Applied Energy xxx (2012) xxx xxx. Contents lists available at SciVerse ScienceDirect. Applied Energy

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1 Applied Energy xxx (212) xxx xxx Contents lists available at SciVerse ScienceDirect Applied Energy journal homepage: A performance analysis of a novel system of a dual loop bottoming organic Rankine cycle (ORC) with a light-duty diesel engine H.G. Zhang, E.H. Wang, B.Y. Fan College of Environmental and Energy Engineering, Beijing University of Technology, Pingleyuan No., 124 Beijing, China highlights " The waste heat characteristic of a light duty diesel engine is analyzed. " A dual loop ORC is designed to simultaneously recover waste heat of exhaust, intake air and coolant. " Effective power and bsfc of combined system is improved greatly over engine s operating region. article info abstract Article history: Received 17 June 212 Received in revised form 5 September 212 Accepted 9 September 212 Available online xxxx Keywords: Organic Rankine cycle Waste heat recovery Diesel engine Performance MAP Dual loop A small-scale organic Rankine cycle (ORC) can be used to harness the waste heat from an internal combustion engine. In this paper, the characteristic of a novel system combining a vehicular light-duty diesel engine with a dual loop ORC, which recovers waste heat from the engine exhaust, intake air, and coolant, is analyzed. A high temperature loop recovers the exhaust heat, whereas a low temperature loop recovers the residual heat from the high temperature loop and the waste heat from both the intake air and the coolant. A performance map of the light-duty diesel engine is created using an engine test bench. The heat waste from the exhaust, the intake air, and the coolant are calculated and compared throughout the engine s entire operating region. Based on these data, the working parameters of the dual loop ORC are defined, and the performance of the combined engine ORC system is evaluated across this entire region. The results show that the net power of the low temperature loop is higher than that of the high temperature loop, and the relative output power improves from 14% to 16% in the peak effective thermal efficiency region and from 38% to 43% in the small load region. In addition, the brake specific fuel consumption (bsfc) of the combined system decreases significantly throughout the engine s operating region. Ó 212 Elsevier Ltd. All rights reserved. 1. Introduction Huge amounts of energy are consumed by internal combustion engines in all types of vehicles, with much of this energy is wasted through the exhaust, the intake air, and the cooling systems. Exacerbating this problem is the fact that these combustion products also cause serious environmental issues. Engine waste-heat recovery could improve the fuel thermal efficiency, minimize fuel consumption, and reduce engine emissions. Using an organic Rankine cycle (ORC) to recover the low-grade wasted heat from these systems is the technology that is the closest to being suitable for mass production. When designing an ORC, special attention must give to the choice of the working fluid and the design of a suitable expander [1 7]. Many researchers have investigated ORC system design and parametric optimization. The dynamic performance and control strategy was investigated by Ref. [8] using a Corresponding author. Tel.: ; fax: address: zhanghongguang@bjut.edu.cn (H.G. Zhang). time-varying model. The results indicate that the steady-state optimization of ORC under various conditions is very important. The parameter optimization and performance comparison was also conducted by Ref. [9] for low-temperature heat source (8 C). When an engine is running, the energy and exergy quantities of the exhaust, the intake air, and the coolant are significantly different. Because of these differences, it is very difficult to design a system that can recover waste heat from all of these systems. Previous investigations have been conducted to solve this problem for various engines [1 15]. However, few of these investigations have concentrated on light-duty diesel engine applications. In this paper, a dual loop ORC system is designed, combining a high temperature (HT) loop and a low temperature (LT) loop to simultaneously recover the waste heat from the exhaust, the intake air, and the coolant of a light-duty diesel engine. The HT loop recovers the exhaust heat, whereas the LT loop recovers the residual heat from the HT loop and the waste heat from both the intake air and the coolant. The two separate loops are coupled through a pre-heater. To evaluate the dual loop system performance when /$ - see front matter Ó 212 Elsevier Ltd. All rights reserved. light-duty diesel engine. Appl Energy (212),

2 2 H.G. Zhang et al. / Applied Energy xxx (212) xxx xxx Nomenclature _W power (kw) _m mass flow rate (kg/s) h enthalpy (kj/kg) s entropy (kj/kg K) _I exergy destruction rate (kw) T temperature (K) P pressure (MPa) _Q heat quantity (kw) mf mass fraction x, y molar amount Greek letters g efficiency Subscript cr critical point bp normal boiling point reference state HT1,HT2,HT2s,HT3,HT4,HT4s state points in HT loop LT1,LT2,LT2s,LT3,LTa,LT4,LTb,LT5,LT6,LT6s state points in LT loop p1 pump 1 p2 pump 2 exh exhaust gas in at the inlet out at the outlet e1 evaporator 1 e2 evaporator 2 s1 expander 1 s2 expander 2 pre pre-heater int intercooler cool coolant c condenser mc mean condensing temperature me mean evaporation temperature f fuel a intake air b brake i indicated misc miscellaneous n net cs combined system HT HT loop LT LT loop Acronyms ORC organic Rankine cycle HT high temperature LT low temperature ODP ozone depletion potential (relative to R11) GWP global warming potential (relative to CO 2 ) SUV sport utility vehicle bsfc brake specific fuel consumption combined with a light-duty diesel engine, the waste heat quantities are first calculated using engine test data. Based on these calculations, the working parameters for the HT and LT loops are determined. R245fa and R134a are selected as the working fluids for the HT loop and the LT loop, respectively. Finally, the performance map of the combined system is calculated and compared to a system with a non-bottoming ORC. 2. System design The waste heat generated by a light-duty diesel engine is found mainly in the exhaust, the intake air, and the coolant. The waste heat carried by the lubrication system can be added to the coolant waste heat if a water-cooled heat exchanger is used. The dual loop ORC designed for this study is shown in Fig. 1. The HT loop recovers the exhaust waste heat, while the LT loop is coupled to recover the residual heat of the HT loop, the waste heat of intake air in the intercooler, and the coolant waste heat. The HT loop consists of a pump (pump 1), an evaporator (evaporator 1), an expander (expander 1), the pre-heater, a reservoir (reservoir 1), and the associated connecting pipes. The LT loop consists of a pump (pump 2), the intercooler, the pre-heater, an evaporator (evaporator 2), an expander (expander 2), the condenser, a reservoir (reservoir 2), and the associated connecting pipes. The LT loop is coupled to the HT loop via the pre-heater, which is used as the condenser for the HT loop. Two single screw expanders were adopted here, which were invented by Beijing University of Technology, China [16,17]. The working fluid of the HT loop was chosen to be R245fa because of its good safety and environmental properties [18]. For the lowtemperature ORC, R134a was selected as the working fluid because of its appropriate critical temperature and pressure. R134a is also an environmentally friendly refrigerant with a zero ODP and a relatively low GWP value [19], widely used in automotive air-conditioners. The properties of these two working fluids are listed in Table 1. The working principle of the dual loop system is illustrated in Fig. 2. After the light-duty diesel engine warms up, the ORC system starts to recover the waste heat. The R245fa is pumped from reservoir 1 to evaporator 1, corresponding to the HT1 to HT2 process. The waste heat from the exhaust is then added, and the working fluid is evaporated to the saturated vapor state, HT3. Subsequently, the R245fa is expanded through expander 1, and the useful work out is used to generate electricity. R245fa is a dry working fluid, therefore, it changes to the superheated state, HT4, after expansion. Upon reaching the pre-heater, the R245fa is transformed into the saturated liquid state, HT1, after transferring its heat to the R134a working fluid. Later, the working fluid returns to reservoir 1 and waits for the next circulation cycle. Meanwhile, in the LT loop, pump 2 pressurizes the R134a from reservoir 2 in preparation to be sent to the intercooler. The corresponding process is shown as moving from LT1 to LT2 in Fig. 2. The R134a is heated to the sub-cooled state LT3 by the intake air in the intercooler. Subsequently, the R134a enters into the pre-heater and changes into the two-phase state, LT4. The coolant then flows out of the engine jacket and heats the R134a to the superheated state LT5 inside of evaporator 2. Overheating is required because R134a is a wet working fluid and overheating guarantees that no liquid is generated during the subsequent expansion process. The R134a remains in the slightly superheated state LT6 after undergoing an expansion process inside expander 2. Later, the fluid is condensed back to the saturated liquid state LT1 in the condenser before flowing back into reservoir 2. The saturation curves of R245fa and R134a are plotted in the T s diagram of Fig. 2. The upper red 1 lines correspond to the HT loop, while the lower blue lines show the LT loop. 1 For interpretation of color in Figs. 2 and 3, the reader is referred to the web version of this article. light-duty diesel engine. Appl Energy (212),

3 H.G. Zhang et al. / Applied Energy xxx (212) xxx xxx 3 Reservior2 LT1 Pump2 LT2 Intercooler LT3 T cool,in T air,in Fan T air,out Preheater T cool,out LT4 Evaporator2 Engine HT4 Expander1 Reservior1 HT1 LT5 Compressor Turbine HT3 Evaporator1 Pump1 HT2 T exh,in T exh,out LT6 Expander2 Intake path Exhaust path LT ORC circuit HT ORC circuit Coolant path Fig. 1. Schematic of a dual loop ORC system combined with a light-duty diesel engine. Table 1 Properties of the selected working fluids. Organic fluid Molecular weight (kg/kmol) T cr (K) P cr (MPa) T bp (K) ASHRAE 34 safety group Atmospheric life time (yr) ODP GWP ( yr) R245fa B R134a A Temperature (K) LT Entropy (kj/kg.k) 3. Mathematical model LT1 HT2 HT1 a LT3 R245fa The thermodynamic model for the system described in this paper was developed using the first and second law methods. In the HT loop, the process HT1 to HT2 is expressed as LT4 R134a Fig. 2. T s plots of the HT and LT loops. b HT3 HT4 LT5 LT6 _W p1 ¼ _m HT ðh HT2 h HT1 Þ¼ _m HT ðh HT2s h HT1 Þ=g p1 _I p1 ¼ T _m HT ðs HT2 s HT1 Þ The heat transfer process between the exhaust gas and the R245fa fluid in evaporator 1 is denoted as _Q exh ¼ _m exh ðh exh;in h exh;out Þ¼ _m HT ðh HT3 h HT2 Þ ð3þ The exergy destruction rate in evaporator 1 is calculated as _I e1 ¼ T _m HT ðs HT3 s HT2 ÞþT _m exh ðs exh;out s exh;in Þ When combustion flue gas is cooled in a heat recovery application, the temperature must not be allowed to drop below the acid dew point [2]. For this reason, it is desirable for the cooled exhaust temperature to be set above C. In this study, the exhaust temperature at the outlet of evaporator 1 is specified as 15 C. The exhaust temperature at the inlet of evaporator 1 is measured via an engine performance test. The enthalpy and entropy are calculated based on the components and temperature of the exhaust gas mixture. Thus, Eq. (3) provides the heat quantity transferred from the exhaust gas. The output work performed by expander 1 is calculated as _W s1 ¼ _m HT ðh HT3 h HT4 Þ¼ _m HT ðh HT3 h HT4s Þg s1 The exergy destruction rate of expander 1 is expressed as _I s1 ¼ T _m HT ðs HT4 s HT3 Þ ð1þ ð2þ ð4þ ð5þ ð6þ light-duty diesel engine. Appl Energy (212),

4 4 H.G. Zhang et al. / Applied Energy xxx (212) xxx xxx The heat rejection rate for the process HT4 to HT1 is denoted as _Q pre ¼ _m HT ðh HT4 h HT1 Þ In the LT loop, the work of pump 2 in the process from LT1 to LT2 is expressed as _W p2 ¼ _m LT ðh LT2 h LT1 Þ¼ _m LT ðh LT2s h LT1 Þ=g p2 ð8þ The exergy destruction rate of pump 2 is calculated as _I p2 ¼ T _m LT ðs LT2 s LT1 Þ The heat transfer in the intercooler is denoted as _Q int ¼ _m a ðh a;in h a;out Þ¼ _m LT ðh LT3 h LT2 Þ ð7þ ð9þ ð1þ The exergy destruction rate in the intercooler is represented as _I int ¼ T _m LT ðs LT3 s LT2 ÞþT _m a ðs a;out s a;in Þ ð11þ The heat added to the R134a in the pre-heater is calculated as _Q pre ¼ _m LT ðh LT4 h LT3 Þ¼ _m HT ðh HT4 h HT1 Þ The exergy destruction rate in the pre-heater is defined as _I pre ¼ T _m HT ðs HT1 s HT4 ÞþT _m LT ðs LT4 s LT3 Þ The heat added in evaporator 2 is calculated as _Q e2 ¼ _m LT ðh LT5 h LT4 Þ ð12þ ð13þ ð14þ Because the heat added from evaporator 2 is completely provided by the coolant waste heat, the following assumption holds true: _Q e2 ¼ Q _ cool ð15þ where Q _ cool is the heat rejected by the coolant, without considering the intercooler. Like the exhaust waste heat, this value can be measured via an engine performance test. If the mean temperature of the R134a in evaporator 2 is assumed to be equal to the evaporation temperature, the exergy destruction rate in evaporator 2 becomes _I e2 ¼ T _m LT ðs LT5 s LT4 Þ h LT5 h LT4 ð16þ T me To maintain a consistent heat transmission from the high temperature side to the low temperature side of the pre-heater, the temperature difference between the condensation temperature of the HT loop and the evaporation temperature of the LT loop is set to 5 K. For this analysis, this means that the working fluid temperature at state LT4 can be described as T LT4 ¼ T HT1 5 ð17þ To ensure the working fluid remains in a superheated state during the expansion process in expander 2, the temperature of the R134a at state LT5 is maintained at T LT5 ¼ T LTb þ 5 ð18þ where T LTb is the evaporation temperature of the R134a at the same pressure in state LT5. The output work generated by expander 2 is expressed as _W s2 ¼ _m LT ðh LT5 h LT6 Þ¼ _m LT ðh LT5 h LT6s Þg s2 The exergy destruction rate of expander 2 is expressed as _I s2 ¼ T _m LT ðs LT6 s LT5 Þ ð19þ ð2þ The heat rejected from the R134a working fluid in the condenser is calculated from _Q c ¼ _m LT ðh LT6 h LT1 Þ ð21þ If the mean temperature of the R134a in the condenser is assumed to be equal to the condensation temperature, the exergy destruction rate of the condenser becomes _I c ¼ T _m LT ðs LT1 s LT6 Þ h LT1 h LT6 T mc ð22þ Given the operating conditions of a dual loop ORC system mounted inside a vehicle, the following are the assumptions for the thermodynamic model used in this paper: (1) All the cycles are operated at steady state conditions, and pressure loss and heat rejection inside the pipes are ignored. (2) The inlet pressure for expander 1 is set to 2.4 MPa. (3) The condensation temperature of the R245fa is set to 75 C. The evaporation temperature of the R134a is 7 C, according to Eq. (18). Because the opening temperature of an engine thermostat valve is normally set above 85 C, the minimum temperature difference between the coolant and the R134a is greater than 1 C. Therefore, this temperature configuration is plausible when considering the temperature limitation at the pinch point. (4) The condensation temperature of the R134a is set to 2 C. (5) The air temperature at the intercooler outlet is set to 3 C. (6) The isentropic efficiencies of pump 1 and pump 2 are set to.8. These values are proper for a positive displacement pump, such as a plunger pump. (7) The isentropic efficiencies of expander 1 and expander 2 are set to.75. These efficiencies are a bit high, but can be achievable. The reason for choosing a high value is that will be helpful to analyze the potential maximum power generated by the ORC. 4. Engine waste heat evaluation To evaluate the dual loop ORC system performance, we first obtain the waste heat quantities of the exhaust, the intake air, and the coolant of the diesel engine. In this research, a four cylinder turbocharged diesel engine was selected for the case study. Table 2 lists the main specifications for this engine. When a vehicle is running, the engine speed and load can vary through a wide range. Therefore, the engine performance test is conducted in an engine test cell to obtain the thermodynamic parameters for the exhaust, the intake air, and the coolant system over all engine operating regions, as defined by the engine speed and output torque. The test procedure is performed according to Ref. [21]. For the measurements described here, the minimum and maximum engine speeds are set to r/min and 4 r/ min, respectively. The intermediate speeds are selected using a step increment of r/min. At each selected engine speed, eight different load values are selected, ranging from a % load to a minimum stable load value. Table 2 Specifications of the R425 diesel engine. Item Parameter Unit Model R425 Displacement L Bore stroke mm Cylinder number 4 Valve number per cylinder 4 Fuel injection equipment Common rail injection system Rated power 15 kw Rated speed 4 r/min Max. torque 34 N m Speed at max. torque 24 r/min light-duty diesel engine. Appl Energy (212),

5 H.G. Zhang et al. / Applied Energy xxx (212) xxx xxx 5 Brake specific fuel consumption (g/kw.h) Effective thermal efficiency (%) (a) 4 (b) Exhaust temperature (K) Exhaust mass flow rate (kg/s) (c) 4 (d) Fig. 3. Performance maps of the R425 diesel engine. One common way to present the operating characteristics of an internal combustion engine over its full load and speed range is to plot brake specific fuel consumption (bsfc) contours on a graph of brake mean effective pressure (or engine torque) versus engine speed. The measured engine performance map is displayed in Fig. 3a. The contours with green lines represent the measured engine power (kw). The color-filled contours with black lines denote the variable shown at the top of the figure. In the middle and high load regions, the bsfc remains at a low level, which is less than g/kw h. The lowest bsfc zone is situated in the high duty range between 18 r/min and 24 r/min, and the minimum bsfc value is 29.3 g/kw h. The measured bsfc data indicate the R425 diesel engine has good fuel economy. The maximum output torque the engine achieves is 34 N m between r/min and 26 r/min, which provides ample reserve torque and is suitable for an SUV. The effective thermal efficiency is defined as the ratio of the output torque at the flywheel end to the fuel combustion energy. The results of the calculations for the effective thermal efficiency are given in Fig. 3b. The effective thermal efficiency reaches a peak greater than 39% in the low bsfc region. Fig. 3c shows the measured exhaust temperature. The exhaust temperature increases slowly with the engine speed, but increases rapidly with the engine load because the amount of combustion energy available in the engine increases significantly due to the large quantity of fuel injected during a high engine load. At the rated power point, the exhaust temperature is 528 C. Fig. 3d shows the mass flow rate of the exhaust, which is the sum of the amounts of intake air and the injected fuel. It can be observed that the mass flow rate of the exhaust increases slowly with engine load, but rapidly with engine speed, which is because the increment of the engine load is primarily dependent on the increase in the injected fuel quantity, whereas the mass flow rate of the intake air essentially remains constant for stable engine speeds. The maximum mass flow rate is.1894 kg/s at the rated power point. The amount of waste heat from the light-duty diesel engine is then evaluated using the measured engine operating parameters. Eq. (23) describes the fuel combustion process according to the conservation of energy equation _m f h f þ _m a h a ¼ _ W i þ _ Q cool þ _ Q misc þ _m exh h exh ð23þ Here _m f and _m a are the fuel and air mass flow rates, respectively. The variables h f and h a are the corresponding inlet enthalpies. The variable W _ i is the indicated power of the engine, which can be calculated using the in-cylinder pressure [22]. In this study, the in-cylinder pressure is measured using a piezoelectric transducer (Kistler light-duty diesel engine. Appl Energy (212),

6 6 H.G. Zhang et al. / Applied Energy xxx (212) xxx xxx 655Bsp) connected to an amplifier (Kistler 11). The variable _ Q cool is the heat transferred to the cooling system, and _ Q misc is the miscellaneous heat loss due to convection and radiation from the engine block. Miscellaneous heat losses are ignored in this study because they normally account for only a small portion of the overall combustion energy. The combustion energy (defined as the enthalpy of the flammable mixture of gases) is calculated from the injected fuel quantity and the intake air mass. The exhaust enthalpy is calculated using an approximation method. Petroleum-derived diesel fuel is composed of approximately 75% saturated hydrocarbons and 25% aromatic hydrocarbons. Thus, the average chemical formula for common diesel fuel can be denoted by C 12 H 23 [23]. Therefore, the combustion process of diesel fuel with air in the cylinders can be expressed simply as xc 12 H 23 þ yðo 2 þ 3:76N 2 Þ!12xCO 2 þ 11:5xH 2 O þ 3:76yN 2 þðy 17:75xÞO 2 ð24þ The molar flow rates of the fuel and the intake air are calculated based on the measured mass flow rates. Consequently, the mass fractions for the components CO 2,H 2 O, N 2, and O 2 are obtained according to the above equation. If the exhaust temperature and pressure are measured, the specific enthalpy for component i (where i represents either CO 2,H 2 O, N 2,orO 2 ) can be computed using the REFPROP software [24]. Typically, the exhaust temperature is between C and 6 C, and the exhaust pressure is slightly higher than atmospheric pressure. Therefore, the exhaust gas can be treated as a mixture of ideal gases [25]. Thus, the specific enthalpy of the exhaust gas can be calculated from h exh ðtþ ¼mf CO2 h CO2 ðtþþmf H2 Oh H2 OðTÞþmf N2 h N2 ðtþþmf O2 h O2 ðtþ ð25þ Finally, the waste heat carried away by the exhaust gas can be calculated using Eq. (3). The waste heat from the exhaust and the coolant are evaluated at each working point in the engine s entire operating region using the method outlined above. Fig. 4a gives the combustion energy, and the indicated power generated by the in-cylinder gas is shown in Fig. 4b. The heat energy from by the exhaust and the waste heat from the coolant are given in Fig. 4c and d, respectively. The combustion energy increases almost linearly with the engine output Fuel combustion energy (kw) Indicated power (kw) (a) 4 (b) Enthalpy of exhaust (kw) Waste heat of coolant (kw) (c) 4 (d) Fig. 4. Waste heat characteristics of the R425 engine. light-duty diesel engine. Appl Energy (212),

7 H.G. Zhang et al. / Applied Energy xxx (212) xxx xxx 7 T air,in (ºC) P air,in (bar) (a) 4 (b) T air,out (ºC) Waste heat of intake air (kw) (c) 4 (d) Fig. 5. Waste heat of the intake air. Table 3 Thermodynamic properties of the working fluids at the rated engine power. Cycles State no. Pressure (MPa) Temperature (K) Enthalpy (kj/kg) Entropy (kj/kg K) HT ORC LT ORC a b power, achieving 391 kw at the rated power point. Note that the indicated power and the waste heat quantity of the exhaust vary in a similar fashion. The reason for this is that the output power of a diesel engine is proportional to the quantity of injected fuel. Table 4 Results of energy loads and exergy destruction rates at the rated engine power. Subsystems E _ (kw) I _ (kw) Pump Evaporator Expander Pre-heater Pump Intercooler Evaporator Expander Condenser The indicated power only accounts for between 22.4% and 44.4% of the overall fuel combustion energy in the engine s operating region. Furthermore, the mechanical efficiency of the engine is between 17.2% and 75.5%. In addition, the exhaust enthalpy accounts for % of the combustion energy and the coolant waste heat accounts for %. At the rated power point, the light-duty diesel engine. Appl Energy (212),

8 .1 8 H.G. Zhang et al. / Applied Energy xxx (212) xxx xxx Mass flow rate of HT ORC (kg/s) Net power of HT ORC (kw) (a) 4 (b) Mass flow rate of LT ORC (kg/s) Net power of LT ORC (kw) (c) 4 (d) Fig. 6. Performances of the HT and LT loops. proportions of the indicated power, the exhaust enthalpy, and the coolant waste heat are 43.4%, 47.8%, and 8.8%, respectively. In particular, the variation in the coolant waste heat is not as regular as that from the other sources because of the thermal inertia of the coolant and the engine block. To evaluate the waste heat produced by the intake air, the air temperatures and pressures at both the inlet and outlet of the intercooler are measured. From these measurements, the corresponding enthalpy of the air can be calculated. Then, the waste heat from the intake air in the intercooler is obtained according to Eq. (1). The measured air temperatures and pressures at the inlet of the intercooler are shown in Fig. 5a and b, respectively. The air temperature increases with the engine power, and achieves 123 C at the rated power point. This result indicates that there is significant heat generated during the boosting process inside the compressor. The air pressure shows a similar tendency, and experiences a maximum of 2.24 bar when the engine speed is 26 r/min and the engine load is %. The measured air temperature at the outlet of the intercooler is shown in Fig. 5c. The intercooler performance is considered good because the air temperature is less than 3 C in most of the engine s operating region. The calculated waste heat from the intake air is shown in Fig. 5d. In the low idle region, the magnitude of the waste heat is very small. However, the waste heat increases with the engine power, and rises to 17.1 kw at the rated power point. In the engine s operating region where the engine output power is greater than 1 kw, the waste heat from the intake air exceeds 2 kw, and the intake air temperature is higher than 6 C. This indicates that the waste heat from the intake air can be used to pre-heat the high-pressure sub-cooled organic working fluid throughout most of the engine s operating region. 5. Combined system performance analysis After evaluating the waste heat for the exhaust, the coolant, and the intake air, the performance of the dual loop ORC system is analyzed at each measured engine operating point using the established mathematical model. The analysis program was written in Matlab [26] and the properties of the working fluids were computed by REFPROP 8. [24]. The results for the thermodynamic properties of the working fluids, where the engine is operating at the rated power, are given in Table 3 and the energy load and the exergy destruction rates are listed in Table 4. These results pro- light-duty diesel engine. Appl Energy (212),

9 H.G. Zhang et al. / Applied Energy xxx (212) xxx xxx 9 vide an estimation of the limited values, which can be used for the ORC component design. The following analysis is based on the outcomes using the energy equations and the exergy destruction rates throughout the engine s operating region will be assessed in the future. The performances for the HT loop and the LT loop are shown in Fig. 6. The mass flow rate of the R245fa is given in Fig. 6a. The results show that the maximum mass flow rate is.485 kg/s at the rated engine power point and that the mass flow rate decreases almost linearly with the engine output power. The net power of the HT loop is given in Fig. 6b, which shows that the maximum net power is kw at the rated engine power point. Note that the rate changes in Fig. 6a and b are very similar to the rate change in Fig. 4c because the mass flow rate and net output power are proportional to the heat exchanged inside of evaporator 1. The mass flow rate of the R134a for the LT loop is given in Fig. 6c. The results show that the maximum mass flow rate is.646 kg/s at the rated engine power point, which is 1.33 times that of the maximum flow rate in the HT loop. The mass flow rate also decreases linearly with the engine output power. Fig. 6d shows the net power of the LT loop. The maximum net power is kw at the rated engine power point, which is 1.71 times larger than the maximum net power of the HT loop. The net power decreases almost linearly with the engine output power. The heat addition from the LT loop is significantly larger than that of the HT loop because the LT loop absorbs the waste heat from the coolant, the high-temperature intake air, and the heat rejected from the HT loop. At the rated power point, the heat addition from the HT loop is kw, whereas the LT loop is kw. In addition, the thermal efficiency of the HT loop is 7.8%, which is lower than that of the LT loop (8.89%). Therefore, the LT loop performance is better than that of the HT loop. The performance map of the combined engine ORC system is then evaluated after the HT and LT loops are analyzed. Fig. 7a provides the overall net power map of the dual loop ORC system, and shows that the overall net power decreases linearly with the engine power. The maximum overall net power is kw at the rated power point, which improves the output power by 2%, relative to the system without the dual loop ORC. The improvement in the effective power over the engine s entire working region is displayed in Fig. 7b. In the high effective thermal efficiency region, the augmentation proportion is lowest (14 16%) because the waste heat quantity ratios are lower. The reason for this is better fuel combustion effects, the engine pumping losses are lower, and the ratio of the output power to the combustion energy is higher than Overall net power of ORCs (kw) Improvement of engine effective power (%) (a) 4 (b) bsfc of combined system (g/kw.h) Improvement of bsfc (%) (c) 4 (d) Fig. 7. Performance of the combined system. light-duty diesel engine. Appl Energy (212),

10 1 H.G. Zhang et al. / Applied Energy xxx (212) xxx xxx in the other regions. In the small load region, the augmentation proportion is at its peak (38 43%) because of the thermal inertia of the engine body and the coolant. Fig. 7c shows the bsfc map of the combined system. The bsfc of a light-duty diesel engine is calculated as bsfc ¼ _m f _W b and the bsfc of the combined system is defined as bsfc cs ¼ _m f _W b þ _ W n;ht þ _ W n;lt ð26þ ð27þ It can be observed that the bsfc of the combined system decreases significantly. In the peak effective thermal efficiency region, the bsfc is reduced from 212 g/kw h to 185 g/kw h. In the high-speed and small-load region, the bsfc is decreased even further from 6 g/kw h to 4 g/kw h. The relative augmentation of the bsfc is given in Fig. 7d. In the peak effective thermal efficiency region, the augmentation ratio is between 12% and 14%, whereas this ratio reaches 25 3% in the small load region for the same reasons that the power augmentation is highest in this region. 6. Conclusions In this study, the waste heat from the exhaust, the intake air, and the coolant of a R425 diesel engine are analyzed using measured data. A novel dual-loop ORC system is designed to recover the waste heat from the exhaust, the intake air, and the coolant. The performance map of the combined system is evaluated over the engine s entire operating region. Based on this analysis, the following can be concluded: 1. The combustion energy is much greater than the engine output power through most of the operating region. The exhaust gas enthalpy ( % of the total combustion energy) is slightly higher than the indicated power ( % of the total combustion energy). The waste heat from the intake air is smaller than that from the coolant. It is, however, still worthwhile to recover this energy in the middle and high engine power region. 2. A dual loop ORC system is designed to recover heat from these three distinct sources simultaneously. An HT loop recovers the exhaust waste heat using R245fa as the working fluid. An LT loop recovers the waste heat from the coolant and the intake air, as well as the residual heat from the HT loop using R134a as the working fluid. The results show that the net power of the LT loop is higher than that of the HT loop (a total of 6.98 kw for the HT loop versus kw for the LT loop at the rated power point). 3. The performance map of the combined system is evaluated using the first law method. In the peak effective thermal efficiency region, the augmentation proportion of the effective power for the combined system is the lowest, at 14 16%, but is highest in the small-load and high-speed region where the augmentation proportion is 38 43%. The bsfc is also found to significantly decrease throughout the engine s operating region. From the viewpoint of power performance and fuel economy, the dual loop ORC system is a promising scheme to recover the waste heat from a vehicular light-duty diesel engine. Acknowledgements This work was sponsored by the National Basic Research (973) Program of China (Grants #211CB7722 and #211CB7174), the National High-Tech Research and Development Program of China (863 Program) (Grant No. 9AA5Z26), and the Funding Project for Academic Human Resources Development in Institutions of Higher Learning Under the Jurisdiction of Beijing Municipality (Grant No. PHR2819). References [1] Fiaschi D, Manfrida G, Maraschiello F. Thermo-fluid dynamics preliminary design of turbo-expanders for ORC cycles. Appl Energy 212;97:61 8. [2] Tempesti D, Manfrida G, Fiaschi D. Thermodynamic analysis of two micro CHP systems operating with geothermal and solar energy. Appl Energy 212;97: [3] De Pascale A, Ferrari C, Melino F, Morini M, Pinelli M. Integration between a thermophotovoltaic generator and an organic Rankine cycle. Appl Energy 212;97: [4] Clemente S, Micheli D, Reini M, Taccani R. Energy efficiency analysis of organic Rankine cycles with scroll expanders for cogenerative applications. Appl Energy 212;97: [5] Roy JP, Mishra MK, Misra A. Performance analysis of an organic Rankine cycle with superheating under different heat source temperature conditions. Appl Energy 211;88: [6] Wang EH, Zhang HG, Fan BY, Ouyang MG, Zhao Y, Mu QH. Study of working fluid selection of organic Rankine cycle (ORC) for engine waste heat recovery. Energy 211;36: [7] Quoilin S, Lemort V, Lebrun J. Experimental study and modeling of an organic Rankine cycle using scroll expander. Appl Energy 21;87: [8] Quoilin S, Aumann R, Grill A, Schuster A, Lemort V, Spliethoff H. Dynamic modeling and optimal control strategy of waste heat recovery. Appl Energy 211;88: [9] Zhang S, Wang H, Guo T. Performance comparison and parametric optimization of subcritical organic Rankine cycle (ORC) and transcritical power cycle system for low-temperature geothermal power generation. Appl Energy 211;88: [1] Bianchi M, De Pascale A. Bottoming cycles for electric energy generation: parametric investigation of available and innovative solutions for the exploitation of low and medium temperature heat sources. Appl Energy 211;88:1 9. [11] Qiu K, Hayden ACS. Integrated thermoelectric and organic Rankine cycles for micro-chp systems. Appl Energy 212;97: [12] Teng H, Regner G, Cowland C. Achieving high engine efficiency for heavy-duty diesel engines by waste heat recovery using supercritical organic-fluid Rankine cycle. SAE ; 6. [13] Freymann R, Strobl W, Obieglo A. The turbosteamer: a system introducing the principle of cogeneration in automotive applications. MTZ 8;69: [14] Ringler J, Seifert M, Guyotot V, Hubner W. Rankine cycle for waste heat recovery of IC engines. SAE ; 9. [15] Wang EH, Zhang HG, Zhao Y, Fan BY, Wu YT, Mu QH. Performance analysis of a novel system combining a dual loop organic Rankine cycle (ORC) with a gasoline engine. Energy 212;43: [16] He W, Wu YT, Ma CF, Ma GY. Performance study on three-stage power system of compressed air vehicle based on single screw expander. Sci China Technol Sci 21;53: [17] Wang W, Wu YT, Ma CF, Liu LD, Yu J. Preliminary experimental study of single screw expander prototype. Appl Therm Eng 211;31: [18] Honyewell. Genertron Ò 245fa applications development guide. Morristown, USA: Honeywell Fluorine Products;. [19] Calm JM, Hourahan GC. Refrigerant data summary. Eng Syst 1;18: [2] Bahadori A. Estimation of combustion flue gas acid dew point during heat recovery and efficiency gain. Appl Therm Eng 211;31: [21] National standard of the People s Republic of China. GB/T ; 1. [22] Heywood JB. Internal combustion engine fundamentals. New York: McGraw- Hill; [23] On the web: < [24] REFPROP version 8.. NIST standard reference database 23. America: The US Secretary of Commerce; 7. [25] Cengel YA, Boles MA. Thermodynamics an engineering approach. 6th ed. London: McGraw-Hill; 8. [26] MATLAB version R14SP3. Matlab user s guide. US: The MathWorks, Inc.; 5. light-duty diesel engine. Appl Energy (212),