# Comparison Study of Four Typical ORC Configurations for Different Waste Heat Characteristics of Engine

Authors: Hua Tian; Zhaoxian He; Xuanang Zhang; Ligeng Li; Jinwen Cai; Xuan Wang; Xingyu Liang; Gequn Shu
Journal: International Journal of Energy Research
 (2023)

---

## Abstract

The organic Rankine cycle (ORC), as the most promising technology for engine waste heat recovery, has many configurations available to match the complicated engine waste heat conditions. This paper is aimed at revealing the characteristics of the engine waste heat condition for which different ORC configurations are matched so as to give selection guidance. The engine waste heat sources are characterized by the exhaust gas temperature (300°C-650°C) and the heat ratio of the high temperature waste heat to low temperature waste heat (0.5-1.5). Four mostly investigated configurations including the single-loop ORC, the cascade dual-loop ORC, the dual-pressure ORC, and the CO2 transcritical Rankine cycle (CTRC) are selected and modified to enable efficient utilization of all waste heat sources. Their thermodynamic performance is optimized for different waste heat characteristics. The results show improvements in net power for all configurations when exhaust gas temperature and heat ratio rise from 300°C, 0.5, to 650°C, 1.5, respectively, of which the high-temperature split CTRC (HTS-CTRC) recorded the largest increase of 183.58%, while the split regenerative ORC (SR-ORC), dual-pressure ORC (D-ORC), and split cascade ORC (SC-ORC) increased by 17.41%, 23.71%, and 35.45%, respectively. Two selection maps are established to provide general references for quickly selecting the configuration according to different waste heat characteristics whatever the engine type and operating conditions.


---

## Body

## 1. Introduction

Worldwide CO2 emissions are growing annually and have surpassed 3,712 billion tons in 2021, which is the primary driver of global warming [1]. The internal combustion engine (ICE), as the main power source of transportation, consumes about 70% of fossil oil and contributed 10% of CO2 emissions in the past 25 years [2]. The thermal balance of ICE indicates that more than 50% of the total fuel combustion heat turns into waste heat released into the environment [3]. If this waste heat is effectively utilized, the efficiency of the engine can be improved greatly thereby saving energy and alleviating the climate change problems [4]. Therefore, many waste heat recovery technologies for ICE have been proposed, of which the organic Rankine cycle (ORC) is the most promising route owing to its advantages such as high efficiency, suitable size, and low cost [5, 6].The waste heat of the engine features multiple categories and large-gradient temperature difference as shown in Table1. They could be classified into two types, the high-temperature waste heat sources including exhaust gas (EG) and exhaust gas recirculation (EGR) and low-temperature waste heat sources including jacket water (JW), charge air (CA), and lubricating oil (LO) [7]; their temperature, especially that of the EG, varies considerably based on different engine types, among which the diesel engine has lower EG temperature compared to other fuel engines. Besides, different engine operating conditions also have different waste heat source conditions. For example, in the study by [8], as the load of a 0.638 L diesel engine goes from 10% to 100%, the waste heat energy rises from 2.92 kW to 9.21 kW and the heat ratio between EG and JW goes from 0.87 to 1.02. With such complicated and varied waste heat source conditions of the engine, there are many ORC configurations available to match [7, 9, 10], and they could be classified into four typical forms, the single-loop ORC, the cascade dual-loop ORC, the dual-pressure ORC, and the CO2-based transcritical Rankine cycle.Table 1
Engine waste heat conditions under full operating region from different references.
Engine parametersTemperature of different waste heat sources (°C)Exhaust gas (EG)Exhaust gas recirculation (EGR)Jacket water (JW)Charge air (CA)8.424 L, diesel [24]205.9-471.1282.3-593.874-8457.31-152.711.596 L, diesel [26]221.85-393.85n.a.9035.85-121.851.02 L, gasoline [46]400-800n.a.86.6-89.1n.a.2.0 L, gasoline [36]649-772n.a.82.5-82.7n.a.8.9 L, natural gas [23]526.85-606.85n.a.79.85-94.8546.85-126.8512.939 L, natural gas [25]439.35-581.85598.85-773.8583.85-91.85n.a.2.3 L, hydrogen [47]204.85-594.96n.a.60-80n.a.The single-loop ORC which focuses on recovering EG was considered initially due to its simplicity and low cost [11–15]. Many researchers have added a regenerator and further utilized the JW to improve the single-loop performance [16–20]. However, it is difficult to effectively utilize all the waste heat of the engine if only a single-loop ORC is used [21, 22]. The cascade dual-loop ORC combines the high-temperature loop (HT-loop) and the low-temperature loop (LT-loop) by using an intermediate heat exchanger, enabling total and gradient recovery of engine waste heat sources. There have been several researches [23–28] on simulations regarding the cascade dual-loop ORC, but its complexity, high cost, and control difficulty are obstacles to practical application. The dual-pressure ORC also has HT-loop and LT-loop, but they are connected by the condenser hence are easier to control [29–32]. Kocher [30] has successfully coupled the dual-pressure ORC mechanically to the diesel engine; their experiment demonstrated that 19.3 kW net power can be generated by the ORC-WHR system, achieving 54% BTE. Nevertheless, these ORC configurations using organic working fluids are at risk of fluid decomposition at high operating temperatures, which constrains ORC to match the HT heat source. The CO2 transcritical Rankine cycle (CTRC) is another attractive route as CO2 is an environment friendly and reliable working fluid with no limitation of decomposition. Substantial efforts [33–38] have been made by Shu’ group for many years to apply CTRC to engine waste heat recovery. They found that the specific heat capacity properties of CO2 make it appropriate for recovering EG and JW simultaneously [38] and have conducted many layout modifications [33, 36–38] as well as experiments [34, 35].The basis for selecting the most suitable one among these configurations is the guarantee of good performance of engine waste heat recovery. Many researchers have compared different configurations in terms of thermodynamics, economic performance, etc. [29, 33, 38–41]. Apostol et al. [39] compared the thermodynamic performance of six ORC configurations including BORC, PHORC, RORC, 2ORC, DORC, and PHRDORC under the fix waste heat condition of a 40 kW turbocharged diesel engine. The result showed that the 2ORC (a dual-pressure form) using R1336mzz in the LT-loop and ethanol in the HT-loop generated the max net power of 5.46 kW. Chen et al. [29] presented a novel dual-pressure ORC called CCE-ORC for recovering EG and JW of a diesel engine. The comparison with the cascade dual-loop ORC showed that the CCE-ORC produced a larger net power under the entire engine operating conditions especial at low engine load and rotating speed. Li et al. [33] proposed three novel split CTRC configurations to overcome the temperature interference problem. The result showed that the high-temperature split (HTS) system is the most competitive configuration compared to the LTS and MTS systems, achieving the maximum net power with best thermoeconomic performance under the given waste heat condition of the diesel engine. Shu et al. [41] proposed a general assessment method for ORC performance named MA-ES, which can provide comprehensive comparison with various goals by combining 1st law, 2nd law, and economic factors. The example case of ORC configuration comparison for engine WHR using the MA-ES showed that the dual-loop ORC outperforms the single-loop ORC in thermodynamic indexes while the single-loop ORC has better economic performance. However, these previous configuration comparison studies have been carried out under the specific waste heat condition, and there is no identification what waste heat conditions these configurations are suitable for. Some efforts have been conducted in other fields to demonstrate that different configurations are suitable for different waste heat conditions [42–45]. Yet their conclusions are specific to a single waste heat source and not applicable to an engine with multigrade waste heat sources.Therefore, this paper is aimed at investigating the thermodynamic performance of different ORC configurations under different waste heat conditions of the engine. To cover most waste heat conditions of the engine, we characterize the engine waste heat sources by their temperature and energy distribution. Four mostly investigated configurations including the single-loop ORC, the cascade dual-loop ORC, the dual-pressure ORC, and the CO2 transcritical power cycle (CTRC) have been selected and modified to enable efficient utilization of all waste heat of the engine. Each configuration is optimized their net power according to different waste heat characteristics. By comparing their net power maps, two selection maps are obtained, which have general applicability for quickly selecting the most suitable configurations according to different waste heat characteristics whatever engine type and operating conditions. In a word, the innovation of this paper is to reveal the characteristic of the engine waste heat condition for which different ORC configurations are matched and provide references when selecting the most suitable ORC configuration for engine waste heat recovery.


## 2. System Description

### 2.1. The Engine Waste Heat Source

Different engines as well as engines at different operating conditions have different waste heat source conditions. In order to investigate a large range of possible applications, we consider the engine waste heat condition to be characterized by EG temperature (300°C~650°C) and the heat ratio (0.5~1.5). The heat ratio (HR) is defined as the energy ratio between the HT waste heat sources and the LT waste heat sources, and it increases along with the engine load and speed as the HT waste heat sources grow faster than the LT waste heat sources [23, 25, 26]. The original waste heat source is introduced by the lowest fuel consumption point of a 12.4 L turbocharged diesel engine for a heavy-duty truck, which meets China VI emission standards. Detailed parameters are listed in Table 2. It is noteworthy that the waste heat of EGR and LO of this engine is included in the JW. Therefore, in this paper, the high-temperature heat source is EG and the low-temperature heat source contains JW and CA. The original waste heat source is redistributed to simulate different engine waste heat conditions as per the following approach: (1) The total amount of the waste heat after redistribution remains at 145.91 kW which is equal to the original one. (2) The amount of the high- and low-temperature waste heat sources is calculated by the following equation:
(1)HeatratioHR=QhighQlow=QegQca+Qjw.Table 2
Main information of SINOTRUK MC13 diesel engine waste heat source.
PropertiesValuesUnitsTemperature of the exhaust gas323.95°CMass flow rate of the exhaust gas0.2354kg/sTemperature of the jacket water inlet and outlet91-94.8°CMass flow rate of the jacket water4.623kg/sTemperature of the charge air inlet and outlet55-112.69°CMass flow rate of the charge air0.2284kg/s(3) The temperature of the JW is kept the same as the reference temperature as it does not vary much between different engines as well as engines at different operating conditions; its effect on ORC performance will be discussed in Section4.2. (4) The CA temperature and mass flow rate are estimated by the fitting relation of experimental data to the EG of the reference engine. Overall, when the HR and EG temperature are determined, the mass flow rate of EG and CA conditions can be calculated, leaving the residual heat of JW. Figure 1 shows the heat redistribution result of the reference waste heat source. It is worth observing that the high waste heat values of CA occur at high HR with low EG temperature because the low heat capacity of CA makes the mass flow rate dominate the amount of its waste heat sizes.Figure 1
The redistribution result of the original waste heat source: (a) exhaust gas, (b) charge air, and (c) jacket water.
(a)(b)(c)


### 2.2. The Engine Waste Heat Recovery Systems

In this paper, four ORC configurations including the split regenerative ORC (SR-ORC), the split cascade ORC (SC-ORC), the dual-pressure ORC (D-ORC), and the high-temperature split CTRC (HTS-CTRC) are proposed. Each configuration contains a regenerator to improve thermal efficiency and reduce the condenser load. Due to the low-temperature characteristic of JW, the higher the evaporative pressure (EP), the lower the utilization of JW when the pinch limitation in the heat exchanger occurs [16]. Therefore, when it comes to recovering waste heat of JW in this paper, the EP is set low to match the JW. Despite the four configurations having different structures, all of them have four base thermodynamic processes: (1) compression process in the pump, (2) heat addition process in the heat exchanger, (3) expansion process in the expander, and (4) heat rejection process in the heat exchanger. Detailed descriptions are as follows.


#### 2.2.1. Split Regenerative ORC (SR-ORC)

Figure2 shows the schematic and corresponding T-s diagram of the SR-ORC. It is modified by the single-loop ORC. Previous research has seldom considered the recovery of CA in the single-loop ORC as the CA temperature range overlaps with that of the regenerator or preheater (JW). This paper proposed a new split structure where the CAC and regenerator are in parallel to solve their temperature interference problem. That is, after being compressed by the pump (1-2), the working fluid is split into two streams: one is heated by CAC (2-C) and the other is heated by the regenerator (2-R). Then, two streams converged into one stream (C+R⟶M) and continue to be heated into the superheat state by JW and EG sequentially (M-5). Finally, the expansion and condensation process (6-1) will be conducted successively to complete a cycle.Figure 2
(a) Schematic diagram of the split regenerative ORC. (b) T-s diagram of the split regenerative ORC (for the original waste heat source).
(a)(b)


#### 2.2.2. Split Cascade ORC (SC-ORC)

The SC-ORC is a cascade dual-loop ORC that contains the HT-loop and LT-loop as is shown in Figure3. Their working fluids are different in that the HT-loop uses higher critical temperature fluids like ethanol and water [48]. The HT-loop is a basic ORC cycle without a regenerator and only utilizes the high-temperature part of EG as the heat source. The JW, CA, low-temperature part of EG, and condensation heat of the HT-loop are utilized as the heat source of the LT-loop. Same as SR-ORC, the LT-loop with a regenerator has a split structure, but notably, if the thermodynamic state after converging is still in the subcooled condition (before 3L), the rest part of the EG waste heat will continue to be utilized as long as the acid dew point has not been reached. As a result, the EG flow would be separated into three parts, allowing a better match between the cycle and EG. The performance of the LT-loop is affected by the HT-loop as they are connected by the intermediate heat exchanger, LT-Evaporator and LT-Preheater 1. It could be also understood that the SC-ORC evolved from SR-ORC by intercepting the high-temperature part of EG to do another cycle. Compared to the previous studies of the cascade dual-loop ORC, the SC-ORC in this paper utilizes the EG waste heat more thoroughly, and the LT-loop has also overcome the issue of temperature interference between the CAC and regenerator like the SR-ORC.Figure 3
(a) Schematic diagram of the split cascade ORC. (b) T-s diagram of the split cascade ORC (for the original waste heat source).
(a)(b)


#### 2.2.3. Dual-Pressure ORC (D-ORC)

As is shown in Figure4. The D-ORC is also a dual-loop form but different from SC-ORC; the HT-loop with regenerator and LT-loop without regenerator use the same working fluid and have a consistent condensing pressure to share the same condenser. The working fluid is divided into two streams: the HT-loop corresponds to high-pressure stream to recover HT waste heat (EG) and the LT-loop corresponds to the low-pressure stream to recover LT waste heat (JW and CA). The two streams will mix together (6L+6H⟶6M) after the expansion process of both loops and then flow into the regenerator. Next, the residual heat would be released into the environment via the condenser, and then, the working fluid is pressurized into two pressure streams to supply the HT-loop and LT-loop via the HT-Pump and LT-Pump, respectively, allowing them to operate separately like the single-loop ORC system. The dual-pressure ORC configuration has been proven to perform well in experiments for engine waste heat recovery by Kocher [30].Figure 4
(a) Schematic diagram of the split cascade ORC [30]. (b) T-s diagram of the split cascade ORC (for the original waste heat source).
(a)(b)


#### 2.2.4. High-Temperature Split CTRC (HTS-CTRC)

Compared to the organic working fluid, CO2 has a lower critical temperature, so the configuration that uses CO2 is called the transcritical cycle. Besides, the pressure ratio of CTRC is lower than that of ORC, and there is much more sensible heat which could be utilized after the expansion process [38]. Adding the regenerator could effectively enhance CTRC performance, but part of the exhaust gas heat cannot be utilized because of temperature interference. Therefore, the split structure should also be considered. In this paper, high-temperature split CTRC (HTS-CTRC) is proposed that was firstly presented by [33], and we further add the CAC to the HTS-CTRC in this paper to enable the use of all engine waste heat sources. The schematic and corresponding T-s diagram of HTS-CTRC are shown in Figure 5. The working fluid is firstly pressurized into the transcritical state by pump (1-2) and heated by JW and CA (2-3). Then, it splits into two streams: one is heated by the regenerator and the other is heated by the low-temperature part of EG. Finally, the two streams converge into one stream and continue to be heated by the high-temperature part of EG (4-5), expanded by expander (5-6), and condensed by the regenerator and condenser (6-1) to complete a cycle.Figure 5
(a) Schematic diagram of the split cascade ORC. (b) T-s diagram of the split cascade ORC (for the original waste heat source).
(a)(b)


### 2.3. Working Fluid Selection

In this paper, R245fa is chosen as the working fluid of SR-ORC, D-ORC, and LT-loop of SC-ORC. Many studies also use R245fa as the working fluid because of its high thermodynamic properties, low global warming potential (GWP), ozone depletion potential (ODP), and low cost [7]. Although alternative working fluids like R123 and R1233zd may have better performance, they are either environmentally unfriendly or too expensive. Ethanol was chosen as the working fluid for the HT-loop of SC-ORC because alcohols are very suitable for recovering the waste heat from the high- to medium-temperature sources [7]. The thermodynamic properties of these working fluids at various states are evaluated using REFPROP V10.0.


## 2.1. The Engine Waste Heat Source

Different engines as well as engines at different operating conditions have different waste heat source conditions. In order to investigate a large range of possible applications, we consider the engine waste heat condition to be characterized by EG temperature (300°C~650°C) and the heat ratio (0.5~1.5). The heat ratio (HR) is defined as the energy ratio between the HT waste heat sources and the LT waste heat sources, and it increases along with the engine load and speed as the HT waste heat sources grow faster than the LT waste heat sources [23, 25, 26]. The original waste heat source is introduced by the lowest fuel consumption point of a 12.4 L turbocharged diesel engine for a heavy-duty truck, which meets China VI emission standards. Detailed parameters are listed in Table 2. It is noteworthy that the waste heat of EGR and LO of this engine is included in the JW. Therefore, in this paper, the high-temperature heat source is EG and the low-temperature heat source contains JW and CA. The original waste heat source is redistributed to simulate different engine waste heat conditions as per the following approach: (1) The total amount of the waste heat after redistribution remains at 145.91 kW which is equal to the original one. (2) The amount of the high- and low-temperature waste heat sources is calculated by the following equation:
(1)HeatratioHR=QhighQlow=QegQca+Qjw.Table 2
Main information of SINOTRUK MC13 diesel engine waste heat source.
PropertiesValuesUnitsTemperature of the exhaust gas323.95°CMass flow rate of the exhaust gas0.2354kg/sTemperature of the jacket water inlet and outlet91-94.8°CMass flow rate of the jacket water4.623kg/sTemperature of the charge air inlet and outlet55-112.69°CMass flow rate of the charge air0.2284kg/s(3) The temperature of the JW is kept the same as the reference temperature as it does not vary much between different engines as well as engines at different operating conditions; its effect on ORC performance will be discussed in Section4.2. (4) The CA temperature and mass flow rate are estimated by the fitting relation of experimental data to the EG of the reference engine. Overall, when the HR and EG temperature are determined, the mass flow rate of EG and CA conditions can be calculated, leaving the residual heat of JW. Figure 1 shows the heat redistribution result of the reference waste heat source. It is worth observing that the high waste heat values of CA occur at high HR with low EG temperature because the low heat capacity of CA makes the mass flow rate dominate the amount of its waste heat sizes.Figure 1
The redistribution result of the original waste heat source: (a) exhaust gas, (b) charge air, and (c) jacket water.
(a)(b)(c)


## 2.2. The Engine Waste Heat Recovery Systems

In this paper, four ORC configurations including the split regenerative ORC (SR-ORC), the split cascade ORC (SC-ORC), the dual-pressure ORC (D-ORC), and the high-temperature split CTRC (HTS-CTRC) are proposed. Each configuration contains a regenerator to improve thermal efficiency and reduce the condenser load. Due to the low-temperature characteristic of JW, the higher the evaporative pressure (EP), the lower the utilization of JW when the pinch limitation in the heat exchanger occurs [16]. Therefore, when it comes to recovering waste heat of JW in this paper, the EP is set low to match the JW. Despite the four configurations having different structures, all of them have four base thermodynamic processes: (1) compression process in the pump, (2) heat addition process in the heat exchanger, (3) expansion process in the expander, and (4) heat rejection process in the heat exchanger. Detailed descriptions are as follows.


### 2.2.1. Split Regenerative ORC (SR-ORC)

Figure2 shows the schematic and corresponding T-s diagram of the SR-ORC. It is modified by the single-loop ORC. Previous research has seldom considered the recovery of CA in the single-loop ORC as the CA temperature range overlaps with that of the regenerator or preheater (JW). This paper proposed a new split structure where the CAC and regenerator are in parallel to solve their temperature interference problem. That is, after being compressed by the pump (1-2), the working fluid is split into two streams: one is heated by CAC (2-C) and the other is heated by the regenerator (2-R). Then, two streams converged into one stream (C+R⟶M) and continue to be heated into the superheat state by JW and EG sequentially (M-5). Finally, the expansion and condensation process (6-1) will be conducted successively to complete a cycle.Figure 2
(a) Schematic diagram of the split regenerative ORC. (b) T-s diagram of the split regenerative ORC (for the original waste heat source).
(a)(b)


### 2.2.2. Split Cascade ORC (SC-ORC)

The SC-ORC is a cascade dual-loop ORC that contains the HT-loop and LT-loop as is shown in Figure3. Their working fluids are different in that the HT-loop uses higher critical temperature fluids like ethanol and water [48]. The HT-loop is a basic ORC cycle without a regenerator and only utilizes the high-temperature part of EG as the heat source. The JW, CA, low-temperature part of EG, and condensation heat of the HT-loop are utilized as the heat source of the LT-loop. Same as SR-ORC, the LT-loop with a regenerator has a split structure, but notably, if the thermodynamic state after converging is still in the subcooled condition (before 3L), the rest part of the EG waste heat will continue to be utilized as long as the acid dew point has not been reached. As a result, the EG flow would be separated into three parts, allowing a better match between the cycle and EG. The performance of the LT-loop is affected by the HT-loop as they are connected by the intermediate heat exchanger, LT-Evaporator and LT-Preheater 1. It could be also understood that the SC-ORC evolved from SR-ORC by intercepting the high-temperature part of EG to do another cycle. Compared to the previous studies of the cascade dual-loop ORC, the SC-ORC in this paper utilizes the EG waste heat more thoroughly, and the LT-loop has also overcome the issue of temperature interference between the CAC and regenerator like the SR-ORC.Figure 3
(a) Schematic diagram of the split cascade ORC. (b) T-s diagram of the split cascade ORC (for the original waste heat source).
(a)(b)


### 2.2.3. Dual-Pressure ORC (D-ORC)

As is shown in Figure4. The D-ORC is also a dual-loop form but different from SC-ORC; the HT-loop with regenerator and LT-loop without regenerator use the same working fluid and have a consistent condensing pressure to share the same condenser. The working fluid is divided into two streams: the HT-loop corresponds to high-pressure stream to recover HT waste heat (EG) and the LT-loop corresponds to the low-pressure stream to recover LT waste heat (JW and CA). The two streams will mix together (6L+6H⟶6M) after the expansion process of both loops and then flow into the regenerator. Next, the residual heat would be released into the environment via the condenser, and then, the working fluid is pressurized into two pressure streams to supply the HT-loop and LT-loop via the HT-Pump and LT-Pump, respectively, allowing them to operate separately like the single-loop ORC system. The dual-pressure ORC configuration has been proven to perform well in experiments for engine waste heat recovery by Kocher [30].Figure 4
(a) Schematic diagram of the split cascade ORC [30]. (b) T-s diagram of the split cascade ORC (for the original waste heat source).
(a)(b)


### 2.2.4. High-Temperature Split CTRC (HTS-CTRC)

Compared to the organic working fluid, CO2 has a lower critical temperature, so the configuration that uses CO2 is called the transcritical cycle. Besides, the pressure ratio of CTRC is lower than that of ORC, and there is much more sensible heat which could be utilized after the expansion process [38]. Adding the regenerator could effectively enhance CTRC performance, but part of the exhaust gas heat cannot be utilized because of temperature interference. Therefore, the split structure should also be considered. In this paper, high-temperature split CTRC (HTS-CTRC) is proposed that was firstly presented by [33], and we further add the CAC to the HTS-CTRC in this paper to enable the use of all engine waste heat sources. The schematic and corresponding T-s diagram of HTS-CTRC are shown in Figure 5. The working fluid is firstly pressurized into the transcritical state by pump (1-2) and heated by JW and CA (2-3). Then, it splits into two streams: one is heated by the regenerator and the other is heated by the low-temperature part of EG. Finally, the two streams converge into one stream and continue to be heated by the high-temperature part of EG (4-5), expanded by expander (5-6), and condensed by the regenerator and condenser (6-1) to complete a cycle.Figure 5
(a) Schematic diagram of the split cascade ORC. (b) T-s diagram of the split cascade ORC (for the original waste heat source).
(a)(b)


## 2.2.1. Split Regenerative ORC (SR-ORC)

Figure2 shows the schematic and corresponding T-s diagram of the SR-ORC. It is modified by the single-loop ORC. Previous research has seldom considered the recovery of CA in the single-loop ORC as the CA temperature range overlaps with that of the regenerator or preheater (JW). This paper proposed a new split structure where the CAC and regenerator are in parallel to solve their temperature interference problem. That is, after being compressed by the pump (1-2), the working fluid is split into two streams: one is heated by CAC (2-C) and the other is heated by the regenerator (2-R). Then, two streams converged into one stream (C+R⟶M) and continue to be heated into the superheat state by JW and EG sequentially (M-5). Finally, the expansion and condensation process (6-1) will be conducted successively to complete a cycle.Figure 2
(a) Schematic diagram of the split regenerative ORC. (b) T-s diagram of the split regenerative ORC (for the original waste heat source).
(a)(b)


## 2.2.2. Split Cascade ORC (SC-ORC)

The SC-ORC is a cascade dual-loop ORC that contains the HT-loop and LT-loop as is shown in Figure3. Their working fluids are different in that the HT-loop uses higher critical temperature fluids like ethanol and water [48]. The HT-loop is a basic ORC cycle without a regenerator and only utilizes the high-temperature part of EG as the heat source. The JW, CA, low-temperature part of EG, and condensation heat of the HT-loop are utilized as the heat source of the LT-loop. Same as SR-ORC, the LT-loop with a regenerator has a split structure, but notably, if the thermodynamic state after converging is still in the subcooled condition (before 3L), the rest part of the EG waste heat will continue to be utilized as long as the acid dew point has not been reached. As a result, the EG flow would be separated into three parts, allowing a better match between the cycle and EG. The performance of the LT-loop is affected by the HT-loop as they are connected by the intermediate heat exchanger, LT-Evaporator and LT-Preheater 1. It could be also understood that the SC-ORC evolved from SR-ORC by intercepting the high-temperature part of EG to do another cycle. Compared to the previous studies of the cascade dual-loop ORC, the SC-ORC in this paper utilizes the EG waste heat more thoroughly, and the LT-loop has also overcome the issue of temperature interference between the CAC and regenerator like the SR-ORC.Figure 3
(a) Schematic diagram of the split cascade ORC. (b) T-s diagram of the split cascade ORC (for the original waste heat source).
(a)(b)


## 2.2.3. Dual-Pressure ORC (D-ORC)

As is shown in Figure4. The D-ORC is also a dual-loop form but different from SC-ORC; the HT-loop with regenerator and LT-loop without regenerator use the same working fluid and have a consistent condensing pressure to share the same condenser. The working fluid is divided into two streams: the HT-loop corresponds to high-pressure stream to recover HT waste heat (EG) and the LT-loop corresponds to the low-pressure stream to recover LT waste heat (JW and CA). The two streams will mix together (6L+6H⟶6M) after the expansion process of both loops and then flow into the regenerator. Next, the residual heat would be released into the environment via the condenser, and then, the working fluid is pressurized into two pressure streams to supply the HT-loop and LT-loop via the HT-Pump and LT-Pump, respectively, allowing them to operate separately like the single-loop ORC system. The dual-pressure ORC configuration has been proven to perform well in experiments for engine waste heat recovery by Kocher [30].Figure 4
(a) Schematic diagram of the split cascade ORC [30]. (b) T-s diagram of the split cascade ORC (for the original waste heat source).
(a)(b)


## 2.2.4. High-Temperature Split CTRC (HTS-CTRC)

Compared to the organic working fluid, CO2 has a lower critical temperature, so the configuration that uses CO2 is called the transcritical cycle. Besides, the pressure ratio of CTRC is lower than that of ORC, and there is much more sensible heat which could be utilized after the expansion process [38]. Adding the regenerator could effectively enhance CTRC performance, but part of the exhaust gas heat cannot be utilized because of temperature interference. Therefore, the split structure should also be considered. In this paper, high-temperature split CTRC (HTS-CTRC) is proposed that was firstly presented by [33], and we further add the CAC to the HTS-CTRC in this paper to enable the use of all engine waste heat sources. The schematic and corresponding T-s diagram of HTS-CTRC are shown in Figure 5. The working fluid is firstly pressurized into the transcritical state by pump (1-2) and heated by JW and CA (2-3). Then, it splits into two streams: one is heated by the regenerator and the other is heated by the low-temperature part of EG. Finally, the two streams converge into one stream and continue to be heated by the high-temperature part of EG (4-5), expanded by expander (5-6), and condensed by the regenerator and condenser (6-1) to complete a cycle.Figure 5
(a) Schematic diagram of the split cascade ORC. (b) T-s diagram of the split cascade ORC (for the original waste heat source).
(a)(b)


## 2.3. Working Fluid Selection

In this paper, R245fa is chosen as the working fluid of SR-ORC, D-ORC, and LT-loop of SC-ORC. Many studies also use R245fa as the working fluid because of its high thermodynamic properties, low global warming potential (GWP), ozone depletion potential (ODP), and low cost [7]. Although alternative working fluids like R123 and R1233zd may have better performance, they are either environmentally unfriendly or too expensive. Ethanol was chosen as the working fluid for the HT-loop of SC-ORC because alcohols are very suitable for recovering the waste heat from the high- to medium-temperature sources [7]. The thermodynamic properties of these working fluids at various states are evaluated using REFPROP V10.0.


## 3. Thermodynamic Model

In this part, the model assumptions, energy model, split model of four ORC systems as well as their model validation are presented.


### 3.1. Assumptions

Some reasonable assumptions before establishing the four ORC configuration models are as follows:(i)
Ambient temperature and pressure are 20°C and 101.3 kPa, respectively(ii)
All components are modelled for steady-state conditions [33](iii)
Heat losses and pressure drops as well as split and convergence losses are ignored [33](iv)
The isentropic efficiencies of the expander and pump are supposed to be 0.7 and 0.8, respectively(v)
The minimum superheat degree on the ORC configuration is set at 5 K to protect the expander(vi)
The pinch point temperature difference of the heat exchanger considering the exhaust gas, charge air, jacket water, and regenerator is set at 30 K,30 K,5 K and 10 K, respectively(vii)
The condenser subcooling is 1 K(viii)
The composition of exhaust gas is 12%CO2, 4%H2O, 73%N2, and 11%O2 while the charge air is 5%CO2, 75%N2, and 24.5%O2(ix)
The acid dew point of the exhaust gas is 85°C based on the model of [49]


### 3.2. Energy Model

Since four ORC systems have the same base thermodynamic processes, the energy balance equations of their components and the system efficiency equation based on the first law of thermodynamics can be expressed in a unified form as listed in Table3. The mass flow rate and the thermodynamic states of these equations can be calculated by the pinch point temperature difference (PPTD) method [12]. Their calculation programs are achieved in the MATLAB environment.Table 3
Thermodynamic evaluation equations [24].
DevicesEquationExpanderPexp=m˙wf·hexp,in−hexp,out,is·ηexpPumpPp=m˙wf·hp,in−hp,out,isηpHeat exchangerQheater=m˙wf·hheater,in−hheater,outSystemPp=Pexp−Pp,Qa=∑Qheater,η1st=PnetQa


### 3.3. Split Model

The split ratio (SR) in SR-ORC, SC-ORC, and HTS-CTRC is defined as the ratio between the regenerator stream mass flow rate and overall mass flow rate of systems. Each mass flow rate is described as follows [33]:
(2)m.orc,reg=SRorc·morc,m.orc,cac=1−SRorc·morc,(3)m.co2,reg=SRco2·mco2,m.co2,LT‐heater=1−SRco2·mco2.The convergence enthalpy can be calculated according to the energy conservation law as follows [33]:
(4)hM=SRorc·hreg,out+1−SRorc·hcac,out,(5)h4=SRco2·hreg,out+1−SRco2·hLT‐heater,out.


### 3.4. Model Validation

The four ORC configuration models proposed in this paper are validated by their closet configurations, that is, the R-ORC [18], the C-ORC [50], the D-ORC [29], and the HTS-CTRC [33]. The model assumption, such as the isentropic efficiencies of the expander and pump as well as the heat source conditions, are adjusted according to the references. As is shown in Table 4, good agreement is achieved between the current calculation and reference ones.Table 4
Comparison of the results of this paper with the calculations in Ref.
CycleNet power (kW)Thermal efficiencyRefPresentRD (%)RefPresentRD (%)R-ORC4.434.410.4%0.0990.0981%C-ORC18.8918.940.2%0.21110.21170.1%D-ORC29.028.950.1%0.11670.11640.2%HTS-CTRC18.8018.810.05%0.10830.10830%


## 3.1. Assumptions

Some reasonable assumptions before establishing the four ORC configuration models are as follows:(i)
Ambient temperature and pressure are 20°C and 101.3 kPa, respectively(ii)
All components are modelled for steady-state conditions [33](iii)
Heat losses and pressure drops as well as split and convergence losses are ignored [33](iv)
The isentropic efficiencies of the expander and pump are supposed to be 0.7 and 0.8, respectively(v)
The minimum superheat degree on the ORC configuration is set at 5 K to protect the expander(vi)
The pinch point temperature difference of the heat exchanger considering the exhaust gas, charge air, jacket water, and regenerator is set at 30 K,30 K,5 K and 10 K, respectively(vii)
The condenser subcooling is 1 K(viii)
The composition of exhaust gas is 12%CO2, 4%H2O, 73%N2, and 11%O2 while the charge air is 5%CO2, 75%N2, and 24.5%O2(ix)
The acid dew point of the exhaust gas is 85°C based on the model of [49]


## 3.2. Energy Model

Since four ORC systems have the same base thermodynamic processes, the energy balance equations of their components and the system efficiency equation based on the first law of thermodynamics can be expressed in a unified form as listed in Table3. The mass flow rate and the thermodynamic states of these equations can be calculated by the pinch point temperature difference (PPTD) method [12]. Their calculation programs are achieved in the MATLAB environment.Table 3
Thermodynamic evaluation equations [24].
DevicesEquationExpanderPexp=m˙wf·hexp,in−hexp,out,is·ηexpPumpPp=m˙wf·hp,in−hp,out,isηpHeat exchangerQheater=m˙wf·hheater,in−hheater,outSystemPp=Pexp−Pp,Qa=∑Qheater,η1st=PnetQa


## 3.3. Split Model

The split ratio (SR) in SR-ORC, SC-ORC, and HTS-CTRC is defined as the ratio between the regenerator stream mass flow rate and overall mass flow rate of systems. Each mass flow rate is described as follows [33]:
(2)m.orc,reg=SRorc·morc,m.orc,cac=1−SRorc·morc,(3)m.co2,reg=SRco2·mco2,m.co2,LT‐heater=1−SRco2·mco2.The convergence enthalpy can be calculated according to the energy conservation law as follows [33]:
(4)hM=SRorc·hreg,out+1−SRorc·hcac,out,(5)h4=SRco2·hreg,out+1−SRco2·hLT‐heater,out.


## 3.4. Model Validation

The four ORC configuration models proposed in this paper are validated by their closet configurations, that is, the R-ORC [18], the C-ORC [50], the D-ORC [29], and the HTS-CTRC [33]. The model assumption, such as the isentropic efficiencies of the expander and pump as well as the heat source conditions, are adjusted according to the references. As is shown in Table 4, good agreement is achieved between the current calculation and reference ones.Table 4
Comparison of the results of this paper with the calculations in Ref.
CycleNet power (kW)Thermal efficiencyRefPresentRD (%)RefPresentRD (%)R-ORC4.434.410.4%0.0990.0981%C-ORC18.8918.940.2%0.21110.21170.1%D-ORC29.028.950.1%0.11670.11640.2%HTS-CTRC18.8018.810.05%0.10830.10830%


## 4. Results and Discussion

In this part, the net power optimization analyses of four configurations under the original waste heat condition are first present in Section4.1, for which those optimization methods are the foundation to analyze the performance of four configurations under different waste heat conditions in Section 4.2. The condensation temperature of all configurations is set to be 25°C, higher than the ambient temperature we have assumed.


### 4.1. Optimization Analysis of Four ORC Configurations

In this section, the net power optimization of four ORC configurations under the original waste heat source is carried out. Each configuration has its own parameters which should be optimized.For the SR-ORC, there are three parameters that could be optimized, the split ratio (SR), the evaporating pressure (EP), and the superheat degree (SD). Due to the pinch limitation, the outlet on the working fluid side of the regenerator and CAC has a different outlet enthalpy depending on the split ratio. As is shown in Figure6(a), the regenerator outlet enthalpy is always higher than that of CAC because the regenerator gains more heat from the condensing process. At small split ratios, despite the acquired high enthalpy of the regenerator outlet, the mass flow rate of the regenerator is relatively tiny causing low regenerative heat usage. In the same way, low waste heat usage of CA appears at high split ratios. Therefore, there is an optimal split ratio that obtains the highest convergence enthalpy leading to the highest mass flow rate and net power. With regard to the EP and SD, the net power of SR-ORC shows the same trend of increase followed by decrease as is displayed in Figure 7(a). The split ratio for each point on this map has been optimized. With the increase in EP, the net power first increases and then decreases sharply because when the evaporating temperature which depends on EP is higher than that of JW, although a higher pressure ratio and thermal efficiency are obtained, a huge part of JW waste heat could not be utilized. On the other side, when the effect of mass reduction is greater than the increase in thermal efficiency, the increase of SD would undermine the performance of SR-ORC. Finally, under the reference heat source condition, the maximum net power of SR-ORC is 18.33 kW with 914 kPa EP and 72°C SD.Figure 6
The distribution of the split and convergence enthalpy: (a) SR-ORC; (b) HTS-CTRC.
(a)(b)Figure 7
Contour maps of four types of ORC configuration net power under reference waste heat condition: (a) SR-ORC, (b) SC-ORC, (c) D-ORC, and (d) HTS-CTRC.
(a)(b)(c)(d)As is discussed in Section2.2.2, the SC-ORC can be modified by SR-ORC by extracting the high-temperature part of EG to construct the HT-loop, so the LT-loop optimization methods are the same as those of the SR-ORC when the residual heat of the HT-loop and EG is determined. Therefore, the first step is to identify the influence of LT-loop parameters. As is shown in Figure 8, since the HT-loop is a basic ORC without the regenerator, better performance occurs at high EP but worse at high SD. Following the energy conservation law, the residual heat of exhaust gas plus HT-loop appears as an opposite pattern to net power. Consequently, an increase or decrease of net power of the HT-loop is accompanied by a decrease or increase of that of the LT-loop. Furthermore, the condensing pressure of the HT-loop has an opposite effect on the HT-loop and LT-loop as is shown in Figure 9. When the condensing pressure of the HT-loop rises, the HT-loop net power decreases while the residual heat of the HT-loop increases, thus improving the LT-loop performance. At the same time, the EP of the LT-loop rises along with the HT-loop condensing pressure causing the net power of the LT-loop to increase rapidly before reaching the EP of the waste heat of JW which could not be fully utilized. Therefore, there is an optimized condensing pressure of the HT-loop. After optimizing the above parameters, the SC-ORC performance depends on the set of the EP and SD of the HT-loop. Surprisingly, the net power of SC-ORC keeps increasing in a wide range of EP and SD as is shown in Figure 7(b). Such a phenomenon can be explained in terms of the waste heat source matching that when the EP and SD of the HT-loop increase; the HT-loop could better match the high-temperature part of the EG. Meanwhile, the LT-loop adjusts its SD to better match the residual heat of the EG and LT-loop. However, if the EP and SD of the HT-loop are too high, the net power out of SC-ORC will decrease as most of the EG waste heat is allocated to the LT-loop. In the extreme case, the SC-ORC will degenerate into SR-ORC once the LT-loop uses all the EG waste heat. For safe operation and practical application in this paper, the max EP and temperature are set at 3100 kPa and 240°C, respectively, which can generate 21.14 kW in SC-ORC.Figure 8
The HT-loop properties of SC-ORC in different EP and SD: (a) net power; (b) the residual heat of exhaust gas plus HT-loop.
(a)(b)Figure 9
The condensing pressure of HT-loop influence on SC-ORC.For the D-ORC, the HT-loop and LT-loop are relatively separated as those mass flow rates depend on their own waste heat source sizes. Since the LT-loop does not utilize the HT waste heat source, the EP of the LT-loop is set much lower compared to that of SR-ORC and SC-ORC to leave enough space ensuring that the working fluid will be in the overheated state. The optimized parameters of the LT-loop are completely dependent on the waste heat condition of JW and CA. In this case, the optimized EP and SD of the LT-loop are 888 kPa and 5°C, respectively. After optimizing the LT-loop parameters, Figure7(c) displays the net power of D-ORC under different EP and SD of the HT-loop. It could be found that each EP has an optimum SD maximizing the net power. However, there is no pinch point limitation at the point 3H (see T-s diagram) so that most EG waste heat could still be recovered under high EP and get a high net power. Similarly, the max EP of the HT-loop is set at 3100 kPa corresponding to 28°C optimized SD which could convert waste heat into 18.27 kW net power.For the HTS-CTRC, the parameters that should be discussed are the split ratio, turbine inlet pressure, and turbine inlet temperature. Similar to the SR-ORC, there is an optimized split ratio with the highest convergence enthalpy by the LT-gas heater and regenerator outlet that can obtain the highest net power as is shown in Figure6(b). It is noteworthy that the optimized split ratio occurs when both of the LT-gas heater and regenerator outlet enthalpy and temperature are the same so that there is no heat transfer loss during the convergence process in practical applications. The same result could be found in [33]. Figure 7(d) depicts the effect of the turbine inlet pressure and temperature under the optimized split ratio. As the turbine inlet pressure increases initially, the enthalpy of the turbine inlet increases, dominating the net power increase. Afterwards, the pump consumption increases and the mass flow rate decreases, resulting in the net power decrease. The same pattern occurs when increasing the turbine inlet temperature. Therefore, maximum net power of HTS-CTRC could be obtained at a certain turbine inlet pressure and temperature. Under the reference heat source condition, HTS-CTRC generates a maximum net power of 12.32 kW when the turbine inlet pressure and temperature are 11,900 kPa and 210°C, respectively.In summary, under the reference waste heat condition, the thermodynamic performance ranking of four configurations is SC-ORC, SR-ORC, D-ORC, and HTS-CTRC, corresponding to 21.14 kW, 18.33 kW, 18.27 kW, and 12.32 kW, respectively. Figure10 shows the total amount of waste heat utilized by the different configurations. It could be found that all configurations are capable of fully absorbing LT waste heat sources. However, only SC-ORC thoroughly utilizes the EG waste heat; the other configurations fail to fully absorb the EG waste heat as per the pinch limitation shown in the T-s diagrams of the four configurations, which is a factor causing the poorer performance of these configurations. Therefore, their performance may vary with the waste heat source, particularly the EG. The next section will continue to study their influence.Figure 10
The waste heat utilization in different ORC configurations.


### 4.2. Matching Analysis for Different Waste Heat Characteristics

In this section, the performance of the four configurations under different waste heat source conditions is carried out. The waste heat source condition is characterized by the temperature of EG and heat ratio as is discussed in Section2.1. It is noteworthy that whatever the increase of the temperature of EG or heat ratio, it is a conversion process from low-grade waste heat to high-grade waste heat. Compared to the low-grade waste heat, the high-grade waste heat is more recoverable because a higher heat absorption temperature could be matched leading to higher thermal efficiency.Figure11 shows the net power of the four configurations under different EG temperatures and heat ratios. All of them adjust the operating parameters to match different heat conditions based on the optimization method mentioned in Section 4.1. As is discussed above, part of the waste heat from EG is unable to be completely absorbed as per the pinch limitation except for the SC-ORC. Therefore, there is less unused EG waste heat leading to better performance when the temperature of EG is increased with a higher optimized SD of SR-ORC and D-ORC as well as a higher optimized expander inlet temperature and pressure of HTS-CTRC. An improvement also appears in the SC-ORC with the increase of the EG temperature because more waste heat from EG is distributed to the HT-loop which possesses a high thermal efficiency. However, when the temperature of EG exceeds 440°C, the HT-loop of SC-ORC gets rid of the pinch point limitation, allowing it to fully absorb the waste heat from the EG. As a result, the LT-loop of SC-ORC no longer obtains the waste heat from EG and is heat exchanger can be eliminated, which means the structure of SC-ORC changes according to different heat source conditions. In this case, increasing the temperature of the exhaust gas does not improve the performance of SC-ORC as long as the maximum pressure and temperature limitations remain at 3100 kPa and 240°C, respectively. On the other hand, when considering increasing the heat ratio, the HT-loop of D-ORC and SC-ORC obtains more waste heat, and the optimized SD of SR-ORC as well as the optimized expander pressure and temperature of HTS-CTRC gets higher; all this enables them to perform better. It is noteworthy that the optimized EP and SD of SC-ORC occurs under the heat source conditions with low EG temperature and high heat ratio, and a negative impact emerges in ORC slowing down the increase rate when the temperature of EG is low corresponding to the low temperature of CA causing part of the CA waste heat to be unutilized, especially the SR-ORC in the high heat ratio.Figure 11
Contour map of four RC configurations under different EG temperatures and heat ratios: (a) SR-ORC, (b) SC-ORC, (c) D-ORC, and (d) HTS-CTRC.
(a)(b)(c)(d)Finally, the HTS-CTRC is the most sensitive configuration with a huge 183.58% improvement when the EG temperature and heat ratio rise from 300°C, 0.5, to 650°C, 1.5, respectively, while the SC-ORC, D-ORC, and SR-ORC are 33.90%, 23.71%, and 17.41%, respectively. To facilitate comparison of the maximum net power achievable for all the considered conditions among these configurations, further selection maps are performed by combining four configurations’ performance maps. Since the net power of SC-ORC is always higher than that of SR-ORC and D-ORC, the first comparison is carried out between the SC-ORC and HTS-CTRC. Figure12(a) shows that SC-ORC has better performance than HTS-CTRC in a wide range of waste heat source conditions except extreme high EG temperature and heat ratio. It is reasonable because the SC-ORC is the most complex configuration with more segmented matching of multiple-grade waste heat from the engine. Therefore, the SC-ORC is an attractive choice if we are not concerned with the security, complexity, and cost of the WHR system. The second comparison is between the SR-ORC, D-ORC, and HT-CTRC, who have increasing complexity to their structures. As is displayed in Figure 12(b), the SR-ORC proves to be preferable to the HTS-CTRC and D-ORC under low heat ratio waste heat characteristics which mostly appears in the engine’s low load and speed operating conditions. With regard to the medium-high heat ratio, the HTS-CTRC is better matched to the high EG temperature while D-ORC is suitable for the low EG temperature. Therefore, the HTS-CTRC is suited to the operating conditions of an engine with a high heat ratio, especially those with high EG temperatures, such as the gasoline engine, natural gas engine, and hydrogen engine, while the D-ORC is a match to the waste heat characteristics of the low EG temperature and high heat ratio such as the diesel engine operating at a high load and speed. The features of the four configurations are summarized in Table 5.Figure 12
The selection maps of four ORC configurations referring to different EG temperatures and heat ratios: (a) SC-ORC vs. HTS-ORC; (b) SR-ORC vs. D-ORC vs. HTS-CTRC.
(a)(b)Table 5
The feature of four ORC configurations for engine waste heat recovery.
WHR system configurationMatching characteristic of waste heat conditionAdvantages and disadvantagesSR-ORCLow heat ratioLess components but poor performance in high heat ratioSC-ORCMostly conditionsLarge output power but more components and difficult to controlD-ORCHigh heat ratio with low EG temperatureLarge output power, easy to control, but poor performance in low EG temperature and heat ratioHTS-CTRCHigh heat ratio with high EG temperatureLarge output power, environmentally friendly, and small component volume but poor performance in low EG temperature and heat ratioThe above comparison is based on the fixed temperature of JW. In fact, Figure13 shows that an enhancement in ORC can be obtained by increasing the JW temperature corresponding to a higher EP, but their increase rates are similar. On the other hand, small temperature variations hardly affect the optimum turbine pressure and temperature of the HTS-CTRC when the waste heat amount of JW remains unchanged, so the HTS-CTRC net power is constant and its area occupied in the selection map would be slightly reduced.Figure 13
The influence of jacket water temperature on four configurations.


## 4.1. Optimization Analysis of Four ORC Configurations

In this section, the net power optimization of four ORC configurations under the original waste heat source is carried out. Each configuration has its own parameters which should be optimized.For the SR-ORC, there are three parameters that could be optimized, the split ratio (SR), the evaporating pressure (EP), and the superheat degree (SD). Due to the pinch limitation, the outlet on the working fluid side of the regenerator and CAC has a different outlet enthalpy depending on the split ratio. As is shown in Figure6(a), the regenerator outlet enthalpy is always higher than that of CAC because the regenerator gains more heat from the condensing process. At small split ratios, despite the acquired high enthalpy of the regenerator outlet, the mass flow rate of the regenerator is relatively tiny causing low regenerative heat usage. In the same way, low waste heat usage of CA appears at high split ratios. Therefore, there is an optimal split ratio that obtains the highest convergence enthalpy leading to the highest mass flow rate and net power. With regard to the EP and SD, the net power of SR-ORC shows the same trend of increase followed by decrease as is displayed in Figure 7(a). The split ratio for each point on this map has been optimized. With the increase in EP, the net power first increases and then decreases sharply because when the evaporating temperature which depends on EP is higher than that of JW, although a higher pressure ratio and thermal efficiency are obtained, a huge part of JW waste heat could not be utilized. On the other side, when the effect of mass reduction is greater than the increase in thermal efficiency, the increase of SD would undermine the performance of SR-ORC. Finally, under the reference heat source condition, the maximum net power of SR-ORC is 18.33 kW with 914 kPa EP and 72°C SD.Figure 6
The distribution of the split and convergence enthalpy: (a) SR-ORC; (b) HTS-CTRC.
(a)(b)Figure 7
Contour maps of four types of ORC configuration net power under reference waste heat condition: (a) SR-ORC, (b) SC-ORC, (c) D-ORC, and (d) HTS-CTRC.
(a)(b)(c)(d)As is discussed in Section2.2.2, the SC-ORC can be modified by SR-ORC by extracting the high-temperature part of EG to construct the HT-loop, so the LT-loop optimization methods are the same as those of the SR-ORC when the residual heat of the HT-loop and EG is determined. Therefore, the first step is to identify the influence of LT-loop parameters. As is shown in Figure 8, since the HT-loop is a basic ORC without the regenerator, better performance occurs at high EP but worse at high SD. Following the energy conservation law, the residual heat of exhaust gas plus HT-loop appears as an opposite pattern to net power. Consequently, an increase or decrease of net power of the HT-loop is accompanied by a decrease or increase of that of the LT-loop. Furthermore, the condensing pressure of the HT-loop has an opposite effect on the HT-loop and LT-loop as is shown in Figure 9. When the condensing pressure of the HT-loop rises, the HT-loop net power decreases while the residual heat of the HT-loop increases, thus improving the LT-loop performance. At the same time, the EP of the LT-loop rises along with the HT-loop condensing pressure causing the net power of the LT-loop to increase rapidly before reaching the EP of the waste heat of JW which could not be fully utilized. Therefore, there is an optimized condensing pressure of the HT-loop. After optimizing the above parameters, the SC-ORC performance depends on the set of the EP and SD of the HT-loop. Surprisingly, the net power of SC-ORC keeps increasing in a wide range of EP and SD as is shown in Figure 7(b). Such a phenomenon can be explained in terms of the waste heat source matching that when the EP and SD of the HT-loop increase; the HT-loop could better match the high-temperature part of the EG. Meanwhile, the LT-loop adjusts its SD to better match the residual heat of the EG and LT-loop. However, if the EP and SD of the HT-loop are too high, the net power out of SC-ORC will decrease as most of the EG waste heat is allocated to the LT-loop. In the extreme case, the SC-ORC will degenerate into SR-ORC once the LT-loop uses all the EG waste heat. For safe operation and practical application in this paper, the max EP and temperature are set at 3100 kPa and 240°C, respectively, which can generate 21.14 kW in SC-ORC.Figure 8
The HT-loop properties of SC-ORC in different EP and SD: (a) net power; (b) the residual heat of exhaust gas plus HT-loop.
(a)(b)Figure 9
The condensing pressure of HT-loop influence on SC-ORC.For the D-ORC, the HT-loop and LT-loop are relatively separated as those mass flow rates depend on their own waste heat source sizes. Since the LT-loop does not utilize the HT waste heat source, the EP of the LT-loop is set much lower compared to that of SR-ORC and SC-ORC to leave enough space ensuring that the working fluid will be in the overheated state. The optimized parameters of the LT-loop are completely dependent on the waste heat condition of JW and CA. In this case, the optimized EP and SD of the LT-loop are 888 kPa and 5°C, respectively. After optimizing the LT-loop parameters, Figure7(c) displays the net power of D-ORC under different EP and SD of the HT-loop. It could be found that each EP has an optimum SD maximizing the net power. However, there is no pinch point limitation at the point 3H (see T-s diagram) so that most EG waste heat could still be recovered under high EP and get a high net power. Similarly, the max EP of the HT-loop is set at 3100 kPa corresponding to 28°C optimized SD which could convert waste heat into 18.27 kW net power.For the HTS-CTRC, the parameters that should be discussed are the split ratio, turbine inlet pressure, and turbine inlet temperature. Similar to the SR-ORC, there is an optimized split ratio with the highest convergence enthalpy by the LT-gas heater and regenerator outlet that can obtain the highest net power as is shown in Figure6(b). It is noteworthy that the optimized split ratio occurs when both of the LT-gas heater and regenerator outlet enthalpy and temperature are the same so that there is no heat transfer loss during the convergence process in practical applications. The same result could be found in [33]. Figure 7(d) depicts the effect of the turbine inlet pressure and temperature under the optimized split ratio. As the turbine inlet pressure increases initially, the enthalpy of the turbine inlet increases, dominating the net power increase. Afterwards, the pump consumption increases and the mass flow rate decreases, resulting in the net power decrease. The same pattern occurs when increasing the turbine inlet temperature. Therefore, maximum net power of HTS-CTRC could be obtained at a certain turbine inlet pressure and temperature. Under the reference heat source condition, HTS-CTRC generates a maximum net power of 12.32 kW when the turbine inlet pressure and temperature are 11,900 kPa and 210°C, respectively.In summary, under the reference waste heat condition, the thermodynamic performance ranking of four configurations is SC-ORC, SR-ORC, D-ORC, and HTS-CTRC, corresponding to 21.14 kW, 18.33 kW, 18.27 kW, and 12.32 kW, respectively. Figure10 shows the total amount of waste heat utilized by the different configurations. It could be found that all configurations are capable of fully absorbing LT waste heat sources. However, only SC-ORC thoroughly utilizes the EG waste heat; the other configurations fail to fully absorb the EG waste heat as per the pinch limitation shown in the T-s diagrams of the four configurations, which is a factor causing the poorer performance of these configurations. Therefore, their performance may vary with the waste heat source, particularly the EG. The next section will continue to study their influence.Figure 10
The waste heat utilization in different ORC configurations.


## 4.2. Matching Analysis for Different Waste Heat Characteristics

In this section, the performance of the four configurations under different waste heat source conditions is carried out. The waste heat source condition is characterized by the temperature of EG and heat ratio as is discussed in Section2.1. It is noteworthy that whatever the increase of the temperature of EG or heat ratio, it is a conversion process from low-grade waste heat to high-grade waste heat. Compared to the low-grade waste heat, the high-grade waste heat is more recoverable because a higher heat absorption temperature could be matched leading to higher thermal efficiency.Figure11 shows the net power of the four configurations under different EG temperatures and heat ratios. All of them adjust the operating parameters to match different heat conditions based on the optimization method mentioned in Section 4.1. As is discussed above, part of the waste heat from EG is unable to be completely absorbed as per the pinch limitation except for the SC-ORC. Therefore, there is less unused EG waste heat leading to better performance when the temperature of EG is increased with a higher optimized SD of SR-ORC and D-ORC as well as a higher optimized expander inlet temperature and pressure of HTS-CTRC. An improvement also appears in the SC-ORC with the increase of the EG temperature because more waste heat from EG is distributed to the HT-loop which possesses a high thermal efficiency. However, when the temperature of EG exceeds 440°C, the HT-loop of SC-ORC gets rid of the pinch point limitation, allowing it to fully absorb the waste heat from the EG. As a result, the LT-loop of SC-ORC no longer obtains the waste heat from EG and is heat exchanger can be eliminated, which means the structure of SC-ORC changes according to different heat source conditions. In this case, increasing the temperature of the exhaust gas does not improve the performance of SC-ORC as long as the maximum pressure and temperature limitations remain at 3100 kPa and 240°C, respectively. On the other hand, when considering increasing the heat ratio, the HT-loop of D-ORC and SC-ORC obtains more waste heat, and the optimized SD of SR-ORC as well as the optimized expander pressure and temperature of HTS-CTRC gets higher; all this enables them to perform better. It is noteworthy that the optimized EP and SD of SC-ORC occurs under the heat source conditions with low EG temperature and high heat ratio, and a negative impact emerges in ORC slowing down the increase rate when the temperature of EG is low corresponding to the low temperature of CA causing part of the CA waste heat to be unutilized, especially the SR-ORC in the high heat ratio.Figure 11
Contour map of four RC configurations under different EG temperatures and heat ratios: (a) SR-ORC, (b) SC-ORC, (c) D-ORC, and (d) HTS-CTRC.
(a)(b)(c)(d)Finally, the HTS-CTRC is the most sensitive configuration with a huge 183.58% improvement when the EG temperature and heat ratio rise from 300°C, 0.5, to 650°C, 1.5, respectively, while the SC-ORC, D-ORC, and SR-ORC are 33.90%, 23.71%, and 17.41%, respectively. To facilitate comparison of the maximum net power achievable for all the considered conditions among these configurations, further selection maps are performed by combining four configurations’ performance maps. Since the net power of SC-ORC is always higher than that of SR-ORC and D-ORC, the first comparison is carried out between the SC-ORC and HTS-CTRC. Figure12(a) shows that SC-ORC has better performance than HTS-CTRC in a wide range of waste heat source conditions except extreme high EG temperature and heat ratio. It is reasonable because the SC-ORC is the most complex configuration with more segmented matching of multiple-grade waste heat from the engine. Therefore, the SC-ORC is an attractive choice if we are not concerned with the security, complexity, and cost of the WHR system. The second comparison is between the SR-ORC, D-ORC, and HT-CTRC, who have increasing complexity to their structures. As is displayed in Figure 12(b), the SR-ORC proves to be preferable to the HTS-CTRC and D-ORC under low heat ratio waste heat characteristics which mostly appears in the engine’s low load and speed operating conditions. With regard to the medium-high heat ratio, the HTS-CTRC is better matched to the high EG temperature while D-ORC is suitable for the low EG temperature. Therefore, the HTS-CTRC is suited to the operating conditions of an engine with a high heat ratio, especially those with high EG temperatures, such as the gasoline engine, natural gas engine, and hydrogen engine, while the D-ORC is a match to the waste heat characteristics of the low EG temperature and high heat ratio such as the diesel engine operating at a high load and speed. The features of the four configurations are summarized in Table 5.Figure 12
The selection maps of four ORC configurations referring to different EG temperatures and heat ratios: (a) SC-ORC vs. HTS-ORC; (b) SR-ORC vs. D-ORC vs. HTS-CTRC.
(a)(b)Table 5
The feature of four ORC configurations for engine waste heat recovery.
WHR system configurationMatching characteristic of waste heat conditionAdvantages and disadvantagesSR-ORCLow heat ratioLess components but poor performance in high heat ratioSC-ORCMostly conditionsLarge output power but more components and difficult to controlD-ORCHigh heat ratio with low EG temperatureLarge output power, easy to control, but poor performance in low EG temperature and heat ratioHTS-CTRCHigh heat ratio with high EG temperatureLarge output power, environmentally friendly, and small component volume but poor performance in low EG temperature and heat ratioThe above comparison is based on the fixed temperature of JW. In fact, Figure13 shows that an enhancement in ORC can be obtained by increasing the JW temperature corresponding to a higher EP, but their increase rates are similar. On the other hand, small temperature variations hardly affect the optimum turbine pressure and temperature of the HTS-CTRC when the waste heat amount of JW remains unchanged, so the HTS-CTRC net power is constant and its area occupied in the selection map would be slightly reduced.Figure 13
The influence of jacket water temperature on four configurations.


## 5. Conclusions

This paper focuses on analyzing the thermodynamic performance of four ORC configurations including the SR-ORC, SC-ORC, D-ORC, and HTS-CTRC under different waste heat conditions of the engine. The engine waste heat sources are characterized by the EG temperature and heat ratio. Each configuration optimizes their net power according to the different waste heat characteristics. Two selection maps have been developed, which could be applied universally to rapidly select the most appropriate ORC configuration based on varying waste heat characteristics, regardless of engine type and operating conditions. The main conclusions are summarized as follows:(1)
The maximum net power of SR-ORC and HTS-CTRC could be obtained by optimizing their operating parameters at any waste heat source conditions, while the D-ORC and SC-ORC have a great potential by increasing their HT-loop EP and SD (only SC-ORC) over a wide range of waste heat conditions(2)
The operating parameters of the four configurations adjust to match different EG temperatures and heat ratios. Increasing the EG temperature and heat ratio can improve their performance. The HTS-CTRC is the most improved configuration with the net power increased by 183.58% when the EG temperature and heat ratio rise from 300°C, 0.5, to 650°C, 1.5, respectively, while the SR-ORC, D-ORC, and SC-ORC increased by 17.41%, 23.71%, and 35.45%, respectively(3)
Two selection maps indicate different ORC configurations suitable for different waste heat conditions. Without considering the system complexity and high cost, the SC-ORC is an attractive choice for most waste heat conditions. The HTS-CTRC is suitable for the waste heat condition of the high EG temperature and heat ratio, such as from gasoline, natural gas, and hydrogen fuel engines, while the D-ORC is the optimum for the low EG temperature and high heat ratio, such as from diesel engines at high operating load and speed. The SR-ORC is appropriate for waste heat conditions with a low heat ratio, corresponding to the low operating load and speed of any engine



---

*Source: 8865282-2023-03-13.xml*