Energy Conversion and Management 228 (2021) 113677

Contents lists available at ScienceDirect

Energy Conversion and Management

journal homepage: www.elsevier.com/locate/enconman

Investigating the effects of integrating an absorption heat transformer with

a combined cooling, heating and power system: A thermodynamic and
economic analysis
Saman Khalilzadeh a, Alireza Hossein Nezhad a, *, Alessandro Romagnoli b,
Bakytzhan Akhmetov b
a
Department of Mechanical Engineering, University of Sistan and Baluchestan, Zahedan, Iran
b
School of Mechanical and Aerospace Engineering, Nanyang Technological University, Singapore 639798, Singapore

A R T I C L E I N F O A B S T R A C T

Keywords: Regarding the significance of low-temperature heat sources in combined cooling, heating, and power (CCHP)
Absorption heat transformer systems, it is necessary to find practical and economical ways to maximize the efficient use of available energies.
Organic Rankine cycle Integrating the auxiliary cycle with the main system and changing the configuration are among the useful so­
Absorption refrigeration cycle
lutions to achieve this purpose. In this work, a combination of both solutions is presented, accordingly, an ab­
Low-temperature heat source
sorption heat transformer cycle (AHT) is integrated with a conventional CCHP system, which is made of organic
Combined cooling heating and power
Rankine cycle, absorption refrigeration cycle, and a heat exchanger. Three conventional CCHP systems with
sequential (Configuraion1), parallel (Configuraion2) and a combination of sequential and parallel (Config­
uraion3) configurations are considered. It is shown that among the Configuraion1, Configuraion2, and Config­
uraion3, Configuraion1 has the highest energy efficiency; therefore, it is selected to be compared with the
proposed integrated system with AHT. This integration affects the outputs production and also energy con­
sumption; therefore, four different modes (standpoints) are taken into account to compare the production of
power (mode A), cooling (Mode B), heating (mode C), and energy consumption (mode D) of these two systems.
Also, from the exergy, environmental, and economic aspects, these two systems are compared. Regarding mode
D, it is indicated that the energy and exergy efficiencies, energy consumption and carbon emission of the pro­
posed system are improved 17.68%, 17.68%, 15.03%, and 15.02%, respectively. It is noted that for the heat
source stream temperature range of 90 ◦ C–120 ◦ C, the proposed system has better performance than the
Configuration1. Despite the additional costs of integrating AHT with the CCHP system, the proposed system has
517018.1 US$/year cost saving (which is 32.65% of the investment cost of Configuration1) if biomass is used as a
heat source, which indicates its higher economic advantage than the Configuration1. Also, comparing the pro­
posed system with Configuration1, the amount of power, cooling and heating productions are improved 27.98%,
102.15%, and 36.87%, respectively.

have been considered in technological improvements of energy systems

1. Introduction (e.g. combined cooling, heating, and power systems). However, finding
cost-efficient ways to increase the maximum use of available waste heat
As it is known, using fossil fuels has destructive effects on the envi­ sources and renewable energies has still been highlighted. Some prac­
ronment. Consequently, science society has started focusing on this tical and cost-efficient ways are integrating conventional energy systems
problem with remarkable attention and performed various research to with auxiliary systems, changing the configuration of energy systems
propose sustainable solutions. Replacing fossil fuels with renewable and/or a combination of both.
energy sources and waste energy can be considered as promising Producing power is one of the suitable solutions to utilize low-
achievements and advances in this field. Some renewable energies and temperature heat sources. In such cases, organic Rankine cycles (ORC)
waste heat have mid- or low-temperature values, and for decades, they are suggested due to the low boiling point of their working fluid. To

* Corresponding author.
E-mail addresses: saman.khalilzadeh@pgs.usb.ac.ir (S. Khalilzadeh), nezhadd@hamoon.usb.ac.ir (A. Hossein Nezhad), A.Romagnoli@ntu.edu.sg (A. Romagnoli),
bakytzhan.akhmetov@ntu.edu.sg (B. Akhmetov).

https://doi.org/10.1016/j.enconman.2020.113677
Received 9 August 2020; Received in revised form 15 November 2020; Accepted 17 November 2020
Available online 27 November 2020
0196-8904/© 2020 Elsevier Ltd. All rights reserved.
S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Nomenclature TCI Total cost investment (US$)

TDI The total depreciable investment (US$)
A The heat transfer area (m2) U Overall heat transfer coefficient (W/m2◦ K)
AC Annual cost (US$) WC Working capital (US$)
AHT Absorption heat transformer W˙ Produced or consumed power (kW)
ANCF Annual net cash flow (US$/year) X Mass concentration of NH3 in NH3-Water or LiBr in LiBr-
AP The average annual profit (US$/year) Water (kg/kg)
AWS Average wind speed (m/s) YD Exergy destruction ratio (%)
ARC Absorption refrigeration cycle Ż The total capital investment cost rate (US$/h)
ARR Average rate of return (%)
BD The annual book depreciation (US$/year) Greek letters
BM Biomass η Efficiency
CDE Carbon dioxide emission (g/kWh) μ Salvage percentage (%)
COP Coefficient of performance ε Exergy efficiency of components
CRF Current recovery factor τ operating hours (hour)
ex Specific exergy (J/kg) Ω Total operating and maintenance cost factor
Ėx Rate of exergy (kW) subscripts
h Specific enthalpy (J/kg) D Destruction
HE Heat exchanger avg Average
i Interest rate (%) ch Chemical
ICR Investment cost rate (US$/h) F Fuel
LiBr Lithium Bromide gen Generator
LMTD Logarithmic mean temperature difference abs Absorber
M Molecular mass eva Evaporator
m˙ Mass flow rate (kg/s) P Product
n Life time of the system (year) ph Physical
NG Natural gas cond Condenser
NH3 Ammonia tot Total
NPV Net present value (US$) k kth component
ORC Organic Rankine cycle sys System
P Pressure (kPa) ex Exergy
PW Present worth (US$) Elect. Electrical
PWF Present worth factor Heat Heating
PP Payback period (year) cool Cooling
Q˙ Rate of heat transfer (kW) CI Capital investment
Q Quality of the stream OM Operating maintenance
s Specific entropy (J/kg K)
SV Salvage value (US$)

achieve higher efficiency, Sun et al. [1] considered a double-pressure as lower COP (coefficient of performance) than ARC. An Ejector refrig­
ORC system and optimized it from thermodynamic and economic eration cycle was modeled and optimized thermally by Riaz et al. [5] to
standpoints. Regarding the geothermal water with a temperature of utilize very low-grade heat sources (60–100 ◦ C). According to the
120 ◦ C as a low-temperature heat source, it was deduced that the high considered case study in Singapore, the COP and cooling production
and low inlet pressures of the turbine have significant roles in improving were evaluated as 0.3 MW and 1.8 MW, respectively, and the related
the efficiency of the system. A small-scale ORC system driven by a low- cost-saving was calculated to be 0.42 S$ million. In another work carried
temperature heat source (120 ◦ C) with the working fluid of R245fa was out by Cola et al. [6], a textile waste heat with a temperature of 170 ◦ C
studied experimentally by Lin et al. [2]. In addition to the effective was considered and three cooling systems, including ARC, ORC-VCC,
parameters on the system performance, the researchers attempted to TEG-VCC (thermoelectric generator), were investigated and compared
show that the scroll expander was appropriate for such systems and heat from viewpoints of COP, cooling production and specific investment
sources. Examining various conditions, the maximum net power was cost. Combining the three mentioned parameters and also comparing
attained up to 6.2 kW. Taking into account the important role of working due to unique index, the authors presented a scoring method in which
fluids in ORC, a study was performed to find out which group of organic the importance of each parameter was determined by the weighting
fluids has better performance [3]. Results demonstrated that wet fluids factors. As a consequence, a suitable range was obtained for each
having a steep saturated vapor curve in the T-s diagram would perform system.
better than dry fluids. Another popular way to use low-temperature heat sources, other
Besides power production using ORC, cooling production applying than producing power using ORC and producing cooling, is to supply
low-temperature heat sources is another appealing way. Lillo et al. [4] some required energy for multi-generation systems. In the following,
analyzed a hybrid ejector refrigerating system from a thermoeconomic some of the recent studies in this field are represented. A combined
standpoint. They endeavored to prove that the mentioned system can cooling and power system with a new configuration was proposed by
perform as efficient as conventional refrigeration systems driven by low- Wang et al. [7]. The system included an absorption power cycle (APC)
temperature heat sources like ORC-VCC (ORC-vapor compression cycle) and an ejector refrigeration cycle (ERC) in which the produced energy of
and absorption refrigeration cycle (ARC). Finally, results revealed that steam in ERC was provided by passing the stream from the generator to
the proposed system had lower investment costs than ORC-VCC as well the absorber of APC. Using the low-grade heat source with the

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S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

temperature of 110 ◦ C, 23.89 kW cooling, and 18.52 kW power were configurations, respectively. In the study [21], two novel configurations
produced, however, the energy and exergy efficiencies were determined of the integrated systems were suggested and compared with the stan­
to be 21.7% and 52.22%, respectively. Combining ORC with ERC, Zhang dard biomass-based combined cycle from energy, exergy, economic and
et al. [8]. presented a novel CCHP system to use low-temperature heat environmental standpoints. In both configurations, produced syngas in
sources. The system performance was assessed thermodynamically and the gasifier (from biomass) was utilized in the Brayton cycle, however,
the results showed that the exergy efficiency and COP of the system were solar energy was used in ORC. In configuration 1, the solar energy used
13.3% and 167.7% higher than the reference system, in which the indirectly to warm up the stream before entering the open feedwater
Brayton cycle was used instead of ORC. In the study [9], a new CCHP heater. On the other hand, in configuration 2, using solar energy, the
system including the Brayton cycle, Rankine cycle, modified Kalina/ produced vapor in the solar cycle was used directly to increase the en­
vapor-compression refrigeration system and water heating unit was ergy of the stream before entering the low-pressure turbine. Although
proposed to use the textile wastewater as the low-temperature heat the configuration 2 had the lowest energy and exergy efficiencies, it had
source. The rendered system was evaluated from energy and economic the lowest Levelized cost of energy (LCOE) and CO2 emission.
viewpoints. Finally, the results demonstrated that the payback period Regarding the literature review energy systems driven by low-
and energy efficiency were obtained as 5 years and 62%, respectively. temperature heat sources have been remarkably concerned. Accord­
Khalilzadeh et al. [10] used the low-temperature waste heat of an R744/ ingly, various studies have been carried out to improve the performance
R717 cascade refrigeration cycle in an integrated system coupled with a of energy systems by presenting new configurations or by integrating
parabolic trough solar collector and an ORC. Therefore, in addition to auxiliary systems. Even though AHT systems are suitable and appealing
producing cooling, power and hot water were also provided. Assuming options as auxiliary systems to be used in energy systems, according to
12 h of solar energy per day, it was reported that the proposed system the best knowledge of authors, there are few studies on the integration of
would produce 132.72 m3/day hot water and 2400 kWh/day cooling. AHT with CCHP systems. In this work, as a practical and economical way
The above studies have shown how researchers tried to utilize low- to enhance the performance of conventional CCHP systems driven by
temperature heat sources in different energy systems. However, in the low-temperature heat sources, a novel CCHP system integrated with
following, the studies focused mostly on the utilization of the available AHT is proposed and studied comprehensively. Therefore, the in­
energy source by introducing auxiliary systems, and the configuration of novations and scientific values of this novel work are named as
energy systems to improve their performance. Absorption heat trans­
former (AHT) systems are recognized as appropriate systems to increase • Integrating an AHT cycle with a basic CCHP system, and accordingly,
the temperature of waste heat or low-temperature heat sources. In proposing a new CCHP system with a novel configuration to achieve
various studies, to enhance the performance and capability of AHTs, better performance especially at very low temperatures.
these systems have been analyzed from viewpoints of energy [11,12], • Investigating the effects of this integration on the produced power,
exergy [13], exergoeconomic [13], and type of working fluid [14,15]. cooling, heating, and consumed energy from the energy and exergy
Meanwhile, according to AHTs special characteristics, in other studies, it standpoints
has been coupled to other systems as an auxiliary system to increase the • Investigating the effects of this integration on carbon dioxide emis­
low-temperature streams and improve the efficiency of the total system. sions, which is important from an environmental perspective.
Parham et al. [16] studied and compared the performance of three • Investigating the effects of this integration on the economic aspects
different types of AHTs (including single, double and triple AHTs) and cost savings
coupled with a water desalination system. It was deduced that although • Comparing the proposed system with basic CCHP systems that have
the triple AHT reached the highest temperature, it had the lowest COP conventional configurations from energy and economic viewpoints
compared with two other cases. Furthermore, assuming all cases
working full time, the single, double and triple AHTs could produce 2. Problem statement
potable water for 853, 796, and 697 residentials, respectively. Behnam
et al. [17] proposed a small-scale trigeneration system consisting of an As it is known, there are many CCHP systems with different config­
AHT, ORC and single-stage evaporation desalination which was driven urations. However, in general, a basic CCHP system consists of a power
by a geothermal heat source with a temperature of 100 ◦ C. The generation cycle and a cooling generation cycle, in addition to some
geothermal heat source supplied the required energy of AHT and ORC components (e.g. heat exchangers) used to generate heating. In the basic
systems, on the other hand, the passing seawater through the absorber of CCHP system of this work, the considered cycles and components to
AHT converted to steam and provided the energy of the desalination produce power, cooling and heating are ORC, ARC, and a heat
system. As a result, the produced fresh water, power, and heating were exchanger, respectively. These cycles and components can be configured
reported as 0.662 kg/s, 161.5 kW, and 246 kW, respectively. As in different ways to make a basic CCHP system. Among all possible
mentioned, one way to improve the performance of an energy system is configurations for a basic CCHP system, the sequential, parallel, and a
to find the best configuration. Roumpedakis et al. [18] considered four combination of sequential and parallel configurations are the most
different configurations to improve the performance of the multi­ common configurations. The mentioned configurations are considered
generation systems based on ORC. Comparing the proposed configura­ as criteria to be compared with the proposed system in this work. To
tions with the ORC-VCC as the reference system, it was revealed that the simplify the expression of their names, the basic CCHP system with
configuration with two expanders had higher energy and exergy effi­ sequential configuration, the basic CCHP system with parallel configu­
ciencies. In the study [19], three different configurations, one sequential ration, and the Basic CCHP system with a combination of sequential and
system and two parallel systems, for a CCHP system driven by solar parallel configuration are considered as Configuration1, Configuration2,
energy were proposed. In the sequential system, solar energy was only and Configuration3, respectively. The proposed system is the integration
used to provide the required energy for the ORC cycle, and the residual of the Configuration3 and an AHT in which the AHT is placed between
energy in the ORC cycle was employed for heating and cooling. Among the heat source and the Configuration3.
the proposed configurations, the sequential configuration had the lowest To investigate and understand the improvements of the proposed
installation space and the payback period. Considering a low-to-mid system, the Configuration1 is selected to be compared with the proposed
grade geothermal heat source, various configurations for a CCHP sys­ system due to its highest energy efficiency and lowest investment cost
tem were presented and compared from energetic and exergetic view­ rate between the considered configurations. To clarify that the proposed
points by Martinez et al. [20]. The hybrid parallel-series cascade and the system has enough advantages compared with the Configuration1, it is
series cascade arrangements had the maximum exergy efficiencies at studied from different viewpoints:
temperature ranges of 80–110 ◦ C and 110–150 ◦ C among the presented

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S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Fig. 1. Schematic of the Configuration1.

• Energy viewpoint: What effects does it have on the production of 2.2. Mode B: effects on cooling production
power, heating, cooling, and also energy efficiency?
• Economic viewpoint: Definitely, integrating an auxiliary cycle with a In this mode, the amount of power, heating, and energy consumption
CCHP system requires more investment costs. Is the proposed system of the proposed integrated system are kept fixed with that of the
more cost-efficient than a Basic system when the same amounts of Configuration1, then, the amount of cooling production of these systems
output productions are demanded? are compared.
• Environmental viewpoint: What effect does the mentioned integra­
tion have on energy consumption and carbon production? 2.3. Mode C: effects on heating production

Integrating an AHT with a CCHP system may affect the production of In this mode, the amount of power, cooling, and energy consumption
heating, cooling, power, energy consumption, economic and environ­ of the proposed integrated system are kept fixed with that of the
mental aspects. To compare the proposed integrated system and the Configuration1, then, the amount of heating production of these systems
Configuration1, four different modes (standpoints) are considered to are compared.
show the effects of this integration on the production of heating, cooling,
power and also the consumption of energy.
2.4. Mode D: effects on energy consumption
In each of these modes, one of the heating, cooling, power, or energy
consumption of the proposed integrated system is changed while the
In this mode, the amount of power, cooling, and heating productions
others are kept fixed and equal with that of the Configuration1, and then
of the proposed integrated system are kept fixed with that of the
comparisons are made between these two systems. In the following,
Configuration1, then, the amount of energy consumption of these sys­
more details about the different modes are presented.
tems are compared.
2.1. Mode A: effects on power production
2.5. Description of the configuration1
In this mode, the amount of cooling, heating, and energy consump­
tion of the proposed integrated system are kept fixed with that of the According to Fig. 1, the considered CCHP system includes ORC, ARC,
Configuration1, then, the amount of power production of these systems and Heat exchanger (HE); its energy is provided by a stream which exits
are compared. the heat source with a mass flow rate of 5 kg/s and a temperature of
120 ◦ C. It should be noted that the mentioned stream is named as heat
source stream (HSS) in the following sections. The energy consumption
rate to generate an HSS with a temperature of 120 ◦ C is assumed as 1890
kW. The Configuration1 has a sequential configuration. Therefore,

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S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Fig. 2. Schematic of the proposed system.

providing the required energy of ORC, the HSS passes through the ORC ORC evaporator to provide the energy of this cycle. On the other hand,
evaporator. Next, it enters the ARC generator, and then, the HEheat to the produced hot water in the AHT condenser is used directly for heating
supply the necessary energy for producing cooling and heating, purposes as well as the produced hot water in HEheat.
respectively. The Configuration1 is designed to produce approximately The exit stream at point 6 mixes with the exit stream from the
102.5 kW of power, 200 kW of cooling, and 776 kW of heating (domestic condenser of ORC (which has been preheated in the condenser of ORC).
hot water at the temperature of 60 ◦ C). Increasing the temperature of the stream at point 33, more mass flow
rate reaches the temperature of 120 ◦ C, while the same amount of energy
as that of the Configuration1 is utilized in the heat source. It should be
2.6. Description of the proposed system noted that in the aforementioned modes, to have the same amount of
productions as those of the Configuration1, the mass flow rate of streams
The schematic of the proposed system is presented in Fig. 2. As can at points 4 and 5 is changed.
be seen, not only an AHT cycle as an auxiliary system is integrated with
the CCHP system, but also, some changes are performed in its configu­ 2.6.1. Absorption heat transformer (AHT)
ration. The HSS is divided into three parts. The entering streams to This cycle is recognized as a heat transformer or a temperature
points 8 and 9 pass through the evaporator and generator of AHT, booster. The purpose of this cycle is to provide higher-temperature heat
respectively. Therefore, the required energy of the AHT cycle is provided by utilizing intermediate-temperature heat. The energy of the stream
to upgrade the temperature of the passing stream through the absorber. entering the evaporator and generator of this cycle is utilized to provide
After releasing energy in the generator and evaporator, the exited a higher-temperature stream at the absorber. A schematic of the pro­
streams are mixed at point 12 and passed via the ARC generator and then posed AHT in this work is presented in Fig. 3; lithium bromide (LiBr) and
HEheat to provide the required energy for producing cooling and heating. water are used as absorbent and refrigerant, respectively. LiBr_H2O
This is while the upgraded stream leaves the absorber and enters the

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S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Fig. 3. Schematic of the absorption heat transformer.

Fig. 4. Schematic of an ORC cycle using regenerator.

solution is heated in the generator by receiving heat from a hot stream, condenser to be changed into saturated liquid by releasing heat. After its
causing absorbent and refrigerant are being separated. The refrigerant pressure is increased by the pumpw, it enters the evaporator and its
leaves the generator as saturation vapor and further enters the phase is changed to saturated vapor by receiving heat from the hot

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S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Fig. 5. Schematic of an absorption refrigeration system.

stream during the constant pressure process. The vapor further enters produces cooling. Water and ammonia are used as the absorbent and
the absorber, which consists of adiabatic and cooled sections. Part of it refrigerant, respectively. Ammonia-water solution is heated in the
enters the adiabatic section and it is absorbed by the LiBr_H2O solution. generator by receiving heat from the hot stream, leading to the sepa­
In this way, its latent heat is released, and thus the temperature of the ration of the absorbent and refrigerant. The refrigerant at its saturation
solution is raised. Next, the other part of the vapor is absorbed by this vapor leaves the generator, and then enters the condenser to be changed
solution, and its latent heat is released, and the temperature of the so­ into saturated liquid by releasing its heat. After its pressure is decreased
lution is raised in the cooled section. Consequently, the released heat is by the expansion valve, it enters the evaporator and its phase is changed
absorbed by another entering stream to the absorber, and thus its tem­ to the saturated vapor by receiving some heat from the cold source
perature is increased. Furthermore, LiBr_H2O solution leaves the during the constant pressure process. It enters absorber, where it is
generator and its pressure is increased by the solution pump. Next, it absorbed by ammonia-water solution, leading to its latent heat release.
passes through the heat exchanger (HE) and its temperature is Ammonia-water solution leaves the generator, then it passes through the
decreased. LiBr_H2O solution leaving the absorber, passes through the heat exchanger where its temperature is decreased. Next, it enters the
HE, therefore its temperature is increased. Its pressure is decreased by expansion valve to obtain low pressure values. Ammonia-water solution
the solution expansion valve while returning to the generator. leaves the absorber and passes through the pump, so its pressure is
increased again. Further, it enters the HE and its temperature is
2.6.2. Organic Rankine cycle increased.
When organic fluids with low boiling points are used, ORCs are
suitable to produce power by utilizing low-temperature waste heat. In 3. Modeling and analysis
the ORC system considered in the current work, firstly, R1234ze as the
working fluid (with GWP < 1) at its saturated liquid phase enters the Investigating and comparing the proposed system, energy, exergy,
pump. Once increased its pressure, it passes through the internal heat and economic analyses are carried out. In this way, all of the proposed
exchanger (IHE) and it is further preheated by the exited stream from the modes are modeled using the Engineering Equation Solver (EES) soft­
turbine. Further, it enters the evaporator, where it receives some heat ware. To simplify the thermodynamic analysis the following assump­
from the hot stream, and exits as a superheated vapor and flows into the tions are considered:
turbine to produce power. Finally, after exiting the turbine and passing
through the IHE, the working fluid enters the condenser to release its • The system operates in a steady-state condition.
heat under a constant pressure process. The schematic of the considered • Pressure drop and heat losses in pipes and heat exchangers are
ORC in this work is shown in Fig. 4. ignored [22,23]
• The reference temperature and pressure are considered as 25 ◦ C and
2.6.3. Absorption refrigeration cycle 101.3 kPa, respectively.
A schematic of an absorption refrigeration system is demonstrated in • Isentropic efficiency of pumps and the turbine of the ORC are
Fig. 5. Absorption refrigeration is a refrigerating process that uses the considered as 85% and 80% [24], respectively.
energy of a heat source to provide cooling. In this work, the heat source • The temperature of the ARC evaporator is defined as − 10 ◦ C.
is the hot stream passing through the generator, and thus the evaporator • The produced heating is hot water at a temperature of 60 ◦ C.

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S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Table 1
Energy balances of the proposed system and Configuration1 considering Figs. 1 and 2.
Energy balance of the proposed system Energy balance of the Configuration1

ORC
evap Q̇eva = ṁ5 (h5 − h6 ) = ṁ28 (h29 − h28 ) Q̇eva = ṁ1 (h1 − h2 ) = ṁ8 (h9 − h8 )
Turbine Ẇturbine = ṁ29 (h29 − h30 ) Ẇturbine = ṁ9 (h9 − h10 )
IHX Q̇IHX = ṁ30 (h30 − h31 ) = ṁ27 (h28 − h27 ) Q̇IHX = ṁ10 (h10 − h11 ) = ṁ7 (h8 − h7 )
Cond Q̇cond = ṁ31 (h31 − h26 ) = ṁ32 (h32′ − h32 ) Q̇cond = ṁ11 (h11 − h6 )
Pump ẆPump,R = ṁ26 (h27 − h26 ) ẆPump,R = ṁ6 (h7 − h6 )

ARC
Abs Q̇abs = ṁ40 h40 + ṁ48 h48 − ṁ35 h35 Q̇abs = ṁ17 h17 + ṁ25 h25 − ṁ12 h12
Pumpsol ẆPump,sol = ṁ35 (h36 − h35 ) ẆPump,sol = ṁ12 (h13 − h12 )
HEsol Q̇HE,sol = ṁ38 (h38 − h39 ) = ṁ36 (h37 − h36 ) Q̇HE,sol = ṁ15 (h15 − h16 ) = ṁ13 (h14 − h13 )
Gen Q̇gen = ṁ37 h37 + ṁ42 h42 − ṁ38 h38 − ṁ41 h41 = ṁ13 (h13 − h14 ) Q̇gen = ṁ14 h14 + ṁ19 h19 − ṁ15 h15 − ṁ18 h18 = ṁ2 (h2 − h3 )
Cond Q̇cond = ṁ43 (h43 − h44 ) Q̇cond = ṁ20 (h20 − h21 )
HEw Q̇HE,w = ṁ44 (h44 − h45 ) = ṁ47 (h47 − h48 ) Q̇HE,w = ṁ21 (h21 − h22 ) = ṁ24 (h25 − h24 )
Eva Q̇eva = ṁ46 (h47 − h46 ) Q̇eva = ṁ23 (h24 − h23 )

AHT
Abs Q̇abs = ṁ2 (h3 − h3 ) = ṁ18 h18 + ṁ25 h25 − ṁ19 h19 –
HEsol Q̇HE,sol = ṁ19 (h19 − h20 ) = ṁ17 (h18 − h17 ) –
Pumpsol ẆPump,sol = ṁ16 (h17 − h16 ) –
Gen Q̇gen = ṁ9 (h9 − h11 ) = ṁ21 h21 − ṁ16 h16 − ṁ22 h22 –
Cond Q̇cond = ṁ22 (h22 − h23 )
Pumpw ẆPump,w = ṁ23 (h24 − h23 ) –
Eva Q̇eva = ṁ25 (h25 − h10 ) = ṁ24 (h25 − h24 ) –

Other –
HEheat Q̇Heat = ṁ14 (h14 − h15 ) Q̇Heat = ṁ3 (h3 − h4 )
Pumpcycle1 ẆPump,cycle1 = ṁ1 (h2 − h1 ) ẆPump,cycle = ṁ4 (h5 − h4 )
Pumpcycle2 ẆPump,cycle2 = ṁ33 (h34 − h33 ) –

3.1. Energy and exergy analysis In the case of the flows without chemical reaction, the chemical
exergy (E˙xch) is taken as zero. In other cases, it is calculated from the
The energy and exergy balances for each component are presented following equation [26]:
according to the first and second laws of thermodynamics, including the [ ]
Xcm 0 1 − Xcm 0
mass conservation in a steady-state condition. Ėxch = ṁ exch,cm + exch,H2 O (5)
Mcm Mcm
In control surface for an open system, work transfer (W˙), heat
transfer (Q˙) and energy associated with the mass transfer are considered
where subscript cm denotes the chemical mixture and it can be LiBr or
as three types of energy transfer [25]:
NH3. ex0ch is the standard chemical exergy and its value for different
∑ ∑
N ∑
N chemical mixtures can be obtained from the reference [26]. X and M are
Q̇k + (ṁi hi )k = Ẇ k + (ṁe he )k (1) the concentration and molecular mass, respectively. The exergy balance
at the steady-state condition is defined as
k i e

where m˙is mass flow rate and h is enthalpy. Subscripts k, i and e indicate ∑N (
T0
) ∑
N ∑
N

kth component, inlet, and outlet, respectively. The energy efficiency of 1− Q̇k + Ėxi,k = Ėxw,k + Ėxe,k + ĖxD,k (6)
T
the system is obtained as i i e

Useful Output Energy where T0 is the ambient temperature and subscripts w, i, e, and D
ηen,sys = (2)
Input Energy indicate the exergy of work, input, output, and destruction, respectively.
Ignoring the rate of exergy losses, the exergy efficiency of each
where the input energy is the HSS, which is constant for all modes except component, εk, is obtained as
mode D, and the useful output energy is the sum of the produced power, ( ) ( ) ( )
cooling and heating energy, which may vary depending on a mode. The εk = 100
ĖxP,k
= 100
ĖxF,k − ĖxD,k
= 100 1 −
ĖxD,k
[%] (7)
maximum useful work that is achievable during the process is named as ĖxF,k ĖxF,k ĖxF,k
the exergy of the system. Ignoring the kinetic and potential exergies, the
exergy of the stream in the steady-state condition, Ėx, is divided into where indices such as P, F, and D indicate the respective product, fuel,
physical and chemical exergies: and destruction. Index k also denotes kth component of the system. To
calculate the exergy destruction ratio for each component, YD,k, the
Ėx = Ėxph + Ėxch (3) following equation is used

where Ėxph is the physical exergy and it is defined as YD,k = 100 × ∑

ĖxD,k
[%] (8)
k ĖxD,k
Ėxph = ṁ[(h − h0 ) − T0 (s − s0 )] (4)
The exergy efficiency of the system is calculated as

8
S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

∑
(Ẇ turbine − i Ẇ Pump i ) + Q̇Heat (1 − T0 /THeat ) + Q̇eva,ARC (T0 /Teva.ARC − 1)
ηex,sys = 100 × ∑ [%] (9)
j Q̇j (1 − T0 /Tj )

of the system components

where i shows the number of the existed pumps, j indicates the
i(i + 1)n
component, to which hot stream heat is transferred. CRF = (12)
(i + 1)n − 1
According to the above-mentioned equations, the energy and exergy
balances for each component of the ORC, AHT, and ARC systems are
where i is the interest rate (10% [27]), and n is the lifetime of the system
presented in Tables 1 and 2.
(15 years [27]). PW is calculated by the following equation:
PW = TCI − SV(PWF) [US$] (13)
3.2. Economic analysis
where PWF and SV are respectively the present worth factor and the
As another important criterion, an economic analysis is carried out.
In this simple analysis payback period, net present value, cost of pro­
Table 3
duced power, heating and cooling, etc. are calculated. The sum of cost Total investment cost of different components [28,29].
rates associated with the capital investment (ŻCI) and operating and
Component Total Investment Cost (US$)
maintenance (ŻOM) is determined as )0.7 ( )3 )(
Turbine ( (
0.05
TCIturbine = 6000 Ẇturbine 1+ 1 + 5exp
Ω(AC) 1 − ηturbine
Ż = Ż CI + Ż OM = [US$/h] (10) ( ))
τ Tin − 866
10.42 ( )
where Ω is the factor including total operating and maintenance cost (Ω Heat AHE 0.78
exchangers, TCIHE = 130
= 1.06 [27]), AC is the annual capital cost and τ is the working hour per 0.093
evaporators,
year: and
generators
AC = PW × CRF [US$] (11) ( )0.71
Pump
TCIpump = 3540 Ẇpump
where CRF and PW are the capital recovery factor and the present value

Table 2
Exergy balances of the proposed system and Configuration1 considering Figs. 1 and 2.
Exergy balance of the proposed system Exergy balance of the Configuration1

ORC
evap Ėx5 + Ėx28 = Ėx6 + Ėx29 + ĖxD,eva Ėx1 + Ėx8 = Ėx2 + Ėx9 + ĖxD,eva
Turbine Ėx29 = Ėx30 + Ẇturbine + ĖxD,turbine Ėx9 = Ėx10 + Ẇturbine + ĖxD,turbine
IHX Ėx30 + Ėx27 = Ėx31 + Ėx28 + ĖxD,IHX Ėx10 + Ėx7 = Ėx11 + Ėx8 + ĖxD,IHX
Cond Ėx31 = Ėx26 + Q̇cond (1 − T0 /Tcond ) + ĖxD,cond Ėx11 = Ėx6 + Q̇cond (1 − T0 /Tcond ) + ĖxD,cond
Pump Ėx26 + Ẇpump,R = Ėx27 + ĖxD,pump,R Ėx6 + Ẇpump,R = Ėx7 + ĖxD,pump,R

ARC
Abs Ėx40 + Ėx48 = Ėx35 + Q̇abs (1 − T0 /Tabs ) + ĖxD,abs Ėx17 + Ėx25 = Ėx12 + Q̇abs (1 − T0 /Tabs ) + ĖxD,abs
Pumpsol Ėx35 + ẆPump,sol = Ėx36 + ĖxD,Pump,sol Ėx12 + ẆPump,sol = Ėx13 + ĖxD,Pump,sol
HEsol Ėx36 + Ėx38 = Ėx37 + Ėx39 + ĖxD,HE,sol Ėx13 + Ėx15 = Ėx14 + Ėx16 + ĖxD,HE,sol
Gen Ėx37 + Ėx42 + Ėx13 = Ėx38 + Ėx41 + Ėx14 + ĖxD,gen Ėx14 + Ėx19 + Ėx2 = Ėx15 + Ėx18 + Ėx3 + ĖxD,gen
Cond Ėx43 = Ėx44 + Q̇cond (1 − T0 /Tcond ) + ĖxD,cond Ėx20 = Ėx21 + Q̇cond (1 − T0 /Tcond ) + ĖxD,cond
HEw Ėx44 + Ėx47 = Ėx45 + Ėx48 + ĖxD,HE,w Ėx21 + Ėx24 = Ėx22 + Ėx25 + ĖxD,HE,w
Eva Ėx46 + Q̇eva (T0 /Teva − 1) = Ėx47 + ĖxD,eva Ėx23 + Q̇eva (T0 /Teva − 1) = Ėx24 + ĖxD,eva

AHT
Abs Ėx2 + Ėx18 + Ėx25 = Ėx3 + Ėx19 + ĖxD,abs –
HEsol Ėx19 + Ėx17 = Ėx18 + Ėx20 + ĖxD,HE,sol –
Pumpsol Ėx16 + ẆPump,sol = Ėx17 + ĖxD,Pump,sol –
Gen Ėx21 + Ėx9 = Ėx16 + Ėx22 + Ėx11 + ĖxD,gen –
Cond Ėx22 = Ėx23 + Q̇cond (1 − T0 /Tcond ) + ĖxD,cond –
Pumpw Ėx23 + ẆPump,w = Ėx24 + ĖxD,Pump,w –
Eva Ėx24 + Ėx8 = Ėx25 + Ėx10 + ĖxD,eva –

Other –
HEheat Ėx14 = Ėx15 + Q̇Heat (1 − T0 /THeat ) + ĖxD,HE,Heat Ėx3 = Ėx4 + Q̇Heat (1 − T0 /THeat ) + ĖxD,HE,Heat
Pumpcycle1 Ėx1 + ẆPump,cycle1 = Ėx2 + ĖxD,Pump,cycle1 Ėx4 + ẆPump,cycle = Ėx5 + ĖxD,Pump,cycle
Pumpcycle2 Ėx33 + ẆPump,cycle2 = Ėx34 + ĖxD,Pump,cycle2 –

9
S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Table 5 presented in Appendix A. It should be noted that the pinch point of ORC
Input data to validate AHT, ARC, and ORC due to the points of Fig. 2. evaporators of the proposed system at modes A-D and Configuration1 are
AHT [31] ARC [32] ORC [33] obtained as 4.39 and 3.38, respectively.
The total depreciable investment (TDI) is the difference between the
m˙16 = 0.194 kg/s T47 = -15 C◦
T26 = 14.5 ◦ C
m˙9 = 1.389 kg/s Q47 = 0.998 ηs,turbine = 58.53% capital investment cost and the salvage value:
m˙8 = 1.389 kg/s Q44 = 0.04 P27 = 640.9 kPa
TDI = TCI − SV [US$] (18)
m˙2 = 1.389 kg/s Q41 = 0.998 Q˙eva = 76.06 kW
T9 = 95 ◦ C ηHE,sol = 80% m˙26 = 0.403 kg/s The annual book depreciation (BD) is calculated as
T8 = 95 ◦ C Q35 = 0.04
T2 = 130 ◦ C TDI
BD = [US$/year] (19)
economic life (n)
salvage value, which is the estimated resale value of an asset at the end The average annual profit AP is the difference between annual net
of its useful life, and are determined as cash flow ANCF and annual depreciation BD:
1 AP = ANCF − BD [US$/year] (20)
PWF = (14)
(1 + i)
The average rate of return ARR on the initial investment is the ratio
SV = μ(TCI) [US$] (15) of the average annual profit to the sum of the total initial investment and
working capital (WC is 15% of TCI [30]):
where μ is the salvage percentage (15% [27]), and TCI is the total capital AP
investment. Below, the TCI (US$) of components used in the proposed ARR = 100 × [%] (21)
TCI + WC
system is presented. To update the TCI values to the original year, the
following equation is used: The payback period (PP) is defined as the length of the time required
for the cash inflows received from a project to recover the original cash
Original cost = cost at reference year ×
cost index for the original year outlays required by the initial investment. Mathematically, the payback
cost index for the reference year period PP is defined by the following relation:
(16)
TDI
The TCI (US$) of all components is presented in table 3. The heat PP = [year] (21)
ANCF
transfer area of heat exchangers is calculated according to the procedure
Net present value (NPV) is the difference between the present value

Table 6
Validation results of mathematical modeling for each cycle.
AHT ORC ARC

Parameter P.W. Ref. [31] Deviation (%) Parameter P.W. Ref. [33] Deviation (%) Parameter P.W. Ref. [32] Deviation (%)

Q˙eva (kW) 34.12 34 0.35 T29 (kW) 118 119.4 1.17 Q˙eva (kW) 2.5 2.8 0.07
COPAHT 0.49 0.47 4.26 W˙turbine (kW) 6.21 6.07 2.31 Q˙gen (kW) 6 6 0.00
= Q˙IHE (kW) 22.86 22.58 1.24 Q˙abs (kW) 5 4.5 0.69
ηORC (%) 7.95 7.98 0.38 COPARC 0.423 0.424 0.67

Table 7
Thermodynamic properties of the proposed system working at mode A.
Point P (kPa) T (◦ C) h (kJ/kg) s (kJ/kgK) m˙ (kg/s) X (kg/kg) Point P (kPa) T (◦ C) h (kJ/kg) s (kJ/kg K) m˙ (kg/s) X (kg/kg)

1 200.00 120.00 503.80 1.53 2.62 – 25 58.27 85.20 2652.0 7.54 0.10 –
2 380.00 120.00 503.90 1.53 2.62 – 26 429.1 20 227.3 1.096 4.44 –
3 380.00 141.80 602.80 1.77 2.62 – 27 3632 21.9 230.5 1.097 4.44 –
4 0.00 0.00 0.00 0.00 0.00 – 28 3632 30 241.6 1.135 4.44 –
5 380.00 141.80 602.80 1.77 2.62 – 29 3632 128.8 467.7 1.761 4.44 –
6 13.96 52.50 219.80 0.74 2.62 – 30 429.1 55.34 432.3 1.788 4.44 –
7 200.00 120.00 503.80 1.53 3.46 – 31 429.1 43.93 421.2 1.754 4.44 –
8 200.00 120.00 503.80 1.53 1.77 – 32 13.96 10.00 41.99 0.151 6.49 –
9 200.00 120.00 503.80 1.53 1.69 – 33 13.96 41.72 174.9 0.595 3.46 –
10 200.00 85.07 356.30 1.14 1.77 – 34 200.00 46.10 193.0 0.653 6.08 –
11 200.00 84.24 352.80 1.13 1.69 – 35 286.79 13.00 − 175.1 0.03 0.85 57.53
12 200.00 84.66 354.60 1.13 3.46 – 36 681.12 13.02 − 174.6 0.03 0.85 57.53
13 200.00 84.66 354.60 1.13 3.46 – 37 681.12 39.12 − 55.64 0.43 0.85 57.53
14 200.00 66.08 276.80 0.91 3.46 – 38 681.12 54.05 3.02 0.64 0.69 47.53
15 4.25 30.00 125.70 0.44 3.46 – 39 681.12 21.23 − 144.0 0.16 0.69 47.53
16 6.86 87.42 259.70 0.48 1.15 62.17 40 286.79 21.31 − 144.0 0.17 0.69 47.53
17 58.27 110.50 259.70 0.60 1.15 62.17 41 681.12 39.12 1351.9 4.71 0.16 –
18 58.27 142.40 319.40 0.75 1.15 62.17 42 681.12 39.11 − 55.78 0.43 0.00 57.53
19 58.27 140.70 339.70 0.68 1.05 68.20 43 681.12 21.07 1288.1 4.50 0.16 –
20 58.27 102.50 274.10 0.51 1.05 68.20 44 681.12 13.00 60.24 0.23 0.16 –
21 6.86 102.50 274.10 0.51 1.05 68.20 45 681.12 − 2.23 − 10.31 − 0.03 0.16 –
22 6.86 85.00 2659.0 8.54 0.10 – 46 286.79 − 10.33 − 10.31 − 0.02 0.16 –
23 6.86 38.64 161.80 0.55 0.10 – 47 286.79 − 10.00 1224.3 4.65 0.16 –
24 58.27 38.64 161.90 0.55 0.10 – 48 286.79 5.59 1294.8 4.92 0.16 –

Points 16–21, X denotes XLiBr.

Points 35–42, X denotes XNH3.

10
S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Table 8
Thermodynamic properties of the Configuration1.
Point P (kPa) T (◦ C) h (kJ/kg) s (kJ/kgK) m˙ (kg/s) X (kg/kg) Point P (kPa) T (◦ C) h (kJ/kg) s (kJ/kg K) m˙ (kg/s) X (kg/kg)

1 200.0 120.0 503.80 1.532 5.00 – 14 681.1 39.1 − 55.6 0.430 0.85 57.53
2 200.0 80.0 335.00 1.075 5.00 – 15 681.1 54.1 3.0 0.638 0.69 47.53
3 200.0 67.1 281.20 0.920 5.00 – 16 681.1 21.2 − 144.0 0.164 0.69 47.53
4 4.3 30.0 125.70 0.440 5.00 – 17 286.8 21.3 − 144.0 0.166 0.69 47.53
5 200.0 30.0 125.80 0.441 5.00 – 18 681.1 39.1 1351.9 4.711 0.16 99.77
6 429.10 20.00 227.30 1.10 4.23 – 19 681.1 39.1 − 55.8 0.429 0.00 57.53
7 3000.0 21.53 229.80 1.10 4.23 – 20 681.1 21.1 1288.1 4.498 0.16 –
8 3000.0 30.00 241.60 1.14 4.23 – 21 681.1 13.0 60.2 0.227 0.16 –
9 3000.0 106.80 441.10 1.70 4.23 – 22 681.1 − 2.2 − 10.3 − 0.026 0.16 –
10 429.10 34.72 412.20 1.73 4.23 – 23 286.8 − 10.3 − 10.3 − 0.022 0.16 –
11 429.10 22.65 400.50 1.69 4.23 – 24 286.8 − 10.0 1224.3 4.653 0.16 –
12 286.8 13.0 − 175.1 0.032 0.85 57.53 25 286.8 5.6 1294.8 4.916 0.16 –
13 681.1 13.0 − 174.6 0.032 0.85 57.53

Points 12–19, X denotes XNH3.

Table 9 Table 10
The main results of energy analysis for various modes. The main results of exergy analysis for ORC components of different modes.
Parameter Mode A Mode B Mode C Mode D Configuration1 Configuration Component ĖxF ĖxP ĖxD YD εk
(kW) (kW) (kW) (%) (%)
Q˙in,source (kW) 1890.00 1890.00 1890.00 1606.05 1890.00
Q˙eva,AHT (kW) 253.80 253.80 253.80 234.10 – Proposed system PumpORC 14.13 11.99 2.14 1.32 84.85
Q˙gen,AHT (kW) 259.70 259.70 259.70 238.10 – (mode A) IHE 3.81 0.18 3.63 2.24 4.72
Q˙abs,AHT (kW) 259.00 259.00 259.00 237.50 – EvaORC 259.20 174.60 84.60 52.12 67.36
Q˙cond,AHT (kW) 254.60 254.60 254.60 234.70 – Turbine 192.90 156.80 36.10 22.24 81.29
Q˙eva,ORC (kW) 1004.0 789.80 784.40 784.10 843.80 CondORC 45.74 9.88 35.86 22.09 21.60
W˙turbine (kW) 156.80 123.40 122.6 122.50 122.20
Proposed system PumpORC 11.12 9.44 1.68 1.32 84.89
Q˙cond,ORC (kW) 860.80 677.50 672.90 672.60 732.40
(mode B) IHE 3.00 0.14 2.86 2.24 4.67
Q˙gen,ARC (kW) 269.40 544.50 269.40 269.40 269.40
EvaORC 204.00 137.50 66.50 52.09 67.40
Q˙eva,ARC (kW) 200.00 404.30 200.00 200.00 200.00
Turbine 151.80 123.40 28.40 22.25 81.29
W˙tot,pump (kW) 20.16 16.92 16.44 16.05 15.77
CondORC 36.00 7.78 28.22 22.11 21.61
W˙net (kW) 140.57 109.84 109.84 109.84 109.84
Q˙Heat (kW) 775.92 775.92 1061.99 775.92 775.92 Proposed system PumpORC 11.05 9.37 1.68 1.32 84.80
COPAHT 0.50 0.50 0.50 0.50 – (mode C) IHE 2.98 0.14 2.84 2.24 4.70
ηORC (%) 14.22 14.22 14.22 14.22 13.20 EvaORC 202.60 136.50 66.10 52.11 67.37
COPARC 0.74 0.74 0.74 0.74 0.74 Turbine 150.80 122.60 28.20 22.23 81.30
ηen,sys (%) 59.07 68.26 72.58 67.60 57.44 CondORC 35.75 7.73 28.02 22.09 21.62

Proposed system PumpORC 11.04 9.37 1.67 1.32 84.87

(mode D) IHE 2.98 0.14 2.84 2.24 4.70
of cash inflows and the present value of cash outflows over a period. if
EvaORC 202.60 136.50 66.10 52.12 67.37
the NPV is positive, the proposed system is acceptable, otherwise, it is Turbine 150.70 122.50 28.20 22.23 81.29
rejected. CondORC 35.74 7.72 28.02 22.09 21.60

∑
n
Yz Configuration1 PumpORC 10.81 9.17 1.64 1.48 84.83
NPV = [US$] (23) IHE 0.62 0.15 0.47 0.42 24.19
(1 + i)z
z=0 EvaORC 178.60 131.60 47.00 42.37 73.68
Turbine 152.10 122.20 29.90 26.95 80.34
where Yz is the net cash flow at the end of zth year. CondORC 18.52 11.79 6.73 6.07 63.66
To calculate ANCF, the cost of the produced power, Celect., the
heating (hot water) at 60 ◦ C, CHeat, and the cooling, Ccool, are calculated
4. Validation
as

Ż tot In this section, the mathematical modeling of the ORC, AHT, and
Celect. = [US$/kWh] (24)
Ẇ net ARC are validated. In this way, the results of the mentioned cycles are
compared with similar experimental works. To solve the mathematical
Ż tot modeling of each cycle in this work, the input data is presented in
CHeat = [US$/kWh] (25)
Q̇heat Table 5 where Q denotes the quality of a stream. Accordingly, the
calculated results of each cycle and also the reference results with their
Ż tot relative deviation are reported in Table 6. The maximum deviation of
Ccool = [US$/kWh] (26) 4.26% shows that the results obtained by the mentioned thermodynamic
Q̇eva,ARC
models in this work have a good agreement with the results of the
∑ considered references.
where Żtot is the total of Ż of all components (Żtot = Żk), and V˙ is the
volumetric flow rate of the produced hot water.
To calculate ANCFtot in a year, the ANCFs from the selling of the 5. Results and discussion
produced power, hot water and cooling should be calculated. Due to the
operating hours of the system, 8000 h, ANCFtot can be determined as: 5.1. Energy and exergy analysis
[ ]
ANCFtot = τ Celect. Ẇ net + Chotwater V̇ 30 + Ccooling Q̇eva.2 [US$/year] (23) According to the aforementioned assumptions, the thermodynamic
properties of all points in the Configuration1 and the proposed system
are determined. Tables 7 and 8 show the thermodynamic properties of

11
S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Table 11 most exergy destruction belongs to the evaporator followed by the tur­
The main results of exergy analysis for ARC components of different modes. bine and condenser. The evaporator of ORC in mode A has the maximum
Configuration Component ĖxF ĖxP ĖxD YD εk exergy destruction where the purpose is to increase the net power pro­
(kW) (kW) (kW) (%) (%) duction, then, the most energy passes through the evaporator. Also as it
Proposed system GenARC 44.93 19.19 25.74 59.56 42.71 is expected, the most exergy destruction in ARC happens in the gener­
(mode A) CondARC 14.31 7.28 7.03 16.27 50.87 ator and the most one belongs to mode B. Regarding the AHT cycle, it is
HEw,ARC 1.26 0.80 0.46 1.06 63.49 seen that the generators have the maximum exergy destruction between
EvaARC 26.62 25.70 0.92 2.13 96.54 the mentioned components. It should be noted that in the aforemen­
Pumpsol, 0.42 0.42 0.01 0.01 98.81
tioned components, the exergy destructions are due to the high-
ARC
HEsol,ARC 4.08 0.36 3.72 8.61 8.82 temperature difference between the hot and cold streams.
AbsARC 9.01 3.67 5.34 12.36 40.73 The exergy efficiencies of the proposed system at modes A-D and the
Proposed system GenARC 101.00 38.80 62.20 63.75 38.42 Configuration1 are obtained as 47.72%, 53.89%, 54.30%, 56.15%, and
(mode B) CondARC 28.93 14.70 14.23 14.58 50.81 47.72%, respectively. The exergy efficiency of the proposed system is
HEw,ARC 2.54 1.61 0.93 0.95 63.39 increased because of the reduction of the waste energy in the condenser
EvaARC 53.80 51.94 1.86 1.91 96.54 of ORC, in addition to the reduction of the mass flow rate and temper­
Pumpsol, 0.85 0.82 0.03 0.03 96.47
ature of the stream entering the HEheat.
ARC
HEsol,ARC 8.25 0.73 7.52 7.71 8.85 Comparing the proposed system with the Configuration1, the relative
AbsARC 18.21 7.41 10.80 11.07 40.69 difference of energy and exergy efficiencies are presented in Fig. 6. The
Proposed system GenARC 50.08 19.19 30.89 63.86 38.32
relative difference indicates the ratio of the difference between the en­
(mode C) CondARC 14.31 7.28 7.03 14.54 50.85 ergy (or exergy) efficiency of the proposed system and the Configura­
HEw,ARC 1.26 0.80 0.46 0.95 63.49 tion1 to the energy (or exergy) efficiency of the Configuration1. It is seen
EvaARC 26.62 25.70 0.92 1.90 96.54 that in the proposed system, the energy and exergy efficiencies are
Pumpsol, 0.42 0.42 0.00 0.01 98.86
increased considerably compared with Configuration1.
ARC
HEsol,ARC 4.08 0.36 3.72 7.69 8.81 At the end of this section, it is concluded that according to the pri­
AbsARC 9.01 3.67 5.35 11.05 40.69 orities of users, each mode is suitable for special conditions and de­
Proposed system GenARC 46.53 19.19 27.34 60.99 41.24
mands; however, the most important point is the improvement of output
(mode D) CondARC 14.31 7.28 7.03 15.69 50.85 productions and efficiencies of the proposed system in all modes.
HEw,ARC 1.26 0.80 0.46 1.03 63.49
EvaARC 26.62 25.70 0.92 2.05 96.54 5.2. Economic analysis
Pumpsol, 0.42 0.42 0.00 0.01 98.86
ARC
HEsol,ARC 4.08 0.36 3.72 8.30 8.81 Until now, it is recognized that integrating an AHT with the CCHP
AbsARC 9.01 3.67 5.35 11.92 40.69 system develops the performance of the system considerably from en­
Configuration1 GenARC 41.97 19.19 22.78 65.23 45.72
ergy, exergy, and environmental standpoints. However, this integration
CondARC 14.31 7.28 7.03 20.13 50.87 and improvement have additional investment costs. Therefore, the
HEw,ARC 1.26 0.80 0.46 1.32 63.49 proposed system would be better than the Configuration1 whenever it
EvaARC 26.62 25.70 0.92 2.63 96.54 also has cost advantages, in addition to the environmental, energy and
Pumpsol, 0.42 0.42 0.01 0.01 98.81
exergy advantages.
ARC
HEsol,ARC 4.08 0.36 3.72 10.65 8.82 The main results of economic analysis in various modes are reported
AbsARC 9.01 3.67 5.34 15.29 40.73 in Table 13. As it is seen, the investment cost of the AHT cycle is almost
2.3–3.5 times lower than that of the Configuration1; therefore, its in­
vestment is not remarkable compared with the Configuration1. How­
the mentioned points in the proposed system working at mode A and the ever, as mentioned earlier, in the proposed system, the mass flow rate of
Configuration1, respectively. The thermodynamic properties of the the HSS is increased. Then, depending on the design of the system due to
proposed system working at other modes are also reported in Appendix the mentioned modes, the size of components changes, and these
B. changes increase the investment cost. Between the various modes, mode
Having the thermodynamic properties of the proposed system in B has the highest investment cost, on the other hand, mode D has the
different modes, the main results of the energy analysis such as required lowest investment cost. Nevertheless, considering NPV and ANCFtot as
input energies, desired outputs, the amount of the produced power, indicators of comparison with the Configuration1, it is clear that the
cooling and heating are presented in Table 9. The production of power at proposed system working at different modes has higher NPV and
mode A, cooling at mode B, and heating at mode C of the proposed ANCFtot. However, between the working modes, mode B has the
system are obtained as 140.57 kW, 404.30 kW, and 1061.99 kW, maximum NPV and ANCFtot, which are developed by 112.40% and
respectively. Therefore, comparing with the Configuration1, the pro­ 112.67%, respectively.
duction of power (mode A), cooling (mode B) and heating (mode C) are Having the main results of economic analysis and also carbon credit,
developed 27.98%, 102.15%, and 36.87%, respectively. Also, it is it is possible to calculate the cost saving if the proposed system is used
indicated that the required energy in heat source is decreased from 1890 instead of the Configuration1.
kW (Configuration1) to 1606.05 kW (mode D) which means during a
year (8000 h) the energy consumption is decreased 2271.6 MWh; and 5.2.1. Carbon credit and cost saving
consequently, it reduces the carbon emission. Considering natural gas As mentioned in previous sections, it is possible to produce the same
(NG), Biomass (BM), and solar energy as heat sources, the carbon amount of heating, cooling, and power as the same as those of the
emissions are reduced by 1056.3 ton/year, 97.7 ton/year, and 68.1 ton/ Configuration1 but by less input energy if an AHT cycle is integrated. In
year, respectively. It should be noted that the carbon emission for NG, other words, the less energy is required, the less fuel and equipment size
BM and solar is reported as 465 g/kWh, 43 g/kWh, and 30 g/kWh [34], are needed in heat source which reduces carbon emission as well as the
respectively. investment cost rate (including investment, operation and maintenance
Another useful tool to investigate the proposed system is the exergy costs). In this section, considering carbon credit, the system investment
analysis. Tables 10–12 show the main results of the exergy analysis for cost rate (ICR), and the cost of energy in the heat source, the saving cost
each component of the proposed system in different modes. In ORC, the during a year is determined for the proposed system. The saving cost,

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S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Table 12
The main results of exergy analysis for AHT components of different modes.
Configuration Component ĖxF (kW) ĖxP (kW) ĖxD (kW) YD (%) εk (%)
Proposed system (mode A) AbsAHT 74.66 72.93 1.73 1.80 97.68
HEsol,AHT 16.27 15.65 0.62 0.64 96.19
Pumpsol,AHT 0.03 0.01 0.02 0.02 20.67
GenAHT 62.78 3.78 59.00 61.33 6.02
CondAHT 26.88 11.84 15.04 15.63 44.05
Pumpw,AHT 0.0062 0.0053 0.0009 0.0009 85.48
EvaAHT 61.36 41.57 19.79 20.57 67.75

Proposed system (mode B) AbsAHT 74.66 72.93 1.73 1.80 97.68

HEsol,AHT 16.27 15.65 0.62 0.64 96.19
Pumpsol,AHT 0.03 0.01 0.02 0.02 20.67
GenAHT 62.78 3.78 59.00 61.33 6.02
CondAHT 26.88 11.84 15.04 15.63 44.05
Pumpw,AHT 0.0062 0.0053 0.0009 0.0009 85.48
EvaAHT 61.36 41.57 19.79 20.57 67.75

Proposed system (mode C) AbsAHT 74.66 72.93 1.73 1.80 97.68

HEsol,AHT 16.27 15.65 0.62 0.64 96.19
Pumpsol,AHT 0.03 0.01 0.02 0.02 20.67
GenAHT 62.78 3.78 59.00 61.33 6.02
CondAHT 26.88 11.84 15.04 15.63 44.05
Pumpw,AHT 0.0062 0.0053 0.0009 0.0009 85.48
EvaAHT 61.36 41.57 19.79 20.57 67.75

Proposed system (mode D) AbsAHT 68.70 66.87 1.83 2.04 97.34

HEsol,AHT 16.57 16.01 0.56 0.62 96.62
Pumpsol,AHT 0.03 0.01 0.02 0.03 20.20
GenAHT 57.56 2.15 55.41 61.65 3.74
CondAHT 24.72 10.91 13.81 15.36 44.13
Pumpw,AHT 0.0060 0.0049 0.0011 0.0012 81.67
EvaAHT 56.58 38.33 18.25 20.30 67.74

Fig. 6. Relative difference of energy and exergy efficiencies of the proposed system compared with the Configuration1 at different modes.

Costsaving, is calculated as [35]

Costen = Ensource × Encost (30)
Costsaving = Costtot,Config1 − Costtot,sys (28)
CostICR = Ż × τ (31)
where Costtot,Config1 and Costtot,sys are total costs of the Configuration1
and the proposed system, respectively. The total cost, Costtot is deter­ where Ensource is the required energy in the heat source and Encost is the
mined as: energy cost. Considering NG, BM and solar as the energy resources, the
Encost is reported as 0.067 $/kWh [36], 0.25 $/kWh [37], and 0.08
Costtot = Costen + CostICR + CDE × CC (29)
$/kWh [38], respectively.
where Costen is the cost of the energy consumption, CostICR is the ICR, To perform a proper investigation about cost saving, it is necessary to
CDE is the carbon dioxide emission, and CC (US$/metric ton of CO2 select the proper mode in which the power, cooling, and heating pro­
equivalent) is the carbon credit. ductions are the same as the Configuration1 (mode D). The obtained

13
S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Table 13
The main results of economic analysis for the configurations using AHT.
Parameter Mode A Mode B Mode C Mode D Configuration1

TCIORC (US$) 1,780,000 1,490,000 1,482,000 1,482,000 1,466,000

TCIARC (US$) 407,233 624,564 401,960 406,269 409,572
TCIAHT (US$) 699,358 699,358 699,358 451,023 0
TCIHeating (US$) 285,798 561,872 259,441 258,385 328,737
TCItot (US$) 3,172,389 3,375,794 2,842,759 2,597,677 1,583,567
Żtot (US$/kWh) 53.29 56.69 47.75 43.62 26.65632
WCtot (US$) 475858.4 506369.1 426413.9 389651.6 237535.1
TDItot (US$) 2,696,531 2,869,425 2,416,345 2,208,025 1,346,032
BDtot (US$) 179768.7 191,295 161089.7 147201.7 89735.46
ANCFtot (US$) 1,278,936 1,360,560 1,146,024 1,046,952 639751.7
AP (US$) 1,099,167 1,169,265 984934.3 899750.3 550016.2
ARR (%) 30.13 30.12 30.13 30.12 30.20
NPV (US$) 6,307,275 6,708,806 5,651,737 5,162,431 3,158,628
Celect. (US$/kWh) 0.379 0.516 0.435 0.397 0.337118
Cheat (US$/kWh) 0.069 0.073 0.045 0.056 0.047724
Ccool (US$/kWh) 0.266 0.138 0.239 0.218 0.18515
PP (year) 2.1 2.1 2.1 2.1 2.1

Fig. 7. Cost saving results by considering various sources of energy.

results of cost saving for the proposed system driven by the mentioned (a), exergy efficiency (b), ICR (c), and energy efficiency (d) are indi­
heat sources are presented in Fig. 7. It is obvious that using the proposed cated. It is seen that reducing the temperature of the stream passing
system even with energy sources which has low carbon emissions, it is through the evaporator of ORC decreases power production. This tem­
still much more economical than the Configuration1 and during a year perature reduction causes reduction in the size of the turbine but not the
the costs are reduced significantly. Also, it is noted that the maximum generator of ARC and HEheat, this is why the ICR is increased. As power
cost saving of 517018.1 US$/year (which is 32.65% of the investment production is the only variable parameter by changing HSS temperature,
cost of Configuration1) achieves when the required energy in the heat decreasing power production reduces energy efficiency. However, as
source is generated by using BM. Finally, it is concluded that in addition another remarkable result, it is seen that the proposed system works
to all advantages of the proposed system in energy, exergy and envi­ better than the Configuration1 even when the HSS temperature is 90 ◦ C.
ronmental viewpoints, it is much more economical than the At an HSS temperature of 90 ◦ C, the proposed system produces 27.35%
Configuration1. more power than the Configuration1.
The change of cooling production, exergy efficiency, ICR, and energy
efficiency of the proposed system at modes B and C are presented in
5.3. Parametric study Figs. 9 and 10, respectively. Regarding Fig. 9, due to the reduction of
HSS temperature, and consequently, the temperature of passing the
In previous works, the influencing parameters on the performance of stream through the generator of ARC, the cooling production (a), exergy
ORC, AHT and ARC have been investigated adequately. However, in the efficiency (b), and energy efficiency (d) decrease, while the ICR (c) in­
present work, the most important parameter affecting the performance creases. As it is seen in Fig. 10, the heating production (a), exergy effi­
of the whole system is the HSS temperature; therefore, the parametric ciency (b), and energy efficiency (d) have a declining behavior due to
study is carried out by changing the temperature of the HSS. Moreover, the reduction of HSS temperature, however, the ICR (c) increases. The
it is tried to consider the behavior of the proposed system compared with increase of ICR in Figs. 9 and 10 happens according to the increase in the
the Configuration1 in lower temperature heat sources. mass flow rate of passing the stream through the evaporator of ORC in
In Fig. 8, the effects of HSS temperature on the production of power

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S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Fig. 8. Comparison of the proposed system at mode A and Configuration1 at different HSS temperatures: a) Power production b) Exergy efficiency c) Investment cost
rate d) Energy efficiency.

lower temperatures to produce the same amount of power as those of higher HSS temperature, a lower HSS mass flow rate is required,
with higher HSS temperature (reason 1), and also, the temperature consequently, the smaller size of components is also required.
difference between the stream passing through the HEheat (reason 2). Until now it was shown that the proposed system has better perfor­
Both reasons increase the size of related components, and thus, increases mance than the Configuration1 even at lower HSS temperatures. How­
the ICR. As another important result, it was clarified in Figs. 9 and 10 ever, to show the effect of HSS temperature change on the mentioned
that the proposed system has higher cooling and heating productions parameters of each mode in Figs. 8–11, it is necessary to perform a
than the Configuration1 in lower HSS temperatures. Moreover, the sensitivity analysis. Accordingly, a simple sensitivity analysis is carried
maximum cooling and heating productions of the Configuration1 out. To perform the sensitivity analysis, the proposed system with the
happen at an HSS temperature of 120 ◦ C which are 291.10 kW and HSS temperature of 120 ◦ C is considered as the base case (BC). More­
1172.37 kW, respectively, however, the minimum mentioned pro­ over, a 5 ◦ C reduction in HSS temperature is applied to see the sensitivity
ductions in the proposed system with an HSS temperature of 90 ◦ C are of the affected parameters. As an indicator, a sensitivity influence co­
235.70 kW and 1045.15 kW. It shows the qualification of the proposed efficient (SIC) is employed as [39]
system even at low HSS temperatures from viewpoint of production.
OPi − OPBC IPi − IPBC
Fig. 11 displays the change of energy consumption rate, exergy ef­ SIC = / (32)
OPBC IPBC
ficiency, ICR, and energy efficiency of the proposed system and
Configuration3 due to the change of HSS temperature. In Configuration1, where OPi is the output value of a parameter corresponding to the input
it is not possible to produce the same amount of power, cooling and value for temperature i (IPi); subscript i shows the HSS temperature at
heating at all mentioned HSS temperatures; therefore, the Configura­ 115 ◦ C, and BC represents the HSS temperature of the base case which is
tion3 is considered instead. The Configuration3 has the same configu­ 120 ◦ C. The sensitivity influence coefficient of the mentioned parame­
ration as the proposed system but without an AHT. Regarding Fig. 11, ters in each mode is demonstrated separately in Fig. 12. It is seen that the
decreasing the HSS temperature increases energy consumption (a) but production of power, cooling, heating, and energy consumption rate are
decreases the exergy and energy efficiencies (b and d). To have the same the most affected parameters in modes A-D, respectively. On the other
amount of power, cooling and heating productions at lower HSS tem­ hand, the least affected parameter in the mentioned modes is the in­
peratures, more HSS should be produced which means more energy is vestment cost rate.
required. This is why the energy consumption rate increases in lower
temperatures. Considering Fig. 11c, by increasing the HSS temperature,
the ICR reduces, it happens because of the mass flow rate of HSS. At a

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S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Fig. 9. Comparison of the proposed system at mode B and Configuration1 at different HSS temperatures: a) Cooling production b) Exergy efficiency c) Investment
cost rate d) Energy efficiency.

5.4. Comparing the proposed system with other conventional basic 6. Conclusions
systems
In this work, an absorption heat transformer cycle is integrated with
In previous sections, it has shown that integrating an AHT cycle as an a conventional CCHP system, which is made of ORC, ARC, and a heat
auxiliary system with a CCHP system has remarkable effects on the exchanger. Three conventional CCHP systems with sequential (Config­
performance and costs of the proposed system in different modes. Also, uraion1), parallel (Configuraion2) and a combination of sequential and
it is noted that each mode has different results in energy and economic parallel (Configuraion3) configurations are considered. It is shown that
viewpoints. In this section, it is tried to compare other conventional among the Configuraion1, Configuraion2, and Configuraion3, Config­
CCHP systems having different configurations with the proposed system. uraion1 has the highest energy efficiency; therefore, it is selected to be
Thus, four systems including the proposed system, Configuration1, compared with the proposed integrated system with AHT. This inte­
Configuration2, and Configuration3 are considered. gration affects the outputs production and also energy consumption;
The Configuration2 has a parallel configuration in which the HSS is therefore, four different modes (standpoints) are considered to compare
divided into three parts and each part provides the required energy of the production of power (mode A), cooling (Mode B), heating (mode C),
producing power, cooling, and heating in related cycles and and energy consumption (mode D) of these two systems. Also, from the
components. exergy, environmental, and economic aspects, these two systems are
Considering the same amount of heating, cooling and power pro­ compared. The important conclusions of this study are summarized as
ductions, the mentioned systems are compared from viewpoints of en­ follows:
ergy efficiency, energy consumption rate, and ICR. The results are
reported in Table 14. It is seen that the proposed system has the highest • Comparing the proposed system with Configuration1, the amount of
energy efficiency and lowest energy consumption rate compared with power, cooling, and heating productions are developed 27.98%,
other mentioned systems. According to the additional costs resulting 102.15%, and 36.87%, respectively.
from the integration of an AHT with a CCHP system, it is rational that • Considering Mode D, it was indicated that the energy and exergy
other configurations have a lower investment cost rate than the pro­ efficiencies, energy consumption and carbon emission of the pro­
posed system. However, it is seen that the proposed system investment posed system are improved 17.68%, 17.68%, 15.03%, and 15.02%,
cost rate is 1.18 times higher than that of the Configuration1. Thus, it is respectively.
concluded that the proposed system has better performance among the • It is noted that for the heat source stream temperature range of 90 ◦ C
mentioned systems due to the highest energy efficiency, lowest energy − 120 ◦ C, the proposed system has better performance than the
consumption and acceptable ICR. Configuration1.

16
S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Fig. 10. Comparison of the proposed system at mode C and Configuration1 at different HSS temperatures: a) Heating production b) Exergy efficiency c) Investment
cost rate d) Energy efficiency.

17
S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Fig. 11. Comparison of the proposed system at mode D and Configuration3 at different HSS temperatures: a) Energy consumption b) Exergy efficiency c) Investment
cost rate d) Energy efficiency.

Fig. 12. Sensitivity influence coefficient of various parameters in Modes A-D.

18
S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Table 14 Validation, Formal analysis, Investigation, Writing - original draft,

Comparing the proposed system with other conventional Basic systems Writing - review & editing, Visualization, Supervision, Project admin­
regarding the same production of heating, cooling, and power. istration, Data curation. Alireza Hossein Nezhad: Conceptualization,
System Energy Energy consumption Investment cost rate Visualization, Supervision, Project administration, Data curation,
efficiency (%) rate (kW) (US$/h) Writing - review & editing. Alessandro Romagnoli: Conceptualization,
Configuration1 57.44 1890.00 37.03 Visualization, Supervision, Project administration, Data curation,
Configuration2 47.16 2301.87 34.99 Writing - review & editing. Bakytzhan Akhmetov: Visualization, Su­
Configuration3 50.11 2166.89 34.88 pervision, Project administration, Data curation, Writing - review &
Proposed 67.60 1606.05 43.62
editing.

• Despite the additional costs of integrating AHT with the CCHP sys­ Declaration of Competing Interest
tem, the proposed system has 517018.1 US$/year cost saving (which
is 32.65% of the investment cost of Configuration1) if biomass is used The authors declare that they have no known competing financial
as a heat source, which indicates its higher economic advantage than interests or personal relationships that could have appeared to influence
the Configuration1. the work reported in this paper.

CRediT authorship contribution statement

Saman Khalilzadeh: Conceptualization, Methodology, Software,

Appendix A. . Calculation procedure to calculate the heat transfer area of shell-and-tube heat exchangers

In this section, according to the heat exchanger type considered in this work, a procedure is presented to calculate the overall heat transfer co­
efficient and heat transfer area of the heat exchangers. At the end of this section, the overall heat transfer coefficients of different heat exchangers are
reported.
The heat exchangers considered in this work are the shell-and-tube type with a single pass shell and a counter flow (Fig. A1). It should be noted that
the tube layout of heat exchangers is considered 30◦ -triangular (Fig. A2). In Fig. A2, Ltp shows the tube pitch, which is the distance between the centers

Fig. A1. Schematic of a shell-and-tube heat exchanger.

Fig. A2. Schematic of a 30◦ - triangular tube layout.

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S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Table A1
Typical film heat transfer coefficient and fouling resistance for shell-and-tube heat exchangers [40].
Fluid conditions α (W/m2K) Rf (m2k/W)

Heat transfer
Ammonia_water 7000 1 × 10− 4
Water 1750 1 × 10− 4
R1234ze 1750 1 × 10− 4
4
LiBr_water 1125 1.5 × 10−

Condensing heat transfer

4
Ammonia_water 12,500 1× 10−
4
Water 12,500 1× 10−
4
R1234ze 5000 1× 10−
4
LiBr_water 1750 1× 10−

Vaporizing heat transfer

4
Ammonia_water 4000 1 × 10−
4
Water 6500 1 × 10−
4
R1234ze 1800 1 × 10−

Table A2
The input data to calculate the overall heat transfer coefficient of heat exchangers.
Parameter Value

Location of hot fluid Tube side

Outside diameter of the tube (mm) 19.0
Tube wall thickness (mm) 2.11
Tube pitch (mm) 23.6
Tube pattern 30◦ - Triangular
Baffle type Single segmental

Table A3
The overall heat transfer coefficient for the heat exchangers of the Configuration1 and proposed system.
Component U (kW/m2 K)

Mode A Mode B Mode C Mode D Configuration1

AHT
Absorber 0.61 0.61 0.61 0.61 –
Evaporator 1.02 1.02 1.02 1.02 –
Solution heat exchanger 0.50 0.50 0.50 0.50 –
Generator 0.31 0.31 0.31 0.31 –
Condensor 1.16 1.16 1.16 1.16 –

ORC
Evaporator 0.78 0.78 0.78 0.78 0.78
Condensor 1.03 1.03 1.03 1.03 1.03
Heat exchanger 0.77 0.77 0.77 0.77 0.77

ARC
Generator 1.02 1.02 1.02 1.02 1.02
Absorber 2.29 2.29 2.29 2.29 2.29
Condensor 1.16 1.16 1.16 1.16 1.16
Evaporator 1.70 1.70 1.70 1.70 1.70
Solution heat exchanger 2.03 2.03 2.03 2.03 2.03
Heat exchanger 2.29 2.29 2.29 2.29 2.29

Other components
Heating heat exchanger 0.77 0.77 0.77 0.77 0.77

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S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

of two adjacent tubes, and Do represents the outside diameter of the tube.
The heat load of a heat exchanger is determined as [40]
Q = UAtot FΔTLMTD (A.1)

where U is the overall heat transfer coefficient, Atot is the total heat transfer area required in the exchanger (calculated based on the outside diameter
of the tube), F is configuration correction factor (which is 1 for a single pass and counter-flow HE [40]), and ΔTLMTD is the logarithmic mean tem­
perature difference:
ΔT2 − ΔT1
ΔTLMTD = (A.2)
lnΔT 2
ΔT1

If there is a parallel flow:

{
ΔT1 = Thot,in − Tcold,in
(A.3)
ΔT2 = Thot,out − Tcold,out

If there is a counter flow:

{
ΔT1 = Thot,in − Tcold,out
(A.4)
ΔT2 = Thot,out − Tcold,in

where hot and cold subscripts represent hot and cold streams, also, in and out subscripts indicate the inlet or outlet of the HE.
The overall heat transfer coefficient is obtained as [40]:
1
U = [( ) ( )( ) ( )( ) ] (A.5)
1 Δxw Ao 1 Ao
αo + Rfo + kw Am
+ Rfi + αi Ai

where Δxw is the wall thickness, Ao and Ai are the outside and inside area of the tube, respectively. αi and αo are the film heat transfer coefficients for
the inside and outside of the tube, respectively. kw is the conductivity of tube wall (in this work, made of copper k = 385 W/m2 ◦ C [41]), Rfi and Rfo are
also the fouling resistances due to the fluid flow inside and outside of the tube, respectively. Am is the effective mean wall heat transfer area and
approximately determined as Am=πLΔxw.
To calculate the overall heat transfer coefficient using Eq. (A.5), having parameters αi, αo, Rfi and Rfo is necessary. In reference [40], there is some
information about the mentioned parameters for various process services and fluid conditions, which can be used to estimate the U in Eq. (A.5).
Table A1 shows the values of αi, αo, Rfi and Rfo considered for the calculation of heat exchangers in this work, which are reported from reference [40].
According to the aforementioned correlation and related values, the Atot is obtained. As stated before, this area is obtained based on the outside
diameter of the tube. To calculate the required outside area of a heat exchanger (Atot,HTA), the following correlation is used:
Atot,HTA = Atot F1 F2 F3 (A.6)

where F1 is the correction factor for the unit cell tube array (which is equal to 1 for 19.0-mm tubes on a 23.6-mm triangular pitch [40]), F2 is the
correction factor for the number of tube passes (which is equal to 1 for one tube pass [40]), and F3 is the correction factor for the shell construction/
tube bundle layout type (which is equal to 1 for a fixed tube sheet [40]). More information about F1, F2, and F3 is presented in Reference [40].
The following algorithm is used to calculate the total heat transfer area of the shell-and-tube heat exchangers in this work:

Step 1: Input the following data: Q, Thot,in, Thot,out, Tcold,in, Tcold,out, Di, Do, Ltp, Δxw, kw, αi, αo, Rfi, Rfo, F, F1, F2, and F3
Step 2: Calculate ΔTLMTD:
Step 2.1: Calculate ΔT1 using equations (A.3) and (A.4)
Step 2.2: Calculate ΔT2 using equations (A.3) and (A.4)
Step 2.3: Calculate ΔTLMTD using equation (A.2)
Step 3:Calculate U:
Step 3.1: Calculate Ao using equation Ao = πLDo
Step 3.2: Calculate Ai using equation Ai = πLDi
Step 3.3: Calculate Am using equation Am = πLΔxw
Step 3.4: Calculate U using equation (A.5)
Step 4: Calculate Atot using equation (A.1)
Step 5: Calculate Atot,HTA using equation (A.6)

The input data to calculate the overall heat transfer coefficients of HEs are reported in Table A2. The obtained results for overall heat transfer
coefficients of various heat exchangers at different modes are presented in Table A3.

Appendix B. . Thermodynamic properties of the proposed system

The thermodynamic properties of the proposed system working at mode B-D are presented in Tables B1–B3.

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S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Table B3
Thermodynamic properties of the proposed system working at mode D.
Point P (kPa) T (◦ C) h (kJ/kg) s (kJ/kgK) m˙ (kg/s) X (kg/kg) Point P (kPa) T (◦ C) h (kJ/kg) s (kJ/kg K) m˙ (kg/s) X (kg/kg)

1 200.0 120.0 503.80 1.53 2.00 – 25 58.27 85.20 2652.0 7.54 0.09 –
2 380.0 120.0 503.90 1.53 2.00 – 26 429.10 20.00 227.30 1.10 3.47 –
3 380.0 141.8 644.40 1.87 2.00 – 27 3632.0 21.90 230.50 1.10 3.47 –
4 200.0 120.2 644.40 1.89 0.05 – 28 3632.0 30.00 241.60 1.14 3.47 –
5 380.0 141.8 644.40 1.87 1.95 – 29 3632.0 128.80 467.70 1.76 3.47 –
6 13.96 52.50 219.80 0.74 1.95 – 30 429.10 55.34 432.30 1.79 3.47 –
7 200.0 120.0 503.80 1.53 3.33 – 31 429.10 43.93 421.20 1.75 3.47 –
8 200.0 120.0 503.80 1.53 1.69 – 32 13.96 10.00 41.99 0.151 6.49 –
9 200.0 120.0 503.80 1.53 1.64 – 33 13.96 41.72 174.9 0.595 3.38 –
10 200.0 86.51 362.4 1.15 1.69 – 34 200.00 47.00 196.9 0.665 5.33 –
11 200.0 85.94 360.0 1.15 1.64 – 35 286.79 13.00 − 175.1 0.03 0.85 57.53
12 200.0 86.23 361.2 1.15 3.33 – 36 681.12 13.02 − 174.6 0.03 0.85 57.53
13 200.0 87.22 365.14 1.16 3.38 – 37 681.12 39.12 − 55.64 0.43 0.85 57.53
14 200.0 68.23 285.7 0.93 3.38 – 38 681.12 54.05 3.02 0.64 0.69 47.53
15 4.3 30.0 125.70 0.44 3.38 – 39 681.12 21.23 − 144.0 0.16 0.69 47.53
16 6.86 88.43 261.90 0.48 1.14 62.60 40 286.79 21.31 − 144.0 0.17 0.69 47.53
17 58.27 110.70 261.90 0.59 1.14 62.60 41 681.12 39.12 1351.9 4.71 0.16 –
18 58.27 143.60 323.00 0.75 1.14 62.60 42 681.12 39.11 − 55.78 0.43 0.00 57.53
19 58.27 141.30 340.70 0.68 1.05 68.20 43 681.12 21.07 1288.1 4.50 0.16 –
20 58.27 102.50 274.10 0.51 1.05 68.20 44 681.12 13.00 60.24 0.23 0.16 –
21 6.86 102.50 274.10 0.51 1.05 68.20 45 681.12 − 2.23 − 10.31 − 0.03 0.16 –
22 6.86 85.00 2659.0 8.54 0.09 – 46 286.79 − 10.33 − 10.31 − 0.02 0.16 –
23 6.86 38.64 161.80 0.55 0.09 – 47 286.79 − 10.00 1224.3 4.65 0.16 –
24 58.27 38.64 161.90 0.55 0.09 – 48 286.79 5.59 1294.8 4.92 0.16 –

Points 16–21, X denotes XLiBr.

Points 35–41, X denotes XNH3.

Table B2
Thermodynamic properties of the proposed system working at mode C
Point P (kPa) T (◦ C) h (kJ/kg) s (kJ/kgK) m˙ (kg/s) X (kg/kg) Point P (kPa) T (◦ C) h (kJ/kg) s (kJ/kg K) m˙ (kg/s) X (kg/kg)

1 200 120.00 503.80 1.53 2.62 – 25 58.27 85.20 2652.0 7.54 0.10 –
2 380 120.00 503.90 1.53 2.62 – 26 429.1 20.0 227.3 1.10 3.47 –
3 380 141.80 602.80 1.77 2.62 – 27 3632.0 21.9 230.5 1.10 3.47 –
4 200 120.20 602.80 1.78 0.57 – 28 3632.0 30.0 241.6 1.14 3.47 –
5 380 141.80 602.80 1.77 2.05 – 29 3632.0 128.8 467.7 1.76 3.47 –
6 13.96 52.50 219.80 0.74 2.05 – 30 429.1 55.3 432.3 1.79 3.47 –
7 200.00 120.00 503.80 1.53 3.46 – 31 429.1 43.9 421.2 1.75 3.47 –
8 200.00 120.00 503.80 1.53 1.77 – 32 13.96 10.00 41.99 0.151 6.49 –
9 200.00 120.00 503.80 1.53 1.69 – 33 13.96 41.72 174.9 0.595 4.03 –
10 200.00 85.07 356.30 1.14 1.77 – 34 200.00 46.10 193.0 0.653 6.08 –
11 200.00 84.24 352.80 1.13 1.69 – 35 286.79 13.00 − 175.1 0.03 0.85 57.53
12 200.00 84.66 354.60 1.13 3.46 – 36 681.12 13.02 − 174.6 0.03 0.85 57.53
13 200.00 93.04 389.8 4.03 4.03 – 37 681.12 39.12 − 55.64 0.43 0.85 57.53
14 200.00 77.13 323.0 4.03 4.03 – 38 681.12 54.05 3.02 0.64 0.69 47.53
15 4.25 30.00 125.70 0.44 4.02 – 39 681.12 21.23 − 144.0 0.16 0.69 47.53
16 6.86 87.42 259.70 0.48 1.15 62.17 40 286.79 21.31 − 144.0 0.17 0.69 47.53
17 58.27 110.50 259.70 0.60 1.15 62.17 41 681.12 39.12 1351.9 4.71 0.16 –
18 58.27 142.40 319.40 0.75 1.15 62.17 42 681.12 39.11 − 55.78 0.43 0.00 57.53
19 58.27 140.70 339.70 0.68 1.05 68.20 43 681.12 21.07 1288.1 4.50 0.16 –
20 58.27 102.50 274.10 0.51 1.05 68.20 44 681.12 13.00 60.24 0.23 0.16 –
21 6.86 102.50 274.10 0.51 1.05 68.20 45 681.12 − 2.23 − 10.31 − 0.03 0.16 –
22 6.86 85.00 2659.0 8.54 0.10 – 46 286.79 − 10.33 − 10.31 − 0.02 0.16 –
23 6.86 38.64 161.80 0.55 0.10 – 47 286.79 − 10.00 1224.3 4.65 0.16 –
24 58.27 38.64 161.90 0.55 0.10 – 48 286.79 5.59 1294.8 4.92 0.16 –

Points 16–21, X denotes XLiBr.

Points 35–41, X denotes XNH3.

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S. Khalilzadeh et al. Energy Conversion and Management 228 (2021) 113677

Table B1
Thermodynamic properties of the proposed system working at mode B.
Point p (kPa) T (◦ C) h (kJ/kg) s (kJ/kgK) m˙ (kg/s) X (kg/kg) Point P (kPa) T (◦ C) h (kJ/kg) s (kJ/kg K) m˙ (kg/s) X (kg/kg)

1 200 120.00 503.80 1.53 2.62 – 25 58.27 85.20 2652.0 7.54 0.10 –
2 380 120.00 503.90 1.53 2.62 – 26 429.10 20.00 227.30 1.10 3.49 –
3 380 141.80 602.80 1.77 2.62 – 27 3632.0 21.90 230.50 1.10 3.49 –
4 200 120.20 602.80 1.78 0.56 – 28 3632.0 30.00 241.60 1.14 3.49 –
5 380 141.80 602.80 1.77 2.06 – 29 3632.0 128.80 467.70 1.76 3.49 –
6 13.96 52.50 219.80 0.74 2.06 – 30 429.10 55.34 432.30 1.79 3.49 –
7 200.00 120.00 503.80 1.53 3.46 – 31 429.10 43.93 421.20 1.75 3.49 –
8 200.00 120.00 503.80 1.53 1.77 – 32 13.96 10.00 41.99 0.151 6.49 –
9 200.00 120.00 503.80 1.53 1.69 – 33 13.96 41.72 174.9 0.595 4.02 –
10 200.00 85.07 356.30 1.14 1.77 – 34 200.00 46.10 193.0 0.653 6.08 –
11 200.00 84.24 352.80 1.13 1.69 – 35 286.79 13.00 − 175.1 0.03 1.72 57.53
12 200.00 84.66 354.60 1.13 3.46 – 36 681.12 13.02 − 174.6 0.03 1.72 57.53
13 200.00 92.86 389.10 1.23 4.02 – 37 681.12 39.12 − 55.64 0.43 1.72 57.53
14 200.00 66.00 253.40 0.84 4.02 – 38 681.12 54.05 3.02 0.64 1.39 47.53
15 4.25 30.00 125.70 0.44 4.02 – 39 681.12 21.23 − 144.0 0.16 1.39 47.53
16 6.86 87.42 259.70 0.48 1.15 62.17 40 286.79 21.31 − 144.0 0.17 1.39 47.53
17 58.27 110.50 259.70 0.60 1.15 62.17 41 681.12 39.12 1351.9 4.71 0.33 –
18 58.27 142.40 319.40 0.75 1.15 62.17 42 681.12 39.11 − 55.78 0.43 0.00 57.53
19 58.27 140.70 339.70 0.68 1.05 68.20 43 681.12 21.07 1288.1 4.50 0.33 –
20 58.27 102.50 274.10 0.51 1.05 68.20 44 681.12 13.00 60.24 0.23 0.33 –
21 6.86 102.50 274.10 0.51 1.05 68.20 45 681.12 − 2.23 − 10.31 − 0.03 0.33 –
22 6.86 85.00 2659.0 8.54 0.10 – 46 286.79 − 10.33 − 10.31 − 0.02 0.33 –
23 6.86 38.64 161.80 0.55 0.10 – 47 286.79 − 10.00 1224.3 4.65 0.33 –
24 58.27 38.64 161.90 0.55 0.10 – 48 286.79 5.59 1294.8 4.92 0.33 –

Points 16–21, X denotes XLiBr.

Points 35–41, X denotes XNH3.

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