can be found in Refs. [20, 37]. 4. Validation In order to simulate the mathematical modeling, a computer program has been developed in Engineering Equation Solver (EES). The thermodynamic properties of LiBr-H2O which used are based on the new set of formulations presented by Patek and Klomfar [38], which are as internal functions in EES. In addition, n-octane thermodynamic properties were obtained using Functions defined in EES based on Fundamental Equation of State, as described by[39]. Since there was lack in experimental data in literature to use in validation of present model of the absorption chiller, the alternative approach was adopted, in which the obtained results of the present model are compared with those reported by Gomri [31]. There is a good agreement between the results from the two works, as seen in Table 3. The proposed system which is presented in Fig. 1 has been modeled and simulated based on the input data given in Table 4. 5. Results and discussion The performance of the CCHP system depends on the performance of each unit as well as the way the units are combined. Thermodynamic properties of the proposed cycle for the base case operating conditions are reported in Table 5.
10
The exergy destruction rates of various components of the proposed cycle which is presented in Fig. 2 provide important information about the sources of irreversibilities. Note that in this figure only components with higher exergy destruction rates are shown. The highest exergy destruction rate is attributable to the solar collectors which is affected by heat transfer irreversibilities between the sun and the collector, between the collector and the ambient air and inside the collector [40] and is found to be 1280 kW. The second highest exergy destruction rate occurs in the ORC evaporator due to high temperature difference and is determined to be 650 kW. In Table 6, the results of the thermodynamic modeling and simulation for present work are compared with the corresponding results reported by [20]. Relative to a single effect absorption chiller, the double effect absorption chiller operates with a higher temperature heat source. Therefore, the inlet temperature of the ORC pump needs to be increased. This measure leads to a decline of 27% in the net electrical power production, due to decrease in differences between lower and upper temperature of ORC cycle, consequently, electrical energy and exergy efficiency decrease, while cooling production under the same input power is increased by 48.5% because of the fact that an absorption chiller with higher coefficient of performance is used and the heating power rises by 20.5% .In fact, increasing ORC pump inlet temperature, decreases input energy to the ORC cycle although it brings about decrease in heating process I power, it increases the amount of available energy to the heating process II. As a result, the cogeneration heat and power production also increases. On the other hand, because of high exergy destruction rate occurred in heat exchanger of heating process II combined heating and power exergy efficiency declines. It is also observed that combined cooling and power efficiency is declined in comparison to base case which is due to the fact that the amount of cooling power is smaller than power production in size. Therefore, its increase could not cover the decline in power production decreases.
11
In the following subsections, the effects of several key parameters, such ORC evaporator pinch point temperature, ORC pump inlet temperature and turbine inlet pressure, are studied. 5.1. Effect of ORC evaporator pinch point temperature Although an evaporator with a lower pinch point temperature (Tpp) typically has higher performance, it incurs increased cost and design requirements. So, studying the effects of this change is an important factor, as demonstrated in Figs. 3-5. In Fig. 3 the effect of ORC evaporator pinch point temperature on the system outputs electrical to heating power ratio (Rh) and the electrical to cooling power ratio (Rc) are presented. It is observed that, with increasing Tpp, both net electrical power and thermal process I power output are reduced while thermal process II power output is increased. In fact, with increasing ORC evaporator pinch point temperature, the amount of input heat to the ORC evaporator is reduced and consequently the inlet temperature and inlet enthalpy of the ORC turbine are decreased. Referring to Fig. 3, heating process I power output decreases and heating process II power output increases from 4086 kW to 3313 kW and 898.4k W to 1745 kW, respectively. In addition, the net electrical power output decreases from 556.4 kW to 476.7 kW. Since in this study the fluid properties at points 5 and 1 control the cooling capacity, the only possible change for point 5 is pressure. Also, the temperature, pressure and mass flow rate of points 5 and 1 are identical. Therefore, the amount of cooling power output remains constant as the ORC evaporator pinch point temperature is varied. It is seen from Fig. 3 that both Rh and Rc decrease with increasing ORC evaporator pinch point temperature, and this is primarily due to the reduction of net electrical power. It is also observed that the electrical to heating power ratio decreases from 0.11 to 0.09 and similarly the electrical to cooling power ratio decreases from 2.36 to 2.03. Note that, as the effect of net
12
heating process power output is more dominant than the effect of net electrical power, Rh changes only slightly. Fig. 4 shows the changes of electrical, combined cooling and power, and combined heating and power efficiencies as ORC evaporator pinch point temperature varies. According to this figure, the electrical efficiency declines from 9.6% to 8.3% as Tpp changes from 5 30
to
. The reason of this drop is the reduction in the rate of heat input to the ORC evaporator,
which leads to a decrease in the turbine inlet temperature and consequently a decrease in turbine power output. Also, as stated before, when the ORC evaporator pinch point temperature increases, the available heat rate at the turbine exit declines, leading to a decrease in the heating process I power output; however, the heating process II power output increases, due to the decrease of ORC evaporator inlet energy. Therefore, the combined heating and power efficiency slightly increases from 96.0% at 5
to 96.1% at 30
as ORC evaporator
pinch point temperature increases. In contrast, the efficiency of the combined cooling and power system is reduced from 13.7% to 12.3%, as Tpp changes from 5
to 30
, mainly
because of the decrease in the net electrical power production. Fig.4 also shows the effect of Tpp on exergy efficiencies. With increasing T pp, the electrical, combined cooling and power and combined heating and power exergy efficiencies of the proposed system are seen to decline. The electrical exergy efficiency and the combined cooling and power exergy efficiency are observed to decrease from 4.8% to 4.1% and from 4.9% to 4.2%, respectively. In addition, the exergy efficiency of the combined heating and power system due to the reduction of input heat to the ORC evaporator decreases from 13.4% to 12.6%. Fig. 5 shows the variation of exergy destruction rate of the cycle components as ORC evaporator pinch point temperature varies from 5
to 30
. It is observed that the highest
rate of exergy destruction rate is related to the solar collectors and is found to be 1280 kW.
13
Due to the fixed operating temperature of the solar collectors, their exergy destruction rate is constant. Referring to Fig. 5, the exergy destruction rate of the ORC evaporator decreases from 694 kW to 625 kW as ORC evaporator pinch point temperature changes from 5
to 30
, due to the reduction of the exergy difference between the entering and exiting flows of the ORC evaporator. In addition, the exergy destruction rate of the heat exchangers for thermal processes I and II decrease from 390 kW to 264 kW and increase from 377 to 746 kW, respectively. As Tpp increases, the exergy destruction rate of the heat exchanger of heating process II changes significantly. Therefore, more care should be exercised in the selection and design of this component. It is also observed from Fig. 5 that the exergy destruction rate of the turbine decreases from 89 kW to 83 kW, with increasing Tpp from 5
to 30
which results
in lower available exergy. 5.2.Effect of ORC pump inlet temperature The effect of varying ORC pump inlet temperature (T1) on the performance of the cycle is described in Figs. 6-8. The temperature of ORC pump inlet temperature is selected to vary from 114
to 145
since in higher temperature of T1, crystallization phenomenon which is
probable in the weak solution entering the absorber is going to happen. [41] Fig. 6 shows the effect of ORC pump inlet temperature variation on system outputs, Rh and Rc. A variation in T1 has a direct impact on the amount of electrical power production due to the fact that with increasing T1, the difference between the minimum and maximum temperatures of the ORC decreases, which causes the electrical power production to decrease from 566 kW at 114
to 387 kW at 145
. Also, for the same reason, the thermal process I
power output is reduced from 3423 kW to 3298 kW. Referring to Fig. 6, an increase of ORC pump inlet temperature is seen to result in an ORC input heat rate decrease, which causes the thermal process II power output to increase from 1231 kW to 1883 kW. In addition, the ORC
14
pump inlet temperature rise causes a reduction in the rate of heat transfer to the HPG. It is also observed from Fig.6 that Rc decreases from 2.4 to 1.6. This drop is primarily due to the significant reduction in electrical power production as the ORC pump inlet temperature increases from 114
to 145
. In addition, as the ORC pump inlet temperature rises, as
mentioned earlier, a significant reduction is observed in the electrical power output and an increase in the thermal process II power output, which causes Rh to decrease from 0.12 at 114 to 0.07 at 145
.
Fig. 7 depicts the effect of ORC pump inlet temperature on the electrical, combined cooling and power and combined heating and power efficiencies. It is observed from Fig. 7 that the electrical efficiency decreases from 9.8% to 6.7%, mainly because of net electrical power output reduction as the ORC pump inlet temperature increases from 114
to 145
.
Similarly, the net electrical power reduction along with thermal process II power output rise cause the combined cooling and power efficiency to decrease from 14% at 114 145
to 10.8% at
. But, the combined heating and power efficiency increases slightly, from 90.5% to
96.5%. The reasons for this phenomenon are the decrease and increase of net electrical power output and thermal process II power output. Fig. 7 also shows the effect of varying ORC pump inlet temperature on the electrical, combined heating and power and combined cooling and power exergy efficiencies. An increase of the ORC pump inlet temperature is observed to lead to an exergy efficiency reduction owing to the decrement of net electrical power. Based on reasons mentioned earlier, as the ORC pump inlet temperature increases from 114
to
145 , the electrical exergy efficiency decreases from 4.9 % to 3.4 %, while the combined cooling and power and combined heating and power exergy efficiencies decrease from 5% to 3.4% and from 12.6% to 12%, respectively.
15
The impact of varying pump inlet temperature on the exergy destruction rate of components is presented in Fig. 8. The exergy destruction rate of the solar collectors is observed to be the highest of all components. With increasing T1 from 114
to 145
the
exergy destruction rate of the evaporator decreases, from 691 kW to 567 kW, owing to the lower available exergy in the ORC evaporator. Also, because of the increase in available energy of thermal process I, the exergy destruction rate of the heat exchanger assigned to thermal process I increases from 268 kW to 408 kW. In addition, the exergy destruction rate of the heat exchanger for thermal process II rises approximately by 286 kW (from 517 kW to 803 kW), due to increase of exergy difference. 5.4.Effect of turbine inlet pressure The analysis results demonstrate that the effect of varying turbine inlet pressure on system performance is insignificant. Therefore, it is not considered further here. Parametric analysis showed how system performance is affected by different key parameters. Critical components with higher exergy destruction rate are recognized and it is investigated that for having higher exergy efficiency for critical components including ORC evaporator, heating process II heat exchanger, heating process I heat exchanger and turbine; the cycle should work under specific conditions which lead to obtain higher exergy efficiency for each component and are presented as follows:
Parametric analysis showed that when cycle operates at lower pinch point temperature of ORC evaporator higher exergy efficiency is obtained for ORC evaporator and turbine while exergy efficiency for heat exchangers of heating process I and II is maximized at higher Tpp.
Moreover, it is deduced that when cycle operates in higher ORC pump inlet temperature, ORC evaporate, heat exchanger of heating process II and turbine operate
16
with higher exergy efficiency whereas it happens for heat exchanger of heating process I when ORC pump inlet temperature is minimum. 6. Conclusions A novel combined cooling heating and power system which is driven by solar energy has been successfully assessed. The main foci of the study were determining the impact of changing the refrigeration subsystem to one with a higher coefficient of performance and analyzing based on first and second laws of thermodynamics the effect on the cycle performance of important working parameters, such as ORC evaporator pinch point temperature, ORC pump inlet temperature and turbine inlet pressure. The main results and the conclusions that can be drawn from them are as follows:
Relative to a single effect absorption chiller, a series-flow double effect absorption chiller increases the cooling power by 48.5% for the same input heat.
For the base case analysis, the electrical efficiency is 8.9%, the combined heating and power efficiency is 96.0% and the combined cooling and power efficiency is 12.9%. Correspondingly, the electrical, the combined heating and power and combined cooling and power exergy efficiencies respectively are determined to be 4.4%, 12.8% and 4.5%. The significant difference between the energy and exergy efficiencies is due to the high exergy destruction rate destroyed by solar collectors.
The main sources of exergy destruction rate are the solar collector and the ORC evaporator, implying that these two components should be carefully selected and designed.
The electrical to the cooling power ratio is sensitive to the changes in the ORC pump inlet temperature. Controlling this parameter can be useful factor improving the electrical power amount.
17
The turbine inlet pressure has little impact on the performance of the CCHP. Therefore, the ORC can operate at lower pressures, reducing capital costs of the cycle.
Raising the ORC evaporator pinch point temperature has negative effect on the turbine and thermal processing I outputs, while it has the opposite effect on the thermal process II output. In general, increasing the ORC evaporator pinch point temperature reduces the energy and exergy efficiencies and, due to the reduction of thermal and electrical power outputs, the electrical to heating power and electrical to cooling power ratios.
An ORC pump inlet temperature increase has the greatest impact on the performance of the system, reducing the electrical and the heating power of thermal process I while increasing the output of thermal process II power. Also, raising the ORC pump inlet temperature, due to increase of ORC pump inlet temperature, reduces the energy and exergy efficiencies.
Nomenclature CCHP
combined cooling heating and power specific exergy (kJ/kg) specific enthalpy (kJ/kg)
HPG
high pressure generator
HTHE
high temperature heat exchanger
LPG
low pressure generator
LTHE
low temperature heat exchanger mass (kg)
ORC
organic Rankine cycle
18
heat (kJ) heat rate (kW) electrical to cooling power ratio electrical to heating power ratio temperature (
)
work rate (kW) x
lithium bromide concentration (%)
Greek symbols efficiency
Subscripts 0
atmospheric condition combined cooling and power combined heating and power collector exit electrical evaporator exergy generator heating inlet inlet
19
water
Superscripts .
rate physical
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[7] Rosen, M. A., Le, M. N., & Dincer, I. (2005). Efficiency analysis of a cogeneration and district energy system. Applied Thermal Engineering, 25(1), 147-159. [8] Kanoglu, M., & Dincer, I. (2009). Performance assessment of cogeneration plants. Energy Conversion and Management, 50(1), 76-81. [9] Arteconi, A., Brandoni, C., & Polonara, F. (2009). Distributed generation and trigeneration: Energy saving opportunities in Italian supermarket sector. Applied Thermal Engineering, 29(8), 1735-1743. [10] Bruno, J. C., Ortega-López, V., & Coronas, A. (2009). Integration of absorption cooling systems into micro gas turbine trigeneration systems using biogas: case study of a sewage treatment plant. Applied Energy, 86(6), 837-847. [11] Khaliq, A. (2009). Exergy analysis of gas turbine trigeneration system for combined production of power heat and refrigeration. International Journal of Refrigeration, 32(3), 534-545. [12] Ameri, M., Behbahaninia, A., & Tanha, A. A. (2010). Thermodynamic analysis of a trigeneration system based on micro-gas turbine with a steam ejector refrigeration system. Energy, 35(5), 2203-2209. [13] Huicochea, A., Rivera, W., Gutiérrez-Urueta, G., Bruno, J. C., & Coronas, A. (2011). Thermodynamic analysis of a trigeneration system consisting of a micro gas turbine and a double effect absorption chiller. Applied Thermal Engineering, 31(16), 3347-3353. [14] Ebrahimi, M., Keshavarz, A., & Jamali, A. (2012). Energy and exergy analyses of a micro-steam CCHP cycle for a residential building. Energy and Buildings, 45, 202-210.
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[15] Sonar, D., Soni, S. L., & Sharma, D. (2014). Micro-trigeneration for energy sustainability: Technologies, tools and trends. Applied Thermal Engineering, 71(2), 790796. [16] Buck, R., & Friedmann, S. (2007). Solar-assisted small solar tower trigeneration systems. Journal of Solar Energy Engineering, 129(4), 349-354. [17] Medrano, M., Castell, A., Fontanals, G., Castellón, C., & Cabeza, L. F. (2008). Economics and climate change emissions analysis of a bioclimatic institutional building with trigeneration and solar support. Applied Thermal Engineering, 28(17), 2227-2235. [18] Fumo, N., Chamra, L. M., & Bortone, V. (2009, January). Potential of solar thermal energy for CCHP systems. In ASME 2009 3rd International Conference on Energy Sustainability collocated with the Heat Transfer and InterPACK09 Conferences (pp. 5763). American Society of Mechanical Engineers. [19] Wang, J., Dai, Y., Gao, L., & Ma, S. (2009). A new combined cooling, heating and power system driven by solar energy. Renewable Energy, 34(12), 2780-2788. [20] Al-Sulaiman, F. A., Dincer, I., & Hamdullahpur, F. (2011). Exergy modeling of a new solar driven trigeneration system. Solar Energy, 85(9), 2228-2243. [21] Raj, N. T., Iniyan, S., & Goic, R. (2011). A review of renewable energy based cogeneration technologies. Renewable and Sustainable Energy Reviews, 15(8), 36403648. [22] Seraj, M., & Siddiqui, M. A. (2013). Performance analysis of parallel flow single and double effect absorption cycles. International Journal of Innovative Research in Science, Engineering and Technology, 2, 1570-1576.
22
[23] Kalogirou, S. (2008). Recent patents in absorption cooling systems. Recent Patents on Mechanical Engineering, 1(1), 58-64. [24] Lawson, M. B., Lithgow, R. A., & Vliet, G. C. (1982). Water-lithium bromide doubleeffect absorption cooling cycle analysis. ASHRAE Trans.;(United States), 88(CONF820112-). [25] Xu, G. P., & Dai, Y. Q. (1997). Theoretical analysis and optimization of a double-effect parallel-flow-type absorption chiller. Applied Thermal Engineering, 17(2), 157-170. [26] Arun, M. B., Maiya, M. P., & Murthy, S. S. (2000). Equilibrium low pressure generator temperatures for double-effect series flow absorption refrigeration systems. Applied Thermal Engineering, 20(3), 227-242. [27] Arun, M. B., Maiya, M. P., & Murthy, S. S. (2001). Performance comparison of doubleeffect parallel-flow and series flow water–lithium bromide absorption systems. Applied Thermal Engineering, 21(12), 1273-1279. [28] Liu, Y. L., & Wang, R. Z. (2004). Performance prediction of a solar/gas driving double effect LiBr–H2O absorption system. Renewable Energy, 29(10), 1677-1695. [29] Torrella, E., Sánchez, D., Cabello, R., Larumbe, J. A., & Llopis, R. (2009). On-site realtime evaluation of an air-conditioning direct-fired double-effect absorption chiller. Applied Energy, 86(6), 968-975. [30] Ravikumar, T. S., Suganthi, L., & Samuel, A. A. (1998). Exergy analysis of solar assisted double effect absorption refrigeration system. Renewable Energy, 14(1), 55-59. [31] Gomri, R., & Hakimi, R. (2008). Second law analysis of double effect vapour absorption cooler system. Energy Conversion and Management, 49(11), 3343-3348.
23
[32] Garousi Farshi, L., Seyed Mahmoudi, S. M., Rosen, M. A., & Yari, M. (2012). A comparative study of the performance characteristics of double‐effect absorption refrigeration systems. International Journal of Energy Research, 36(2), 182-192. [33] Herold, K. E., Radermacher, R., & Klein, S. A. (2016). Absorption chillers and heat pumps. CRC press. [34] Therminol.therminol-66 heat transfer fluid, www.therminol.com;2010. [35] Dincer I, Rosen MA. (2013) Exergy: Energy, environment and sustainable development. 2nd edition Elsevier. [36] Avanessian, T., & Ameri, M. (2014). Energy, exergy, and economic analysis of single and double effect LiBr–H2O absorption chillers. Energy and Buildings, 73, 26-36. [37] Chen, X., Gong, G., Wan, Z., Zhang, C., & Tu, Z. (2016). Performance study of a dual power source residential CCHP system based on PEMFC and PTSC. Energy Conversion and Management, 119, 163-176. [38] Patek, J., & Klomfar, J. (2006). A computationally effective formulation of the thermodynamic properties of LiBr–H2O solutions from 273 to 500 K over full composition range. International Journal of Refrigeration, 29(4), 566-578. [39] Span, R., & Wagner, W. (2003). Equations of state for technical applications. II. Results for nonpolar fluids. International Journal of Thermophysics, 24(1), 41-109. [40] Tyagi, S. K., Wang, S., Singhal, M. K., Kaushik, S. C., & Park, S. R. (2007). Exergy analysis and parametric study of concentrating type solar collectors. International Journal of Thermal Sciences, 46(12), 1304-1310.
24
[41] Farshi, L. G., Mahmoudi, S. S., & Rosen, M. A. (2011). Analysis of crystallization risk in double effect absorption refrigeration systems. Applied Thermal Engineering, 31(10), 1712-1717. [42] Xu, D., Liu, Q., Lei, J., & Jin, H. (2015). Performance of a combined cooling heating and power system with mid-and-low temperature solar thermal energy and methanol decomposition integration. Energy Conversion and Management, 102, 17-25. [43] Karellas, S., & Braimakis, K. (2016). Energy–exergy analysis and economic investigation of a cogeneration and trigeneration ORC–VCC hybrid system utilizing biomass fuel and solar power. Energy Conversion and Management, 107, 103-113.
25
Table 1 Typical properties of Therminol-66 Source: [34] Substance name Composition Density at 25 C Kinematic viscosity at 40 C Normal boiling point Optimum use range Pseudocritical temperature Pseudocritical pressure
Therminol-66 Modified terphenyl 1005 kg/m3 29.6 mm2/s 359 C 0-345 C 569 C 24.3 bar
26
Table 2 Relations of thermodynamic rate balances for system components. Component
Energy rate balance
ORC turbine Heating process I
ORC pump ORC evaporator
Heating process II
Solar pump
HPG
LPG
Condenser Evaporator Absorber LTHE
HTHE Solution pump EV-1 EV-2 EV-3
27
Exergy rate balance
Table 3 Comparison of present modeling and simulation results of absorption chiller with those of Gomri [31]. State point Gomri Present
Gomri
work
Present
Gomri
work
Present work
Present work
1
35
35
0.12702
0.127
0
0
146.6
2
4
4
0.12702
0.127
0
0
146.6
3
4
4
0.12702
0.127
0
0
2508
4
35
35
1.73674
1.737
55.869
55.87
87.59
5
35
35
1.73674
1.737
55.869
55.87
87.62
6
62.49
62.66
1.73674
1.737
55.869
55.87
143.4
7
106.81
107.2
1.73674
1.737
55.869
55.87
236.4
8
130
130
1.67132
1.672
58.056
58.06
288.6
9
76.09
82.86
1.67132
1.672
58.056
58.06
192
10
76.09
82.86
1.67132
1.672
58.056
58.06
192
11
130
130
0.06542
0.06544
0
0
2741
12
82.46
82.46
0.06542
0.06544
0
0
345.2
13
35
35
0.06542
0.06544
0
0
345.2
14
80
80
0.06160
0.06161
0
0
2650
15
80
80
1.60972
1.61
60.278
60.28
195.8
16
48.67
48.5
1.60972
1.61
60.278
60.28
135.6
17
48.67
48.5
1.60972
1.61
60.278
60.28
135.6
28
Table 4 Conditions of simulation for the proposed CCHP. Term
Value
ORC turbine efficiency
80 % a
ORC pump efficiency
80 % a 2000 kPaa
Baseline turbine inlet pressure ORC pump inlet temperature
120
Electrical generator efficiency
95 % a
Electrical motor efficiency
95 % a
Solution pump efficiency
95 % b
Pinch point temperature of ORC evaporator
20
Effectiveness of solution heat exchangers
70 % b 120 o C b
HPG temperature Condensation temperature
35 o C b
Absorber temperature
35 o C b
Evaporation temperature
7 oC
Minimum heat transfer temperature in LPG
5 oC b
b
500 W/m2a
Solar radiation a [20] b [31]
29
Table 5 Thermodynamic properties of cycle states. State point 1
120
86.81
-
229.5
0.6659
217.4
7
2
121
2000
-
233.4
0.6679
240.3
7
3
283.6
2000
-
856
1.941
1942
7
4
223.4
86.81
-
774.8
1.982
1287
7
5
120
86.81
-
257.9
0.7381
265.5
7
6
120
44.72
0
2722
7.849
19.01
0.04918
7
78.58
44.72
0
329
1.058
0.8851
0.04918
8
35
5.627
0
329
1.097
0.3219
0.04918
9
73.58
5.627
0
2637
8.574
4.319
0.05027
10
35
5.627
0
146.6
0.505
0.05938
0.09945
11
7
1.002
0
146.6
0.5246
-0.5221
0.09945
12
7
1.002
0
2513
8.973
-15.6
0.09945
13
120
44.72
0.5563
263.2
0.7133
94.21
1.71
14
77.71
44.72
0.5563
173.8
0.4744
63.21
1.71
15
77.71
5.627
0.5563
173.8
0.4744
63.21
1.71
16
73.58
5.627
0.5731
170.6
0.4359
75
1.66
17
46.57
5.627
0.5731
116.7
0.2739
65.64
1.66
18
46.57
1.002
0.5731
116.7
0.2739
65.64
1.66
19
35
1.002
0.5408
81.15
0.2184
36.21
1.759
20
35
44.72
0.5408
81.18
0.2184
36.26
1.759
30
21
59.59
44.72
0.5408
132.1
0.3776
42.32
1.759
22
100.3
44.72
0.5408
218.9
0.6225
66.67
1.759
23
326.9
54.05
-
1602
1.606
5006
7.1
24
186.8
1.497
-
987.8
0.9701
1994
7.1
25
131.9
0.1997
-
788.8
0.7097
1132
7.1
26
131.9
0.1997
-
788.8
0.7097
1132
7.1
* * * *
*The data in the literature and EES can calculate only vapor pressure of therminol-66.
31
Table 6 Comparison of present results with those of [20]. Parameter COP
Present work
ref [20]
1.18
0.79
Net electrical power
511.4
Cooling power
235.4
158.5
Heating power
5031
4172.7
700
Electrical efficiency
8.9 %
12.5 %
Combined heating and power efficiency
96.0 %
88 %
Combined cooling and power efficiency
12.9 %
15.5 %
Electrical exergy efficiency
4.4
6.2 %
Combined heating and power exergy efficiency
12.8 %
17.6 %
Combined cooling and power exergy efficiency
4.5 %
6.4 %
32
Fig. 1. Schematic of CCHP system.
33
Fig. 2. Exergy destruction rate of components of the proposed cycle (P 3=2000 kPa, T1=120 , Tpp=20
34
).
Fig. 3. Effect of ORC evaporator pinch point temperature (T pp) on cycle outputs, Rc and Rh at P3=2000 kPa, T1=120
35
.
Fig. 4. Effect of ORC evaporator pinch point temperature (T pp) on efficiencies at P3=2000 kPa, T1=120
36
.
Fig. 5. Effect of ORC evaporator pinch point temperature (T pp) on exergy destruction rate of components at P3=2000 kPa, T1=120
37
.
Fig. 6. Effect of ORC pump inlet temperature (T1) on cycle outputs, Rh and Rc at P3=2000 kPa, Tpp=20
38
.
Fig. 7. Effect of ORC pump inlet temperature (T1) on efficiencies at P3=2000 kPa, Tpp=20
39
Fig. 8. Effect of ORC pump inlet temperature (T1) on exergy destruction rate of components at P3=2000 kPa, Tpp=20
40
.