Proposal and Investigation of a New Tower Solar Collector-Based Trigeneration Energy System

Abstract

These days, the low efficiency of solar-based thermal power plants results in uneconomical performance and high-cost uncompetitive industries compared with conventional fossil fuels. In order to overcome such issues, a novel combined cooling–power–heating (trigeneration) system is proposed and analyzed in this paper. This system uses an ammonia–water binary mixture as a working fluid and a solar heat source to produce diverse types of energy for a multi-unit building in a sustainable fashion. In addition to the basic cooling–power cogeneration cycle, a flashing chamber that will boost the flow rate of refrigerant without any additional heat supply is employed. By developing a mathematical model, the system performance is analyzed using varying parameters of solar irradiation, hot oil temperature, process heat pressure, and ambient temperature to investigate the influence on electrical power, cooling capacity, refrigeration exergy, energy utilization factor (EUF), and system exergy efficiency. Increasing direct normal irradiation (DNI) from 500 W/m² to 1000 W/m² reduces the system EUF and exergy efficiency from 53.62% to 43.12% and from 49.02% to 25.65%, respectively. Both power and refrigeration exergy increase with increasing DNI and ambient temperature, while heating exergy remains constant. It is demonstrated that of 100% solar energy supplied, 46.03% is converted into energetic output and 53.97% is recorded as energy loss. The solar exergy supplied is distributed into 8.34% produced exergy, 29.78% exergy loss, and the remaining 61.88% is the destructed exergy. The highest destruction of solar exergy (56.89%) occurs in the central receiver.

1. Introduction

Energy concerns have risen at the global level in the recent past because 70% of the world’s energy requirement is met through consumption of fossil fuels, of which 40% is used in the building sector to meet internal energy demands such as cooling, ventilation, heating, and air conditioning [1]. Problems associated with traditional energy resources like the depletion of fossil fuel reserves, global warming, and increased pollution levels motivated researchers to search for new alternatives to overcome such challenges [2]. In this situation, multi-generation systems driven by renewable sources of energy possess enormous and promising potential as these systems can simultaneously produce diverse types of energy from the same primary energy source in a sustainable manner. Among the renewables, solar energy has a position of pride as it provides a clean and promising option due to its abundant availability and emission-free nature [3,4]. Employment of concentrating solar power (CSP) technology is the only solution for achieving higher temperatures through efficient conversion of solar energy to thermal energy. A typical CSP plant consists of three main subsystems: a solar collector field, a solar receiver, and a power conversion system. There are four CSP families based on the two major solar subsystems, i.e., the collector and the receiver: parabolic trough, linear Fresnel, dish Stirling, and solar tower (also known as central receiver) [5]. Ouagued [6] developed a numerical model of a parabolic trough collector for the Algerian climate. In this model, the receiver is divided into several segments and heat transfer balance equations, which rely on the collector type, optical properties, heat transfer fluid (HTF), and ambient conditions, are applied to each segment. Archer [7] developed a mathematical model of a parabolic trough collector for solar cooling and heating using energy balance correlations between the absorber tube, glass tube, and surroundings. The results of the comparison between the mathematical model and the experimental data indicated some differences, including high glass temperature and low efficiencies. Hachicha [8] developed a numerical model of a parabolic trough receiver for thermal and optical analysis. All the heat energy balance correlations are used in the model. A mathematical–geometric method is applied to estimate heat flux around the receiver. Cagnoli et al. [9] presented the assessment of heat loss from the receiver of a real 1 MWe pilot plant based on the Fresnel collector cooled with thermal oil. The receiver unit, which consists of an absorber tube and a compound parabolic concentrator (CPC), is investigated numerically in order to determine the receiver performance in different wind directions. Sepulveda et al. [10] carried out several energy audits in industries located in the southwest of Europe (with considerable steam consumption), quantifying thermal and energy consumption and defining both work schedules and seasonality. Their analysis ranked tomato industries as the most suitable for LFC technology due to their operation during months with the highest solar isolation, integrating the solar schedule in a 24-h working day. Among the various concentrating solar power (CSP) technologies that can be applied to convert solar radiation into thermal energy, heliostat-based central receiver technology is considered to be more efficient and economically better as it has the largest and most flexible range of operating temperatures. Xu et al. [11] presented energy and exergy analyses of a thermal power plant coupled with the heliostat-based central receiver. They discussed the effect of different operating parameters on the plant’s energetic and exergetic performance. Rabbani et al. [12] attempted to investigate the performance of an integrated system driven by the heliostat-based central receiver CSP source. Their computational results show that tower solar collectors can contribute significantly to the sustainable production of electricity and useful heat due to their potential to maintain higher thermal efficiency while producing different types of useful energy. To achieve greater sustainability, it is important to utilize solar resource more efficiently. This can be achieved by minimizing thermodynamic losses using the plant’s waste heat for useful purposes (hot water production, space heating, refrigeration using absorption chillers, industrial drying, etc.). To this effect, CCHP systems driven by solar energy provide greater economic benefits compared with solar-independent power, heating, and cooling systems. In the application of combined systems, the same solar collectors, pipes, and auxiliary equipment are used to produce three diverse types of energy. CCHP systems are capable of meeting the energy requirements of a building because a building requires electricity to meet the internal energy demands, process heating is required for drying or space conditioning and cold is required for air conditioning. Utilization of stored solar energy to drive the integrated energy systems that generate cooling, powering, and heating are of prime importance because high demand for cooling for air conditioning and refrigeration occurs at peak summer time when high intensity solar radiation is largely available [13,14,15,16]. Several investigations have reported the development and exergetic analysis of CCHP systems. Such systems mainly consist of an organic Rankine cycle (ORC) combined with a heat-operated absorption refrigeration cycle (ARC) and a process heat exchanger; these integrations exhibit tremendous potential for efficient utilization of solar energy by displaying significantly higher energetic and exergetic efficiencies compared with traditional energy conversion systems. The advancement of existing solar-driven combined power, heating, and cooling systems is as follows: Boyaghchi and Heidarnejad [17] proposed a CCHP system of micro solar-type that combines the organic Rankine cycle with absorption refrigeration cycle to meet the domestic energy requirements of a building in summer and winter seasons and the cycle was thermodynamically investigated. Zhao et al. [18] analyzed various configurations affecting the operation and performance of a solar-driven ORC-based CCHP system. They obtained the optimal operating conditions for the system based on detailed thermodynamic investigations. Khaliq et al. [19] examined the influence of selected operating variables on thermodynamic performance of a trigeneration system consisting of an ORC integrated to absorption chiller for producing low temperature refrigeration using a solar energy source. Their results revealed that the type of ORC working fluid and expander inlet pressure significantly affects the CCHP system’s energetic and exergetic performances. Aghaziarati et al. [20] introduced a new type of solar-driven CCHP system comprising an ORC combined with a compression–absorption cooling system and a heat exchanger. The system was developed to meet all the energy needs of a hospital, including electricity, cooling, and heating. The energy- and exergy-based efficiencies of the system were determined to be 89.39% and 8.70%, respectively. Research conducted on the development and analysis of thermodynamic cycles designed for multi-objective generation systems reveals that the cycle’s productivity can be further increased with the utilization of binary mixtures. This is because binary mixtures accompanied by the varying temperature processes of evaporation and condensation at the steady state pressure raise the system performance because of temperature matching in the evaporator and condenser, which in turn reduces thermodynamic irreversibility [21,22,23]. Numerous investigations have been conducted in the field of combined cooling and power cycles operated on binary mixtures as the working fluid and herein thermal cooling cycle is combined with the Kalina power cycle, which shows excellent thermodynamic performance compared with organic Rankine cycle in regard to efficiencies. In this regard, many investigators have proposed novel integration of Kalina cycle with other energy conversion systems to obtain the versatility of not only mechanical power production but also heating and refrigeration of a thermal load while using the thermal efficiency characteristics of the Kalina cycle [24,25,26]. The energy and exergy performance of the integration of Kalina cycle with vapor absorption cycle for simultaneous generation of power and refrigeration has been studied to a certain level and is reported by the following investigators: Yu et al. [27] developed a combined cycle utilizing ammonia–water mixture for the purpose of producing refrigeration and electric power. The system developed provides a promising cooling to power ratio by superimposing the modified Kalina cycle with the absorption cooling cycle. Cao et al. [28] proposed a novel cooling–power cycle where the Kalina cycle is integrated with the absorption refrigeration cycle and the exergy indicating parameter of such a cogeneration cycle was maximum at an optimum value of the turbine entry pressure. Chen et al. [29] utilized the ammonia–water mixture for combined generation of power and cooling by utilizing exhaust heat from the engine. By employing the aqua–ammonia mixture as a working fluid at a high pressure of 100 bar, they generated additional power of 92.86 kW with an efficiency of 19.76%. Shokati et al. [30] analyzed various types of configurations that combined the Kalina cycle with an absorption chilling machine. Cogeneration of power and cooling makes it possible to effectively exploit the energy resource, to some extent, as it produces two different forms of energy from the single energy supply. A trigeneration system, which simultaneously generates cooling, heating, and power from a single energy resource, has emerged in the recent past with the aim of utilizing the given energy resource more effectively and meeting the diverse types of energy demands of industrial and residential sectors. Only a few studies have investigated Kalina cycle-based trigeneration systems using energy and exergy methods. Furthermore, Kalina cycle-based trigeneration systems powered by solar energy have been scarcely considered and analyzed. Almatrafi and Khaliq [31] presented an analysis based on the energy and exergy theory of a solar-driven trigeneration system comprising the Kalina cycle and absorption cooling cycle for the combined production of power, heating, and low temperature refrigeration. The simulation studies revealed that the energy efficiency of trigeneration increased from 15.90% to 21.27% and its exergy efficiency increased from 11.80% to 17.25% following the increase in solar fluid temperature from 433K to 453K. Moradpoor and Ebrahimi [32] developed a novel Kalina cycle-based trigeneration system whose condenser was used to supply heating and an absorption cooling cycle was employed to produce refrigeration power. The system was parametrically investigated from a thermo-environmental point of view. The results showed that the proposed trigeneration system saves up to 45% of the energy supply in comparison to a single generation system. Ebrahimi-Moghadam et al. [33] assessed the energy and exergy performances of a novel trigeneration district energy system. From their robust parametric investigation, the optimum values of the outputs of the trigeneration system were determined to be 1025.9 kW (power), 1642.3 kW (heating), and 304.9 kW (cooling).

Investigations reported in the literature on the development and assessment of combined energy systems comprising a Kalina cycle and an absorption chiller integrated with a solar energy source reveals a low performance of absorption refrigeration cycle because a significant amount of energy input is required to attain the separation process. Introduction of a flash chamber in the previously investigated cogeneration cycle assists in generating additional refrigerant without further supply of any heat via flashing that leads to a greater output of cooling and power [34]. Exergy analysis of such modified combined energy systems can provide further insights. Little research is available on the integration of the parabolic trough solar collector system with the Kalina cycle and absorption cooling cycle with a maximum of two concentration ratios (without flashing) and to date, no study has reported the development and analysis of the integration of tower solar collector, binary mixture-operated combined power and cooling cycle employing a flash chamber along with a separator. Therefore, the present research aims to present thermodynamic modeling and exergy analysis of a solar-based trigeneration system. The sub-objectives are:

• To model a new trigeneration system based on a tower solar collector, a process heat exchanger, and a cogeneration cycle employing the concept of separation and flashing for simultaneous generation of heating, power, and predominantly high cooling output.
• To perform exergy analysis of the system and parametric investigation of the effects of varying selected operating parameters on overall energy and exergy efficiencies.
• To quantify the distribution of irreversibilities in system components with the aim of identifying the weak spots, meaning the component accounts for the highest exergy loss whose reduction helps to improve system performance.

2. Materials and Methods

2.1. System Description

The solar-driven trigeneration system developed in this study is a combination of a tower solar collector, a process heat exchanger (PH), and a Kalina cycle integrated with a vapor-absorption cooling cycle employing the concept of flashing. Figure 1 depicts the layout of the system of trigeneration considered for investigation. The solar field considered consists of a small heliostat field (Helio) and a central receiver (CR), where the heliostats are two-axis tracking mirrors that reflect the sun onto a fixed spot on top of a tower during the course of the day. The concentrated rays that strike the receiver increase its temperature and this in turn heats up the oil (Duratherm 600)

Table 1. Properties of solar heat transfer fluid (Duratherm600 Oil) [19].

Property	Value
Temperature considered (°C)	200
Viscosity (Centistoke)	5.6 at 100 °C
Density of oil (g/mL)	0.693 at 260 °C
Thermal conductivity (W. K−1.m−1)	0.13 at 260 °C
Specific heat (Average) (kJ. K−1.kg−1)	2.33
Thermal expansion coefficient (/°C)	0.012

. The amount of solar energy used to heat the oil may be referred to as solar energy gain. The oil passes through the tubes embedded in the central receiver, gets heated, and then enters the process heat exchanger to generate heating for high temperature use. In the configuration where the Kalina cycle combines with the absorption refrigeration cycle, the absorber (Abs), solution pump (SP), and solution heat exchanger (SHX) are connected with the generator (Gen) in a series. At the outlet of the separator (Sep), a flashing chamber (FC) is employed in the line of solution considered to be weak and the other outlet goes to the superheater (SH) as depicted in Figure 1. The flashing chamber and superheater outlets are connected to the exit and entry of the expander (Exp), respectively. The condenser (Cond) is used to condense the ammonia vapor in the cycle, producing cooling and power together. Furthermore, the line is connected to the evaporator through the sub-cooler (SC). A solution heat exchanger is employed after the flashing chamber (FC) to preheat the strong solution concentration that is directed to the generator (Gen). The refrigerant is generated in two parts: (i) the process following the thermal separation process and (ii) a process of flashing. Three pressure levels are found; first (HP) generator pressure, second (IP) condensing pressure, and third (LP) sink pressure. The intermediate pressure is the same as the condensing pressure and is proposed for flashing the aqua–ammonia mixture in the cooling–power cycle (Kalina power cycle combined with the absorption cooling cycle). The ammonia-rich strong solution exiting the Abs is delivered to the high-pressure generator through SHX. The separator splits the ammonia-rich strong solution into ammonia vapor and a weak concentration solution. The high-pressure ammonia vapor enters the Exp through the superheater. Expansion of this high-pressure ammonia vapor in the expander produces the power output. The weak concentration solution leaving the generator passes through the flashing chamber, where additional ammonia vapor is generated. Pressure in the flashing chamber and condenser are equally maintained. To reduce the pressure before the flashing chamber, a Throttling valve 1 (TV1) is employed. When the weak aqua–ammonia solution acquires low pressure and high temperature, flashing occurs, generating additional ammonia vapor. Flashing generates ammonia vapor at the pressure attained by the ammonia vapor leaving the expander; therefore, both these streams combine before entering the condenser. After condensation, ammonia vapor mixture changes to liquid phase that is then subcooled and its pressure reduced by throttling (via TV3) to generate the desired refrigeration effect. The refrigerant flowing rate is the sum of expander outlet vapor and ammonia vapor leaving the flashing process, leading to generation of a greater refrigeration effect $\left({\dot{Q}}_{Eva}\right)$ at the evaporator $\left(Eva\right)$. The weak ammonia solution in the flashing chamber outlet and the ammonia vapor in the outlet of the sub-cooler are mixed and the mixture goes to the absorber to generate a solution rich in aqua–ammonia in order to complete the trigeneration cycle.

2.2. Thermodynamic Modeling of the Proposed Trigeneration System

In order to perform thermodynamic evaluation of the developed trigeneration system based on the laws of conservation of energy and mass, several conditions are assumed. The system is assumed to work under steady state and reduction in pressure due to friction in all the heat transfer components of trigeneration is not considered. The isentropic efficiencies of the pumps and expander are fixed (95%). Heat loss at the superheater (SH), heat exchanger generating the process heat (PH), generator (Gen), solution heat exchanger (SHX), condenser (Cond), absorber (abs), and evaporator (ev) was not considered during the investigation. Flow across the throttling valves (TV1, TV2, TV3) is assumed to be isenthalpic. For the fluid used for conversion of solar energy to thermal energy and the cogeneration cycle working fluid, which is NH₃-H₂O binary mixture, only physical exergy is considered in the computational analysis. The concentrations of ammonia vapor generated during flashing and separation are equal. The kinetic and potential exergy rates of the flow stream are not considered due to insignificant variations in the velocity and height of control volume. The reference pressure, P₀, and temperature, T₀, designated for the assumed environment were taken as 1.013 bar and 25 °C, respectively.

The solar heat source consists of the heliostat (Helio)-based central receiver (CR), where tubes are embedded for the conversion of solar radiation into heat, which is absorbed by solar heat transfer fluid and used to drive the proposed trigeneration system. The equations employed for modeling the tower solar collector applied in this system can be found in Refs. [14,31]. The enthalpy, entropy, and exergy of the aqua–ammonia (NH₃-H₂O) mixture at state points of the cooling–power cogeneration cycle are extremely important for thermodynamic evaluation of the proposed trigeneration system and were computed by applying REFPROP 9.1 [35]. To investigate the performance of the developed cooling–heating–power trigeneration system, energy and exergy analyses were considered and are mentioned here. The mass, energy, and exergy balance equations, as well as the components, are formulated for the entire system. For evaluation of algebraic equations, the Engineering Equation Solver (EES) software, which is considered one of the most robust tools for thermodynamic investigations, was employed [36].

2.2.1. Energetic and Exergetic Analyses

The newly developed trigeneration system considered in this study was investigated from exergy and energy perspectives. The law of conservation of energy derived from the first law of thermodynamics discusses the quantity of energy conversion and has been found incapable of determining the work potential or quality of an energy resource in relation to the reference environment [11]. Exergetic analysis not only reveals the amount of useful work potential, or exergy, associated with the energy resource, it also demonstrates the distribution of irreversibility of the system’s components and determines the components that contribute most to system inefficiencies [19,22].

Considering the steady state condition of the control volume, mass rate balances of the total mass and the mixture of NH₃-H₂O for components of the system as well as the energy balance relation can be expressed as [21,28]:

$${\displaystyle \sum }{\dot{m}}_{in}={\displaystyle \sum }{\dot{m}}_{out}$$

(1)

$${\displaystyle \sum }{\left(\dot{m}X\right)}_{in}={\displaystyle \sum }{\left(\dot{m}X\right)}_{out}$$

(2)

$${\displaystyle \sum }{\left(\dot{m}h\right)}_{in}-{\displaystyle \sum }{\left(\dot{m}h\right)}_{out}+{\displaystyle \sum }{\dot{Q}}_{in}-{\displaystyle \sum }{\dot{Q}}_{out}+\dot{W}=0$$

(3)

where $\dot{m}$ and X represent the mass flow rate and ammonia concentration, respectively, $\dot{Q}$ and $\dot{W}$ represent the heat transfer rate and the work transfer rate, respectively, and ‘h’ indicates the specific enthalpy.

Mass/energy balance expressions of the individual elements are developed through application of the aforementioned equations on the system components and are shown in

Table 2. Equations of Energy balance rates of the key elements of the system.

Component	Energy Balance Rate
Heliostat (Halio)	${\dot{Q}}_{solar}={\dot{Q}}_{CR}+{\dot{Q}}_{loss,Helio}$
Central Receiver (CR)	${\dot{m}}_5{h}_5+{\dot{Q}}_{CR}={\dot{m}}_1{h}_1+{\dot{Q}}_{loss,CR}$
Process Heater (PH)	${\dot{m}}_1{h}_1={\dot{m}}_2{h}_2+{\dot{Q}}_P ; \& {\dot{Q}}_P=\left({\dot{m}}_{27}{h}_{27}-{\dot{m}}_{26}{h}_{26}\right)$
Expander (Exp)	${\dot{m}}_6{h}_6={\dot{m}}_7{h}_7+{\dot{W}}_{Exp}$
Superheater (SH)	${\dot{m}}_2{h}_2+{\dot{m}}_{19}{h}_{19}={\dot{m}}_3{h}_3+{\dot{m}}_6{h}_6$
Evaporator (Eva))	${\dot{m}}_{11}{h}_{11}+{\dot{Q}}_{Eva}={\dot{m}}_{12}{h}_{12}$
Condenser (Cond)	${\dot{m}}_8{h}_8={\dot{m}}_2{h}_2+{\dot{Q}}_{Cond} ;$&  ${\dot{Q}}_{Cond}=\left({\dot{m}}_{29}{h}_{29}-{\dot{m}}_{28}{h}_{28}\right)$
Separator (Sep)	${\dot{m}}_{18}{h}_{18}={\dot{m}}_{19}{h}_{19}+{\dot{m}}_{20}{h}_{20}$
Subcooler (SC)	${\dot{m}}_9{h}_9+{\dot{m}}_{12}{h}_{12}={\dot{m}}_{10}{h}_{10}+{\dot{m}}_{13}{h}_{13}$
Generator (Gen)	${\dot{m}}_3{h}_3+{\dot{m}}_{17}{h}_{17}={\dot{m}}_4{h}_4+{\dot{m}}_{18}{h}_{18}$
Flash chamber (FC)	${\dot{m}}_{21}{h}_{21}={\dot{m}}_{22}{h}_{22}+{\dot{m}}_{23}{h}_{23}$
Solution heat exchanger (SHX)	${\dot{m}}_{22}{h}_{22}+{\dot{m}}_{16}{h}_{16}={\dot{m}}_{17}{h}_{17}+{\dot{m}}_{24}{h}_{24}$
Absorber (Abs)	${\dot{m}}_{13}{h}_{13}+{\dot{m}}_{25}{h}_{25}={\dot{m}}_{15}{h}_{15}+{\dot{Q}}_{Abs}$; &  ${\dot{Q}}_{Abs}=\left({\dot{m}}_{33}{h}_{33}-{\dot{m}}_{32}{h}_{32}\right)$
Throttle valve1 (TV1)	${\dot{m}}_{20}{h}_{20}={\dot{m}}_{21}{h}_{21}$
Throttle valve2 (TV2)	${\dot{m}}_{24}{h}_{24}={\dot{m}}_{25}{h}_{25}$
Throttle valve3 (TV3)	${\dot{m}}_{10}{h}_{10}={\dot{m}}_{11}{h}_{11}$
Solar fluid pump (SFP)	${\dot{m}}_4{h}_4+{\dot{W}}_{SFP}={\dot{m}}_5{h}_5$
Solution pump (SP)	${\dot{m}}_{15}{h}_{15}+{\dot{W}}_{SP}={\dot{m}}_{16}{h}_{16}$

.

The energy balance provides no information on energy resource degradation during the process and is silent about the quantification of the quality of various energy and material streams considered as input flow and output flow of the control volume; therefore, in order to overcome these limitations, exergy analysis is also incorporated as it determines the system inefficiencies by pinpointing the sources of irreversibilities. Exergy can be regarded as the maximum theoretical useful work achieved through the integration of the system and its environment when the system reaches a dead state.

In general, the balance of exergy relation is applicable to each component of the system following the assumptions described above [28] and can be presented as follows:

$${\displaystyle \sum }\left(1-\frac{{\mathrm{T}}_0}{\mathrm{T}}\right){\dot{\mathrm{Q}}}_{\mathrm{j}}-{\displaystyle \sum }\dot{\mathrm{W}}+{ {\displaystyle \sum }\dot{\mathrm{m}}}_{\mathrm{i}}{\mathrm{e}}_{{\mathrm{x}}_{\mathrm{i}}}-{{\displaystyle \sum }\dot{\mathrm{m}}}_{\mathrm{o}}{\mathrm{e}}_{{\mathrm{x}}_{\mathrm{o}}}-{\dot{\mathrm{E}}}_{\mathrm{x},\mathrm{D}}=0$$

(4)

Total exergy of a flowing stream is mainly the sum of its physical and chemical exergies and may be expressed as:

$${\dot{\mathrm{E}}}_{\mathrm{x},\mathrm{tot}}={\dot{\mathrm{E}}}_{\mathrm{x},\mathrm{ph}}+{\dot{\mathrm{E}}}_{\mathrm{x},\mathrm{ch}}$$

(5)

The components ${\mathrm{e}}_{\mathrm{x},\mathrm{ph}}$ and ${\mathrm{e}}_{\mathrm{x},\mathrm{ch}}$ are the specific physical and chemical exergies which can be defined as:

$${\mathrm{e}}_{\mathrm{x},\mathrm{ph}}=\left\{\left(h-h_0\right)-T_0\left(s-s_0\right)\right\}$$

(6)

$$\begin{array}{ll}{\mathrm{e}}_{\mathrm{x},\mathrm{ch}}(N{H}_3/H_{2}O)& =\dot{m}[\left(\frac{X}{MN{H}_3}\right)e_{ch,}^{0}N{H}_3-T_0\left(\frac{X}{M{H}_{2}O}\right)e_{ch,}^{0}N{H}_{2}O\\ & +R{T}_0{\displaystyle \sum }Xln\left(X\right){\displaystyle ]}\end{array}$$

(7)

where $e_{ch,}^{0}N{H}_3$ and $e_{ch,}^{0}N{H}_{2}O$ refer to standard chemical exergies of ammonia–water. The term $R{T}_0{\displaystyle \sum }Xln\left(X\right)$ is the chemical exergy associated with mixing binary fluids, which was found to be insignificant in this investigation. The above expressions were useful in formulating the component energy and exergy balances along with the exergy efficiencies of the components in Figure 2, as depicted in Table 1.

Determining the balance of exergy in view of the above expression provides formulation for the evaluation of exergy dissipation in each element of the system considered, as reported in

Table 3. Equations of exergy balance for the key element system [24,30]).

Component	Exergy Balance Rate
Heliostat (Helio)	${\dot{E}}_{solar}={\dot{E}}_{CR}+{\dot{E}}_{d,Halio}$
Central Receiver (CR)	${\dot{m}}_5{e}_5+{\dot{E}}_{CR}={\dot{m}}_1{e}_1+{\dot{Q}}_{loss,CR}\left(1-\frac{T_0}{T_{CR}}\right)+{\dot{E}}_{d,CR}$
Process Heater (PH)	${\dot{m}}_1{e}_1={\dot{m}}_2{e}_2+{\dot{E}}_P+{\dot{E}}_{d,P} ; \& {\dot{E}}_P={\dot{Q}}_P\left(1-\frac{T_0}{T_P}\right)$
Expander (Exp)	${\dot{m}}_6{e}_6={\dot{m}}_7{e}_7+{\dot{W}}_{Exp}+{\dot{E}}_{d,Exp}$
Superheater (SH)	${\dot{m}}_2{e}_2+{\dot{m}}_{19}{e}_{19}={\dot{m}}_3{e}_3+{\dot{m}}_6{e}_6+{\dot{E}}_{d,SH}$
Evaporator (Eva)	${\dot{m}}_{11}{e}_{11}={\dot{m}}_{12}{e}_{12}+{\dot{Q}}_{Eva}\left(\frac{T_0}{T_{ev}}-1\right)+{\dot{E}}_{d.Eva}$
Condenser (cond)	${\dot{m}}_1{e}_1={\dot{m}}_2{e}_2+{\dot{E}}_{d,Cond}$
Separator (Sep)	${\dot{m}}_{18}{e}_{18}={\dot{m}}_{19}{e}_{19}+{\dot{m}}_{20}{e}_{20}+{\dot{E}}_{d,Sep}$
Subcooler (SC)	${\dot{m}}_9{e}_9+{\dot{m}}_{12}{e}_{12}={\dot{m}}_{10}{e}_{10}+{\dot{m}}_{13}{e}_{13}+{\dot{E}}_{d,SC}$
Generator (Gen)	${\dot{m}}_3{e}_3+{\dot{m}}_{17}{e}_{17}={\dot{m}}_4{e}_4+{\dot{m}}_{18}{e}_{18}+{\dot{E}}_{d,Gen}$
Flash chamber (FC)	${\dot{m}}_{21}{e}_{21}={\dot{m}}_{22}{e}_{22}+{\dot{m}}_{23}{e}_{23}+{\dot{E}}_{d,FC}$
Solution heat exchanger (SHX)	${\dot{m}}_{22}{e}_{22}+{\dot{m}}_{16}{e}_{16}={\dot{m}}_{17}{e}_{17}+{\dot{m}}_{24}{e}_{24}+{\dot{E}}_{d,SHX}$
Absorber (Abs)	${\dot{m}}_{14}{e}_{14}={\dot{m}}_{15}{e}_{15}+{\dot{E}}_{d,Abs}$
Throttle valve1 (TV1)	${\dot{m}}_{20}{e}_{20}={\dot{m}}_{21}{e}_{21}+{\dot{E}}_{d,TV1}$
Throttle valve2 (TV2)	${\dot{m}}_{24}{e}_{24}={\dot{m}}_{25}{e}_{25}+{\dot{E}}_{d,TV2}$
Throttle valve3 (TV3)	${\dot{m}}_{10}{e}_{10}={\dot{m}}_{11}{e}_{11}+{\dot{E}}_{d,TV3}$
Solar fluid pump (SFP)	${\dot{m}}_4{e}_4+{\dot{W}}_{SFP}={\dot{m}}_5{e}_5+{\dot{E}}_{d,SFP}$
Solution pump (SP)	${\dot{m}}_{15}{e}_{15}+{\dot{W}}_{SP}={\dot{m}}_5{e}_5+{\dot{E}}_{d,SP}$

below.

2.2.2. Overall Performance Analyses

To evaluate the performance of a proposed tower solar collector-driven trigeneration system from energy and exergy perspectives, the following parameters indicating its overall performance were considered: system energy utilization factor (EUF), which can be stated as the ratio of energy delivered to energy required to drive the system and is expressed as defined in [31]:

$${\mathsf{\eta }}_{\mathrm{EUF}, \mathrm{Trigen}}=\frac{{\dot{W}}_{\mathit{el}}+{\dot{Q}}_P+{\dot{\mathrm{Q}}}_{\mathrm{Eva}}}{{\dot{\mathrm{Q}}}_{\mathrm{solar}}}$$

(8)

where ${\dot{W}}_{\mathit{el}}$ represents the electrical power generated, ${\dot{Q}}_P$ is the rate of process heat generated, ${\dot{Q}}_{\mathit{ev}}$ is the rate of cooling produced, and ${\dot{\mathrm{Q}}}_{\mathrm{solar}}$ is solar heat required to operate the trigeneration system.

The electrical power generated by the system can be obtained as described in [19] using the following equation:

$${\dot{W}}_{\mathit{el}}={\mathsf{\eta }}_{\mathrm{gen}}{\dot{\mathrm{W}}}_{\mathrm{Exp}}-\frac{{\dot{\mathrm{W}}}_{\mathrm{SFP}}}{{\mathsf{\eta }}_{\mathrm{motor}}}-\frac{{\dot{\mathrm{W}}}_{\mathrm{SP}}}{{\mathsf{\eta }}_{\mathrm{motor}}}$$

(9)

${\dot{\mathrm{W}}}_{\mathrm{Exp}}$ is the power generated by the expander, ${\dot{\mathrm{W}}}_{\mathrm{SFP}}$, and ${\dot{\mathrm{W}}}_{\mathrm{SP}}$, denote the power consumed by the solar fluid pump (SFP) and solution pump (SP), respectively. ${\mathsf{\eta }}_{\mathrm{gen}}$ is the efficiency of the electrical generator coupled to the expander and ${\mathsf{\eta }}_{\mathrm{motor}}$ is the efficiency of the motor coupled to the shaft of the pumps.

The efficiencies of the electrical generator and motor were considered to be 95%.

The process heat load, ${\dot{Q}}_P,$ is given by:

$${\dot{Q}}_P={\dot{m}}_{27}{h}_{27}-{\dot{m}}_{26}{h}_{26}$$

(10)

The cooling effect (${\dot{\mathrm{Q}}}_{\mathrm{cooling}}$) delivered by the NH₃-H₂O operated cycle can be determined using the following equation:

$${\dot{\mathrm{Q}}}_{\mathrm{Eva}}={\dot{m}}_{12}{h}_{12}-{\dot{m}}_{11}{h}_{11}$$

(11)

The EUF of the trigeneration system, which is expressed by Eq. (8), assesses the performance based on input/output variables and weights the electrical power, process heating, cooling effect, and rate of heat supplied equally, although four energies are distinct from a quality point of view [19,27]. Contrary to EUF, efficiency based on exergy is considered the more appropriate criterion as it determines the quality of energy reserves that can be transformed into useful outputs by the system employed for energy conversion [33].

The ratio of exergy produced to exergy accompanied by primary energy input is known as the exergy efficiency. When the energy required to operate the energy conversion system comes from solar, then the parameter exergy efficiency needs to be analyzed because in such systems, the remaining solar fluid exergy is returned to the receiver instead of being released into the environment. Therefore, the efficiency of the trigeneration operating system mode to produce electricity, process heating, and cooling based on exergy may be evaluated as:

$${\mathsf{\eta }}_{\mathrm{ex}, \mathrm{Trigen}}=\frac{{\dot{W}}_{\mathit{el}}+{\dot{E}}_{x,P}+{\dot{\mathrm{E}}}_{x,Eva}}{{\dot{E}}_{x,solar}}$$

(12)

where ${\dot{\mathrm{E}}}_{x,P}$ is the process heating exergy (${\dot{\mathrm{Q}}}_P$) and may be expressed as:

$${\dot{E}}_{x,P}={\dot{\mathrm{Q}}}_P\left(1-\frac{T_o}{T_P}\right)$$

(13)

where $T_P$ denotes the temperature of process heat generation.

${\dot{E}}_{x,ev}$ is the cooling exergy rate and is described as the variation in the exergy of refrigerant that occurs while passing through the evaporator of the trigeneration system. It can be expressed as:

$${\dot{\mathrm{E}}}_{x,\mathit{Eva}}={\dot{\mathrm{Q}}}_{Eva}\left(\frac{T_o}{T_{Eva}}-1\right)$$

(14)

$${\dot{E}}_{x,\mathit{solar}}={\dot{\mathrm{Q}}}_{\mathrm{solar}}{\left(1-\frac{{\mathrm{T}}_{\mathrm{o}}}{{\mathrm{T}}_{\mathrm{s}}}\right)}_{ }$$

(15)

where ${\mathrm{T}}_{\mathrm{s}}$ is the apparent sun temperature, which is not determined but is assumed in line with the values of the sun’s temperature for the Gulf region reported in the literature [11,19,31], and ${\dot{\mathrm{Q}}}_{\mathrm{solar}}$ is the thermal power associated with falling solar radiation which we consider the primary energy supply for operating the trigeneration system and is given by:

$${\dot{\mathrm{Q}}}_{\mathrm{solar}}={\mathrm{A}}_{\mathrm{field}}\mathrm{q}$$

(16)

where q is the direct normal irradiation (DNI), which is described as the solar energy that falls on a one square meter area, and ${\mathrm{A}}_{\mathrm{field}}$ is the heliostat field.

In this analysis, exergy accompanied by incoming solar heat and supplied to the trigeneration system is divided into exergy produced and destroyed. Exergy loss is also considered and computed accordingly. The employment of global balance of exergy [28] gives:

$${\dot{\mathrm{E}}}_{\mathrm{x},\mathrm{F},\mathrm{k}}={\dot{\mathrm{E}}}_{\mathrm{x},\mathrm{P},\mathrm{k}}+{\dot{\mathrm{E}}}_{\mathrm{x},\mathrm{D},\mathrm{k}}+{\dot{\mathrm{E}}}_{\mathrm{x},\mathrm{L},\mathrm{k}}$$

(17)

where ${\dot{\mathrm{E}}}_{\mathrm{x},\mathrm{p},\mathrm{k}}$  and ${\dot{\mathrm{E}}}_{\mathrm{x},\mathrm{F},\mathrm{k}}$  indicaticate the rates of product exergy and fuel exergy. The terms ${\dot{\mathrm{E}}}_{\mathrm{x},\mathrm{L},\mathrm{k}}$ and ${\dot{\mathrm{E}}}_{\mathrm{x},\mathrm{D},\mathrm{k}}$ represent the loss of exergy and the dissipation of exergy.

In the present investigation, fuel (F) is the heat accompanying the solar energy, therefore, the exergy rate accompanying the collected solar energy $\left({\dot{\mathrm{E}}}_{x,\mathit{solar}}\right)$  is ${\dot{\mathrm{E}}}_{\mathrm{x},\mathrm{F},\mathrm{k}}$. The share of supplied solar exergy transformed into exergy generated, exergy dissipated, and exegy lost from the proposed trigeneration system is computed and presented.

3. Results and Discussion

In this paper, the proposed integration of the tower solar collector, the process heat exchanger, and the cooling–power cogeneration cycle is thermodynamically simulated using EES and a theoretical investigation is carried out to identify the operating variables affecting the performance of the system’s first and second laws. Further, the results of the distribution of solar exergy supplied to the proposed system in the form of exergy generated and exergy dissipated due to the irreversibility assigned to components and the exergy accompanied by thermal release into the atmosphere are also determined. Baseline properties for the trigeneration system’s operation conditions are defined in

Table 4. Assumptive operating conditions for energy and exergy analyses of the developed CCHP system in the present study [11,31,34].

Operating Parameter	Value
Restricted dead state temperature (K)	298
Restricted dead state pressure (MPa)	0.10132
Variation in the outlet temperature of hot oil (K)	433–503
Inlet temperature of hot oil (K)	388
Solar radiations collected on area considered DNI (W/m2)	500–1000
Temperature of Sun (K)	5000
Heliostat field area (m2)	10,000
Isentropic efficiency of expander (%)	95
Concentration ratio	300
Working fluid considered for Kalina cycle	NH3-H2O solution
Isentropic efficiency of pump (%)	95
Heliostat’s energy efficiency (%)	75
Exergy based efficiency of heliostat field (%)	75
Solution heat exchanger and superheater effectiveness (%)	100

, with relevant references from the literature cited. Solar irradiation and ambient temperature have been found to play a dominant role in the type of system employed for power generation or multi-objective generation as they support the selection of suitable solar-based energy systems for operation in different weather conditions. The power generated, cooling effect, cooling exergy, and the efficiencies based on energy and exergy of the cooling–power–heating system of trigeneration were ascertained by varying direct normal irradiation (DNI), hot oil supply temperature, process heat pressure, and ambient temperature.

3.1. Impact of Solar Irradiance and Ambient Temperature

To evaluate the performance of a proposed CCHP system, the alteration of solar irradiance (DNI) and ambient temperature must be properly considered as these two parameters change with seasons and play an important role in the selection of solar-based energy generation systems for operation under different climatic conditions. The energy efficiency of the solar power tower field changes with changes in ambient temperature and solar irradiance [11]. The ranges of ambient temperature and solar irradiance are taken from Refs. [11,31] and reflect the range of alteration of these parameters in Saudi Arabia. Figure 2 shows that the energy efficiency of the solar field or solar collector increases with increase in DNI. This can be described by considering the heat loss mechanism from the receiver as two major losses, convection and emission, depend mainly on the surface temperature of the receiver, which increases slightly after a large increase in DNI. Consequently, heat loss from the receiver increases only slightly when input solar energy increases proportionally, and this results in an increase in thermal efficiency of solar field with DNI. Since the exergy assigned to energy losses in the receiver is significantly large due to the fact that solar isolation is designated as very high quality energy and large irreversibilities occur when the solar exergy (very high temperature) is converted to thermal energy of significantly lower temperature in the receiver, increase in the percentage of exergy efficiency of the solar field is greater than its energy efficiency. Figure 2 further shows the impact of changing ambient temperature on energy and exergy efficiency of the solar field and indicates that both energy and exergy efficiencies improve based on the increase in ambient temperature. This is due to the fact that increasing the ambient temperature at a given DNI results in a decrease in the difference in temperature between the receiver surface and ambient temperature and hence heat losses are reduced, which in turn increases the collector energy efficiency. The increase in exergy efficiency of the collector is observed due to a decrease in heat transfer irreversibilites at the receiver with increasing ambient temperature.

Figure 3 shows the influence of promoting DNI on CCHP system’s energetic and exergetic outputs. We determined that the rise in DNI of both electrical power (${\dot{W}}_{el}$) and cooling capacity (${\dot{Q}}_{ev}$) are promoted while process heat (${\dot{Q}}_p$) remains unaltered. Electrical power and cooling capacity increased from 3.087 kW to 26.03 kW and from 216.4 kW to 1824 kW, respectively, as DNI increased from 500 W/m² to 1000 W/m². This is due to the fact that increasing the mass flow rate of the system working fluid increases with increasing DNI, leading to the generation of greater system outputs. This means larger availability of DNI results in greater energetic output of the trigeneration system. As the mass flow rate and the pressure of the process heater for a given enthalpy change in solar heat transfer fluid (oil) are unchanged, the process heating capacity remains constant at 2461 kW. Further, power exergy (${\dot{W}}_{el}$) and cooling exergy (${\dot{E}}_{x, ev}$) improve significantly with the increase in DNI while process heating exergy (${\dot{E}}_{x, p}$) remains unchanged. Since the power exergy of the system is generated electrical power and cooling exergy is significantly less than the cooling capacity, cooling exergy improves significantly from 11.27 kW to 95.02 kW following an increase in cooling capacity at a constant process heating exergy (${\dot{E}}_{x, p}$) of 2272 kW corresponding to a change in DNI from 500 W/m² to 1000 W/m².

Figure 4 depicts the changes in EUF and exergy efficiency of CCHP for different DNI conditions. According to this display, the EUF declines from 53.62% to 43.12% as DNI changes from 500 W/m² to 1000 W/m². The reason for this decline is the increase in the amount of solar energy required to drive the system to produce power, heating, and cooling. Since the rate of increase in solar heat supply dominates over the rate of increase in electrical power and cooling capacity, the system EUF reduces as DNI increases. The exergy efficiency of trigeneration system shares the same declining trend and is sharply reduced from 49.02% to 25.65% as DNI increases from 500 W/m² to 1000 W/m². The exergy efficiency of the system is less than its energy efficiency or EUF because exergies for process heating and cooling are significantly less than the amount of process heat and cooling produced by the system. The overall energy and exergy efficiencies of the developed trigeneration system were 53.62% and 25.65%, which are on the lower side considering trigeneration studies reported in the literature [16], where energy and exergy efficiencies were 60.12% and 26.54%, respectively. This difference between the present results and the previously reported ones is due to the use of different solar collector technologies (tower solar collector instead of a parabolic trough collector) and the employment of a distinct working fluid (NH₃-H₂O instead of a hydrocarbon) for the generation of power and cooling.

The influence of ambient temperature on energy and exergy outputs of the system is displayed in Figure 5, which shows that promotion of ambient temperature results in an increase in cooling capacity (${\dot{Q}}_{ev}$) and a simultaneous increase in cooling exergy (${\dot{E}}_{x, ev}$), while electrical power (${\dot{W}}_{el}$) generation is reduced. The increase in cooling capacity at higher ambient temperatures accounts for the benefit in the proposed system because greater cooling is needed when days are hotter. Consideration of specific cooling capacity of the evaporator and the mass flow rate can justify the growing cooling capacity with the promotion of atmospheric temperature. Increase in ambient temperature decreases the specific cooling capacity and increases the mass flow rate in such a way that refrigeration, which is the product of these two, is increased as ambient temperature rises. Power output and power exergy decrease because increase in ambient temperature reduces changes in expander pressure.

Figure 6 shows the evolution of EUF and exergy efficiency of the system versus ambient temperature. System EUF increases with the rise in ambient temperature because increased cooling capacity(${\dot{Q}}_{ev}$) is greater than the decrease in expander power (${\dot{W}}_{el}$) for a given supply of solar heat to the system. The opposite trend is observed for the exergy efficiency of the system due to the fact that an increase in exergy accompanied by cooling capacity (${\dot{E}}_{x, ev}$) is less than the reduction in power produced by the expander (${\dot{W}}_{el}$). Moreover, increase in ambient temperature results in a reduction of the differences in temperature between the collector and the environment, which reduces exergetic losses due to heat transfer at the collector and an increase in solar exergy supplied to the system contributes further to the decline of the system’s exergy efficiency.

3.2. Impact of Hot Oil Supply Temperature

The temperature of hot oil supplied to the CCHP system is considered to play an important role in the analysis of single-generation or multi-generation systems. Therefore, the proposed system is examined at different inputs of hot oil supply temperatures, from 172.5 °C to 228 °C, as shown in Figure 7. Since the difference in the temperature of hot oil passing through the heat exchanger is unaltered, the amount of process heat generated (${\dot{Q}}_p$) remains constant. Increased hot oil temperature results in an increased mass flow rate of fluid at the expander entry, raising the power output (${\dot{W}}_{el}$). Due to the introduction of the weak solution from the exit of the generator for flashing, an additional refrigerant, which increases when the hot oil supply temperature is increased, is produced, hence it offers an increase in cooling capacity (${\dot{Q}}_{ev}$). Referring to Figure 7, the increase in the temperature of the hot oil supplied leads to an increase in power exergy (${\dot{W}}_{el}$) and cooling exergy (${\dot{E}}_{x, ev}$), for obvious reasons. Since the cooling capacity is much greater than the exergy assigned to it, the refrigeration exergy also shows an increasing trend, although the increase is lower than the rise in cooling capacity. The effect of an increase in the hot oil exit supply temperature on EUF and exergy efficiency of the system is displayed in Figure 8. The calculated energy and exergy efficiencies of the system decreased as oil temperature increased due to the large difference in temperature between the system boundary and the environment, leading to greater heat loss. The trend of decreasing energy and exergy efficiencies of the system correlates with the trends observed in Refs. [12,37], where system efficiencies also decreased with increasing temperature of the solar heat transfer fluid due to a higher rate of heat loss from the receiver of the heliostat field.

3.3. Impact of Process Heat Pressure

Relative to a CCHP system combining ORC with absorption chiller, tower solar collector-based process heat exchanger combined with NH₃-H₂O cooling cogeneration cycle operates with a high-pressure process heat exchanger. Moreover, it is considered an important design parameter. The impact of process heat pressure on cycle performance is discussed and shown in Figure 9. Computation of results revealed that the influence of varying the process heat pressure on the system’s first law performance is insignificant. This is due to considering a constant temperature difference between hot oil supply at the inlet and outlet states of the heat exchanger. In this situation, the variation of process heat pressure, the enthalpy, and mass flow rate of steam vary in such a way that their product, which is the amount of process heat generated, remains constant with increasing process heat pressure. Therefore, its impact on the system’s first law performance was not investigated. Analysis carried out after employing the second law determined that an increase in process heat pressure from 100 kPa to 250 kPa would lead to an increase in process heating exergy (${\dot{E}}_{x,p}$) from 2263 kW to 2275 kW owing to reduction in the entropy of steam as pressure increases. Since process heating exergy is one of the most valuable outputs of the proposed CCHP system, an increase in process heat pressure leads to an enhancement of system exergy efficiency, although the increase is small, as shown in Figure 9.

The distributions of solar energy and exergy input in the proposed CCHP system are displayed in Figure 10. These depictions clearly demonstrate how solar exergy and energy are utilized, lost, and dissipated by system components. It shows that from the 100% solar energy collected by the system, 46.7% is delivered as energetic output composed of 15.11% cooling capacity (${\dot{Q}}_{ev}$), 30.7% process heating (${\dot{Q}}_p$), and 0.22% power output (${\dot{W}}_{el}$). The remaining energy is lost via heat transfer to the surroundings of the condenser, absorber, central receiver, and heliostat. The majority of energy loss occurs in the heliostat (25%) and central receiver (17.19%), while the absorber (24.5%) and condenser (17.5%) are also significant sources of energy loss from the system. Since various loss mechanisms such as losses due to cosine efficiency, blocking and shading, mirror reflectivity, and tracking error are associated with solar energy incident on the heliostats, the highest energy loss is observed at the heliostat. Because of the significant transfer of energy from the receiver surface to the environment by reflective, emissive, convective, and conductive heat loss, the receiver exhibits the second-largest energy loss. Because of the significant heat carried away by the cooling fluid, energy loss from the absorber and condenser is also significant compared with other system components.

Analysis of the second law provides qualitative exergy flow representation for the developed CCHP and the resulting exergy shows that from the 100% solar exergy supplied to the system, around 7.27% is converted to process heating exergy (${\dot{E}}_{x,p}$), 0.84% is converted to refrigeration exergy (${\dot{E}}_{x, ev}$), and 0.23% is converted to power exergy (${\dot{W}}_{el}$). The remaining solar exergy is dissipated through generation of entropy in various system components and losses to the atmosphere via thermal exhaust. Net entropy increases during the high thermal transformation process but at the same time the exergy should not be negative. The positive exergy loss indicates the feasibility of the system proposed in this study. Further discretization of solar exergy pertaining to system components shows that the central receiver contributes the most to exergy destruction in the system (56.89%), followed by the heliostat (25%), process heater (3.7%), absorber (2.95%), and condenser (1.83%). The exergy destroyed in the central receiver of the present system is 56.89% compared with 55.47% obtained from Ref. [12]. This little deviation between the calculated results and reported results in the literature is observed due to the application of different solar heat transfer fluids and because of the larger temperature differences between the system boundary and the surroundings. The calculated exergy loss at the heliostat is 25%, which is perfectly aligned with the amount of exergy loss at the heliostats determined by previous investigators [14,38]. This similarity is obtained because most of the investigators considered the same field efficiency of 75%. Computation of exergy destruction in the components indicates that even a small improvement in central receiver or heliostat efficiency would result in an efficient and climate-friendly operation of a solar-based energy generation system. Dissipation of exergy in the rest of the system components was not significant, and therefore, not worth mentioning. The central receiver and heliostat provide the largest exergy dissipation because in the heliostat field, solar insolation possess high quality energy due to extremely high temperature (T = 5000 K) and this high quality insolation is absorbed by solar heat transfer fluid in the form of thermal energy at lower temperatures (around 550–700 K). This large temperature difference results in high convective and radiative energy losses at the collector which is of high quality and hence a lot of solar exergy is lost at the solar field. The next largest solar exergy dissipation is attributed to process heater (3.7%) and this mainly occurs because of the considerable difference in temperature between the hot supply oil and the process heating fluid. The exergy destroyed in the process heater of the present system is lower (4.3%) than the value (3.7%) reported in the literature [20]. This can be attributed to the use of different working fluids and the distinct position of the process heater in the trigeneration system. The dissipation of exergy caused by the absorber is considerably greater because of the higher mass flow rate of the working fluid. This is because flashing generates additional refrigerant, which mixes with the refrigerant produced at the evaporator during cooling; thus, the mass flow rate of stream entering the absorber is large, leading to greater loss of exergy. Due to the lack of mixing of fluid streams and a lower temperature difference, the loss of exergy at the condenser is lower than the absorber, as mentioned above.

3.4. Validation of Proposed Methodology

Since experimental data from a similar trigeneration system (combined cooling–power–heating) is lacking in the literature, an alternative approach in which the results derived from the present model are compared with results obtained from existing technologies available in Ref. [38] was adopted to evaluate the accuracy of the proposed modeling. Due to the fact that a system that is similar to the present model was not found in the literature, direct comparison of the overall efficiencies and component performance parameters with other trigeneration systems was not viable. Hence, a comparison of the energy efficiency or the cycle energy utilization factor and output parameters of the major subsystem (cooling–power) cycle has been provided. The parameters considered for validation of the results obtained in the present study are mentioned in

Table 5. Validation of the proposed methodology using existing technologies [38].

Description	Lu and Goswami (2003) [38]	Present Methodology with Same Working Conditions	Deviation, %
Source pressure, bar	13.00	12.80	1.50
Sink pressure, bar	5.50	5.45	0.91
Strong solution concentration, kg/kg	0.6733	0.6737	−0.06
Turbine exit temperature, °C	7.60	7.20	5.20
Generator heat input, kW	272.90	295.90	−8.40
Absorber heat rejection, kW	269.10	301.40	−12.00
Power output, kW	21.00	20.4	2.85
Vapor quality at turbine exit, %	93.93	96.00	−2.20
Energy Utilization Factor (EUF)	13.26	13.47	1.58
Cooling output, kW	16.20	16.90	−4.32
Total output, kW	37.20	37.84	6.66
Hot fluid supply temperature, °C	87.00	85.00	2.30
Hot fluid exit temperature, °C	57.20	54.30	5.00
Hot fluid mass flow rate, kg/s	2.183	2.30	−5.36

. Simulation using the Engineering Equation Solver (EES) was conducted for the same operating conditions considered in Ref. [38] using the proposed methodology. The cooling output and power output produced by the evaporator and turbine of the present system were 16.9 kW and 20.4 kW, respectively. These values are similar to the reported cooling output value of 16.2 kW and power output value of 21 kW. The cooling output value in the present work was a little higher because the installation of a flashing system generated additional refrigerant that increases the cooling capacity. Due to the higher vapor quality at the turbine outlet, which reduces the difference in enthalpy during expansion, the power output of the present system was a little lower than the reported value in Ref. [38]. The EUF (Energy Utilization Factor) of the proposed cooling–power system was 13.47, which is a little higher than the EUF value of 13. 26 of the system used for validation in Ref. [38]. This deviation occurred because more cooling energy was produced due to the employment of a flashing chamber. Apart from generator heat supply and absorber heat rejection, the remaining deviations in state point’s values and output are less than 5%, which is positive in the proposed methodology. Maintaining high turbine efficiency and low temperature difference in the absorber will reduce the higher deviation in generator heat supply and absorber heat rejection. The analysis and comparison of results show good performance of the model developed in this study, as the computed results correlate with values from the literature. However, minor deviation with a low percentage, which is attributed to small differences in the assumptive operating conditions, is observed.

4. Conclusions

This study introduces a new solar-based trigenerative system and investigates its performance and outputs under different operative variables of DNI, ambient temperature, hot oil supply temperature, and process heat pressure. The exergy efficiencies of the solar collector increased the DNI from 500 W/m² to 1000 W/m², reducing system EUF and exergy efficiency from 53.62% to 43.12% and from 49.02% to 25.65%, respectively. It was noted that increasing DNI and ambient temperature resulted in higher power exergy and exergy of refrigeration while keeping the heating exergy constant. Raising the hot oil supply temperature has a positive impact on power output and cooling capacity, but has the opposite effect on the EUF and exergy efficiency of the system. Employment of a flash chamber resulted in generation of additional refrigerant without additional heat input, which in turn enhanced cooling; therefore, the proposed system resulted in higher EUF compared with the conventional cooling–heating–power trigeneration plant. Exergy analysis of the system illustrates that the central receiver contributes to most of the exergy dissipation (56.89%), the process heater accounts for the second-largest dissipation (3.7%), while exergy destruction caused by other components of the system is less than 1%. Heliostat produces the largest exergy loss (25%), followed by absorber (2.95%), and condenser (1.83%). This data implies that the central receiver and the heliostat are the largest sources of overall system inefficiency and therefore, these two components require special attention during selection and design. Discretization of 100% radiative solar energy/exergy supplied to the proposed system produces energetic output of 46.03% and exergetic output of 8.34%, while the remaining is energy/exergy losses discharged to the atmosphere; the exergy destroyed due to the operation of system components occurs in an irreversible fashion.