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A performance analysis of a novel system of a dual loop bottoming organic Rankine cycle (ORC) with a light-duty diesel engine
Applied Energy 102 (2013) 1504–1513
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Applied Energy journal homepage: www.elsevier.com/locate/apenergy
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. 100, 100124 Beijing, China
h i g h l i g h t s " 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.
a r t i c l e
i n f o
Article history: Received 17 June 2012 Received in revised form 5 September 2012 Accepted 9 September 2012 Available online 9 October 2012 Keywords: Organic Rankine cycle Waste heat recovery Diesel engine Performance MAP Dual loop
a b s t r a c t 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 2012 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.: +86 10 6739 2469; fax: +86 10 6739 2774. E-mail address: [email protected] (H.G. Zhang). 0306-2619/$ - see front matter Ó 2012 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.apenergy.2012.09.018
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 (80– 100 °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 [10–15]. 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
H.G. Zhang et al. / Applied Energy 102 (2013) 1504–1513
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Nomenclature _ W _ m h s I_ T P Q_ mf x, y
power (kW) mass flow rate (kg/s) enthalpy (kJ/kg) entropy (kJ/kg K) exergy destruction rate (kW) temperature (K) pressure (MPa) heat quantity (kW) mass fraction molar amount
Greek letters g efficiency Subscript cr critical point bp normal boiling point 0 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
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-condi-
s1 s2 pre int cool c mc me f a b i misc n cs HT LT
expander 1 expander 2 pre-heater intercooler coolant condenser mean condensing temperature mean evaporation temperature fuel intake air brake indicated miscellaneous net combined system HT loop 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 CO2) SUV sport utility vehicle bsfc brake specific fuel consumption
tioners. 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 red1 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.
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Reservior2 Pump2 LT2 Intercooler LT1
LT3 Tcool,in
Tair,in
Fan
Tcool,out
Preheater
Tair,out
LT4
Evaporator2 Reservior1 HT4 Engine
LT5
HT1
Expander1
Pump1 HT3
Compressor
HT2
Turbine Evaporator1 Texh,in
Texh,out
Expander2 LT6 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)
Tcr (K)
Pcr (MPa)
Tbp (K)
ASHRAE 34 safety group
Atmospheric life time (yr)
ODP
GWP (100 yr)
R245fa R134a
134.05 102.03
427.2 374.21
3.639 4.059
288.05 247.08
B1 A1
7.2 13.8
0 0
950 1300
440
R245fa
420
HT3
Temperature (K)
400
ð1Þ
_ HT ðsHT2 sHT1 Þ I_p1 ¼ T 0 m
ð2Þ
The heat transfer process between the exhaust gas and the R245fa fluid in evaporator 1 is denoted as
380
HT2 360
HT4
_ exh ðhexh;in hexh;out Þ ¼ m _ HT ðhHT3 hHT2 Þ Q_ exh ¼ m
HT1
340
a
b
LT4
_ HT ðsHT3 sHT2 Þ þ T 0 m _ exh ðsexh;out sexh;in Þ I_e1 ¼ T 0 m
LT2 300
LT6 LT1
280
R134a 260 1
1.2
1.4
1.6
1.8
ð3Þ
The exergy destruction rate in evaporator 1 is calculated as
LT5
LT3
320
240 0.8
_ p1 ¼ m _ HT ðhHT2 hHT1 Þ ¼ m _ HT ðhHT2s hHT1 Þ=gp1 W
2
Entropy (kJ/kg.K) Fig. 2. T–s plots of the HT and LT loops.
3. Mathematical model 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
ð4Þ
When combustion flue gas is cooled in a heat recovery application, the temperature must not be allowed to drop below the acid dew point [20]. For this reason, it is desirable for the cooled exhaust temperature to be set above 100 °C. In this study, the exhaust temperature at the outlet of evaporator 1 is specified as 105 °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
_ s1 ¼ m _ HT ðhHT3 hHT4 Þ ¼ m _ HT ðhHT3 hHT4s Þgs1 W
ð5Þ
The exergy destruction rate of expander 1 is expressed as
_ HT ðsHT4 sHT3 Þ I_s1 ¼ T 0 m
ð6Þ
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The heat rejection rate for the process HT4 to HT1 is denoted as
_ HT ðhHT4 hHT1 Þ Q_ pre ¼ m
ð7Þ
In the LT loop, the work of pump 2 in the process from LT1 to LT2 is expressed as
_ p2 ¼ m _ LT ðhLT2 hLT1 Þ ¼ m _ LT ðhLT2s hLT1 Þ=gp2 W
ð8Þ
The exergy destruction rate of pump 2 is calculated as
_ LT ðsLT2 sLT1 Þ I_p2 ¼ T 0 m
ð9Þ
The heat transfer in the intercooler is denoted as
_ a ðha;in ha;out Þ ¼ m _ LT ðhLT3 hLT2 Þ Q_ int ¼ m
ð10Þ
The exergy destruction rate in the intercooler is represented as
_ LT ðsLT3 sLT2 Þ þ T 0 m _ a ðsa;out sa;in Þ I_int ¼ T 0 m
ð11Þ
The heat added to the R134a in the pre-heater is calculated as
_ LT ðhLT4 hLT3 Þ ¼ m _ HT ðhHT4 hHT1 Þ Q_ pre ¼ m
ð12Þ
The exergy destruction rate in the pre-heater is defined as
_ HT ðsHT1 sHT4 Þ þ T 0 m _ LT ðsLT4 sLT3 Þ I_pre ¼ T 0 m
ð13Þ
The heat added in evaporator 2 is calculated as
_ LT ðhLT5 hLT4 Þ Q_ e2 ¼ m
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
h hLT4 _ LT ðsLT5 sLT4 Þ LT5 I_e2 ¼ T 0 m T me
ð16Þ
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
ð19Þ
The exergy destruction rate of expander 2 is expressed as
_ LT ðsLT6 sLT5 Þ I_s2 ¼ T 0 m
ð20Þ
The heat rejected from the R134a working fluid in the condenser is calculated from
_ LT ðhLT6 hLT1 Þ Q_ c ¼ m
h hLT6 _ LT ðsLT1 sLT6 Þ LT1 I_c ¼ T 0 m 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 70 °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 10 °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 20 °C. (5) The air temperature at the intercooler outlet is set to 30 °C. (6) The isentropic efficiencies of pump 1 and pump 2 are set to 0.8. 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 0.75. 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 1000 r/min and 4000 r/ min, respectively. The intermediate speeds are selected using a step increment of 200 r/min. At each selected engine speed, eight different load values are selected, ranging from a 100% load to a minimum stable load value.
ð18Þ
where TLTb is the evaporation temperature of the R134a at the same pressure in state LT5. The output work generated by expander 2 is expressed as
_ s2 ¼ m _ LT ðhLT5 hLT6 Þ ¼ m _ LT ðhLT5 hLT6s Þgs2 W
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
ð21Þ
Table 2 Specifications of the R425 diesel engine. Item
Parameter
Unit
Model Displacement Bore stroke Cylinder number Valve number per cylinder Fuel injection equipment Rated power Rated speed Max. torque Speed at max. torque
R425 2.499 92 94 4 4 Common rail injection system 105 4000 340 2000 2400
– L mm – – – kW r/min Nm r/min
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Brake specific fuel consumption (g/kW.h)
220
200
24
150 100
20
250
40
30
350
0 1000
260
280
10
50
400
300
500
2000
3000
250
33
30
100
27
23
50 0 1000
4000
2000
0.16
0.1
500
100
0.07
550
150
0.14
600
200
0.12
650
250
0.18
70 0
0.05
Engine Torque (N.m)
300
0.03
Engine Torque (N.m)
4000
Exhaust mass flow rate (kg/s)
350
74 5
50
3000
(b)
250
100
20
Engine speed (r/min)
78 0
150
32
600
Exhaust temperature (K)
200
34
150
(a)
300
35
200
Engine speed (r/min)
350
36
90
70
60
50
0
225
80
37
5 21
Engine Torque (N.m)
23 0
21
Engine Torque (N.m)
250
38
300
0
300
Effective thermal efficiency (%)
350
39
350
50
450 0 1000
2000
3000
4000
0 1000
2000
3000
Engine speed (r/min)
Engine speed (r/min)
(c)
(d)
4000
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 250 g/kW h. The lowest bsfc zone is situated in the high duty range between 1800 r/min and 2400 r/min, and the minimum bsfc value is 209.3 g/kW h. The measured bsfc data indicate the R425 diesel engine has good fuel economy. The maximum output torque the engine achieves is 340 N m between 2000 r/min and 2600 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 0.1894 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
_ i þ Q_ cool þ Q_ misc þ m _ a ha ¼ W _ exh hexh _ f hf þ m m
ð23Þ
_ f and m _ a are the fuel and air mass flow rates, respectively. Here m The variables hf and ha are the corresponding inlet enthalpies. The _ i is the indicated power of the engine, which can be calvariable W culated using the in-cylinder pressure [22]. In this study, the in-cylinder pressure is measured using a piezoelectric transducer (Kistler
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6055Bsp) connected to an amplifier (Kistler 5011). 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 C12H23 [23]. Therefore, the combustion process of diesel fuel with air in the cylinders can be expressed simply as
xC12 H23 þ yðO2 þ 3:76N2 Þ ! 12xCO2 þ 11:5xH2 O þ 3:76yN2 þ ðy 17:75xÞO2
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 CO2, H2O, N2, and O2 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 CO2, H2O, N2, or O2) can be computed using the REFPROP software [24]. Typically, the exhaust temperature is between 200 °C and 600 °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
hexh ðTÞ ¼ mfCO2 hCO2 ðTÞ þ mfH2 O hH2 O ðTÞ þ mfN2 hN2 ðTÞ þ mfO2 hO2 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
Indicated power (kW)
300
300
30
0 1000
100
20 50
2000
3000
10
0 1000
4000
2000
Engine speed (r/min)
Waste heat of coolant (kW)
300
300
160
250
5 17
0 14
120
200
0 10
80 60
52
250
42 200
32
150
22
Engine Torque (N.m)
350
40
Engine Torque (N.m)
Enthalpy of exhaust (kW)
100
12
20 50
50
12 0 1000
4000
(b)
350
100
3000
Engine speed (r/min)
(a)
150
0 16
0 12
150
60
50
80
100
200
60
0 20
0 15
150
0 14
0 35
0 30 0 25
200
250
0 10
250
40
Engine Torque (N.m)
350
0 10
Engine Torque (N.m)
Fuel combustion energy (kW) 350
2000
3000
Engine speed (r/min)
4000
4
0 1000
2000
3000
Engine speed (r/min)
(c)
(d) Fig. 4. Waste heat characteristics of the R425 engine.
4000
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Pair,in (bar)
300
300
250
120
5 11
200
0 11
150
50
0 1000
2000
3000
0 1000
4000
1.5
100
2000
Engine speed (r/min)
Waste heat of intake air (kW)
300
300
250
200
14
8
30
16
200
250
11
150
5
26
30
2
Engine Torque (N.m)
350
22
Engine Torque (N.m)
T air,out (ºC)
100
1
18 2000
3000
4000
0.5
50
50 0 1000
4000
(b)
350
100
3000
Engine speed (r/min)
(a)
150
2
1.1
40
50
2.1
9 1.
0 10
100
200
1.7
80
150
2.2
250
1.3
Engine Torque (N.m)
350
60
Engine Torque (N.m)
T air,in (ºC) 350
0 1000
2000
3000
4000
Engine speed (r/min)
Engine speed (r/min)
(c)
(d) Fig. 5. Waste heat of the intake air.
Table 3 Thermodynamic properties of the working fluids at the rated engine power.
Table 4 Results of energy loads and exergy destruction rates at the rated engine power.
Cycles
State no.
Pressure (MPa)
Temperature (K)
Enthalpy (kJ/kg)
Entropy (kJ/kg K)
Subsystems
E_ (kW)
I_ (kW)
HT ORC
1 2 3 4
0.695 2.4 2.4 0.695
348.15 349.38 404.38 359.79
301.87 303.65 487.97 471.82
1.327 1.328 1.800 1.815
Pump 1 Evaporator 1 Expander 1 Pre-heater
0.8614 89.45 7.836 82.48
0.1325 19.39 2.001 2.104
LT ORC
1 2 3 a 4 b 5 6
0.572 2.117 2.117 2.117 2.117 2.117 2.117 0.572
293.15 294.16 312.69 343.15 343.15 343.15 348.15 299.75
227.47 229.02 255.53 304.28 383.12 428.65 436.26 416.28
1.096 1.097 1.185 1.333 1.563 1.696 1.718 1.740
Pump 2 Intercooler Evaporator 2 Expander 2 Condenser
1.002 17.14 34.35 12.92 122.05
0.1765 28.78 1.099 3.971 0.0305
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.
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 30.5–54.5% of the combustion energy and the coolant waste heat accounts for 5.9–30.8%. At the rated power point, the
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Net power of HT ORC (kW)
300
300
250
0.1
100 50
01 0. 0 1000
0. 03
0.1 5
0.0 6
6
150
0.2 5
250
5
0. 2
45 0.
200
35 0.
0. 3
Engine Torque (N.m)
350
0.4
Engine Torque (N.m)
Mass flow rate of HT ORC (kg/s) 350
3 150
3000
1. 5
100
0.6
0. 3
0 1000
4000
2000
3000
(a)
(b) Net power of LT ORC (kW)
300
300
250
0. 5
200
0.4 0.3 0.2
0 1000
0.0 4
250
3000
4000
11
7 150
5
3
100 50
2000
9
200
0.1
50
Engine Torque (N.m)
350
6 0.
Engine Torque (N.m)
Mass flow rate of LT ORC (kg/s) 350
100
4000
Engine speed (r/min)
Engine speed (r/min)
150
2
1 50
2000
4
200
1
0 1000
Engine speed (r/min)
2000
3000
4000
Engine speed (r/min)
(c)
(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. (10). 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 2600 r/min and the engine load is 100%. 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 30 °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 10 kW, the waste heat from the intake air exceeds 2 kW, and the intake air temperature is higher than 60 °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 8.0 [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-
H.G. Zhang et al. / Applied Energy 102 (2013) 1504–1513
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 0.485 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 6.975 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 0.646 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 11.913 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 89.45 kW, whereas the LT loop is 133.97 kW. In addition, the thermal efficiency of the HT loop is 7.80%, 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 18.89 kW at the rated power point, which improves the output power by 20%, 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 Improvement of engine effective power (%)
300
300
200
10
150
7
4
100 50 0 1000
1
2
200
15
250
18
16
16
13
17
250
150 100
19
14
20
50
2000
3000
2000
Engine speed (r/min)
3000
210
220
100
260
50 0 1000
2000
300
3000
13 15
13
200
14
19 0
200
250
14
300
Engine Torque (N.m)
300
0 20
4000
Improvement of bsfc (%) 350
18 4
Engine Torque (N.m)
bsfc of combined system (g/kW.h)
150
42
(b)
350
186
26 34
22
Engine speed (r/min)
(a)
250
30
38
0 1000
4000
17
350
Engine Torque (N.m)
Engine Torque (N.m)
Overall net power of ORCs (kW) 350
16
1512
.5
15
150 100
16
12
18 50
400 4000
0 1000
25 2000
3000
Engine speed (r/min)
Engine speed (r/min)
(c)
(d) Fig. 7. Performance of the combined system.
17
20 30 4000
H.G. Zhang et al. / Applied Energy 102 (2013) 1504–1513
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
_f m _ Wb
ð26Þ
and the bsfc of the combined system is defined as
bsfccs ¼
_f m _ _ _ n;LT W b þ W n;HT þ W
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 600 g/kW h to 400 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 25–30% 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 (30.5–54.5% of the total combustion energy) is slightly higher than the indicated power (22.4–44.4% 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 11.91 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.
1513
Acknowledgements This work was sponsored by the National Basic Research (973) Program of China (Grants #2011CB707202 and #2011CB710704), the National High-Tech Research and Development Program of China (863 Program) (Grant No. 2009AA05Z206), and the Funding Project for Academic Human Resources Development in Institutions of Higher Learning Under the Jurisdiction of Beijing Municipality (Grant No. PHR201008019). References [1] Fiaschi D, Manfrida G, Maraschiello F. Thermo-fluid dynamics preliminary design of turbo-expanders for ORC cycles. Appl Energy 2012;97:601–8. [2] Tempesti D, Manfrida G, Fiaschi D. Thermodynamic analysis of two micro CHP systems operating with geothermal and solar energy. Appl Energy 2012;97:609–17. [3] De Pascale A, Ferrari C, Melino F, Morini M, Pinelli M. Integration between a thermophotovoltaic generator and an organic Rankine cycle. Appl Energy 2012;97:695–703. [4] Clemente S, Micheli D, Reini M, Taccani R. Energy efficiency analysis of organic Rankine cycles with scroll expanders for cogenerative applications. Appl Energy 2012;97:792–801. [5] Roy JP, Mishra MK, Misra A. Performance analysis of an organic Rankine cycle with superheating under different heat source temperature conditions. Appl Energy 2011;88:2995–3004. [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 2011;36:3406–18. [7] Quoilin S, Lemort V, Lebrun J. Experimental study and modeling of an organic Rankine cycle using scroll expander. Appl Energy 2010;87:1260–8. [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 2011;88:2183–90. [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 2011;88:2740–54. [10] 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 2011;88:1500–9. [11] Qiu K, Hayden ACS. Integrated thermoelectric and organic Rankine cycles for micro-CHP systems. Appl Energy 2012;97:667–72. [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 2006-01-3522; 2006. [13] Freymann R, Strobl W, Obieglo A. The turbosteamer: a system introducing the principle of cogeneration in automotive applications. MTZ 2008;69:404–12. [14] Ringler J, Seifert M, Guyotot V, Hubner W. Rankine cycle for waste heat recovery of IC engines. SAE 2009-01-0174; 2009. [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 2012;43:385–95. [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 2010;53:2299–303. [17] Wang W, Wu YT, Ma CF, Liu LD, Yu J. Preliminary experimental study of single screw expander prototype. Appl Therm Eng 2011;31:3684–8. [18] Honyewell. GenertronÒ245fa applications development guide. Morristown, USA: Honeywell Fluorine Products; 2000. [19] Calm JM, Hourahan GC. Refrigerant data summary. Eng Syst 2001;18:74–88. [20] Bahadori A. Estimation of combustion flue gas acid dew point during heat recovery and efficiency gain. Appl Therm Eng 2011;31:1457–62. [21] National standard of the People’s Republic of China. GB/T 18297-2001; 2001. [22] Heywood JB. Internal combustion engine fundamentals. New York: McGrawHill; 1998. [23] On the web:. [24] REFPROP version 8.0. NIST standard reference database 23. America: The US Secretary of Commerce; 2007. [25] Cengel YA, Boles MA. Thermodynamics – an engineering approach. 6th ed. London: McGraw-Hill; 2008. [26] MATLAB version R14SP3. Matlab user’s guide. US: The MathWorks, Inc.; 2005.
Contents lists available at SciVerse ScienceDirect
Applied Energy journal homepage: www.elsevier.com/locate/apenergy
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. 100, 100124 Beijing, China
h i g h l i g h t s " 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.
a r t i c l e
i n f o
Article history: Received 17 June 2012 Received in revised form 5 September 2012 Accepted 9 September 2012 Available online 9 October 2012 Keywords: Organic Rankine cycle Waste heat recovery Diesel engine Performance MAP Dual loop
a b s t r a c t 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 2012 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.: +86 10 6739 2469; fax: +86 10 6739 2774. E-mail address: [email protected] (H.G. Zhang). 0306-2619/$ - see front matter Ó 2012 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.apenergy.2012.09.018
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 (80– 100 °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 [10–15]. 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
H.G. Zhang et al. / Applied Energy 102 (2013) 1504–1513
1505
Nomenclature _ W _ m h s I_ T P Q_ mf x, y
power (kW) mass flow rate (kg/s) enthalpy (kJ/kg) entropy (kJ/kg K) exergy destruction rate (kW) temperature (K) pressure (MPa) heat quantity (kW) mass fraction molar amount
Greek letters g efficiency Subscript cr critical point bp normal boiling point 0 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
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-condi-
s1 s2 pre int cool c mc me f a b i misc n cs HT LT
expander 1 expander 2 pre-heater intercooler coolant condenser mean condensing temperature mean evaporation temperature fuel intake air brake indicated miscellaneous net combined system HT loop 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 CO2) SUV sport utility vehicle bsfc brake specific fuel consumption
tioners. 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 red1 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.
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Reservior2 Pump2 LT2 Intercooler LT1
LT3 Tcool,in
Tair,in
Fan
Tcool,out
Preheater
Tair,out
LT4
Evaporator2 Reservior1 HT4 Engine
LT5
HT1
Expander1
Pump1 HT3
Compressor
HT2
Turbine Evaporator1 Texh,in
Texh,out
Expander2 LT6 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)
Tcr (K)
Pcr (MPa)
Tbp (K)
ASHRAE 34 safety group
Atmospheric life time (yr)
ODP
GWP (100 yr)
R245fa R134a
134.05 102.03
427.2 374.21
3.639 4.059
288.05 247.08
B1 A1
7.2 13.8
0 0
950 1300
440
R245fa
420
HT3
Temperature (K)
400
ð1Þ
_ HT ðsHT2 sHT1 Þ I_p1 ¼ T 0 m
ð2Þ
The heat transfer process between the exhaust gas and the R245fa fluid in evaporator 1 is denoted as
380
HT2 360
HT4
_ exh ðhexh;in hexh;out Þ ¼ m _ HT ðhHT3 hHT2 Þ Q_ exh ¼ m
HT1
340
a
b
LT4
_ HT ðsHT3 sHT2 Þ þ T 0 m _ exh ðsexh;out sexh;in Þ I_e1 ¼ T 0 m
LT2 300
LT6 LT1
280
R134a 260 1
1.2
1.4
1.6
1.8
ð3Þ
The exergy destruction rate in evaporator 1 is calculated as
LT5
LT3
320
240 0.8
_ p1 ¼ m _ HT ðhHT2 hHT1 Þ ¼ m _ HT ðhHT2s hHT1 Þ=gp1 W
2
Entropy (kJ/kg.K) Fig. 2. T–s plots of the HT and LT loops.
3. Mathematical model 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
ð4Þ
When combustion flue gas is cooled in a heat recovery application, the temperature must not be allowed to drop below the acid dew point [20]. For this reason, it is desirable for the cooled exhaust temperature to be set above 100 °C. In this study, the exhaust temperature at the outlet of evaporator 1 is specified as 105 °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
_ s1 ¼ m _ HT ðhHT3 hHT4 Þ ¼ m _ HT ðhHT3 hHT4s Þgs1 W
ð5Þ
The exergy destruction rate of expander 1 is expressed as
_ HT ðsHT4 sHT3 Þ I_s1 ¼ T 0 m
ð6Þ
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H.G. Zhang et al. / Applied Energy 102 (2013) 1504–1513
The heat rejection rate for the process HT4 to HT1 is denoted as
_ HT ðhHT4 hHT1 Þ Q_ pre ¼ m
ð7Þ
In the LT loop, the work of pump 2 in the process from LT1 to LT2 is expressed as
_ p2 ¼ m _ LT ðhLT2 hLT1 Þ ¼ m _ LT ðhLT2s hLT1 Þ=gp2 W
ð8Þ
The exergy destruction rate of pump 2 is calculated as
_ LT ðsLT2 sLT1 Þ I_p2 ¼ T 0 m
ð9Þ
The heat transfer in the intercooler is denoted as
_ a ðha;in ha;out Þ ¼ m _ LT ðhLT3 hLT2 Þ Q_ int ¼ m
ð10Þ
The exergy destruction rate in the intercooler is represented as
_ LT ðsLT3 sLT2 Þ þ T 0 m _ a ðsa;out sa;in Þ I_int ¼ T 0 m
ð11Þ
The heat added to the R134a in the pre-heater is calculated as
_ LT ðhLT4 hLT3 Þ ¼ m _ HT ðhHT4 hHT1 Þ Q_ pre ¼ m
ð12Þ
The exergy destruction rate in the pre-heater is defined as
_ HT ðsHT1 sHT4 Þ þ T 0 m _ LT ðsLT4 sLT3 Þ I_pre ¼ T 0 m
ð13Þ
The heat added in evaporator 2 is calculated as
_ LT ðhLT5 hLT4 Þ Q_ e2 ¼ m
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
h hLT4 _ LT ðsLT5 sLT4 Þ LT5 I_e2 ¼ T 0 m T me
ð16Þ
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
ð19Þ
The exergy destruction rate of expander 2 is expressed as
_ LT ðsLT6 sLT5 Þ I_s2 ¼ T 0 m
ð20Þ
The heat rejected from the R134a working fluid in the condenser is calculated from
_ LT ðhLT6 hLT1 Þ Q_ c ¼ m
h hLT6 _ LT ðsLT1 sLT6 Þ LT1 I_c ¼ T 0 m 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 70 °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 10 °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 20 °C. (5) The air temperature at the intercooler outlet is set to 30 °C. (6) The isentropic efficiencies of pump 1 and pump 2 are set to 0.8. 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 0.75. 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 1000 r/min and 4000 r/ min, respectively. The intermediate speeds are selected using a step increment of 200 r/min. At each selected engine speed, eight different load values are selected, ranging from a 100% load to a minimum stable load value.
ð18Þ
where TLTb is the evaporation temperature of the R134a at the same pressure in state LT5. The output work generated by expander 2 is expressed as
_ s2 ¼ m _ LT ðhLT5 hLT6 Þ ¼ m _ LT ðhLT5 hLT6s Þgs2 W
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
ð21Þ
Table 2 Specifications of the R425 diesel engine. Item
Parameter
Unit
Model Displacement Bore stroke Cylinder number Valve number per cylinder Fuel injection equipment Rated power Rated speed Max. torque Speed at max. torque
R425 2.499 92 94 4 4 Common rail injection system 105 4000 340 2000 2400
– L mm – – – kW r/min Nm r/min
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Brake specific fuel consumption (g/kW.h)
220
200
24
150 100
20
250
40
30
350
0 1000
260
280
10
50
400
300
500
2000
3000
250
33
30
100
27
23
50 0 1000
4000
2000
0.16
0.1
500
100
0.07
550
150
0.14
600
200
0.12
650
250
0.18
70 0
0.05
Engine Torque (N.m)
300
0.03
Engine Torque (N.m)
4000
Exhaust mass flow rate (kg/s)
350
74 5
50
3000
(b)
250
100
20
Engine speed (r/min)
78 0
150
32
600
Exhaust temperature (K)
200
34
150
(a)
300
35
200
Engine speed (r/min)
350
36
90
70
60
50
0
225
80
37
5 21
Engine Torque (N.m)
23 0
21
Engine Torque (N.m)
250
38
300
0
300
Effective thermal efficiency (%)
350
39
350
50
450 0 1000
2000
3000
4000
0 1000
2000
3000
Engine speed (r/min)
Engine speed (r/min)
(c)
(d)
4000
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 250 g/kW h. The lowest bsfc zone is situated in the high duty range between 1800 r/min and 2400 r/min, and the minimum bsfc value is 209.3 g/kW h. The measured bsfc data indicate the R425 diesel engine has good fuel economy. The maximum output torque the engine achieves is 340 N m between 2000 r/min and 2600 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 0.1894 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
_ i þ Q_ cool þ Q_ misc þ m _ a ha ¼ W _ exh hexh _ f hf þ m m
ð23Þ
_ f and m _ a are the fuel and air mass flow rates, respectively. Here m The variables hf and ha are the corresponding inlet enthalpies. The _ i is the indicated power of the engine, which can be calvariable W culated using the in-cylinder pressure [22]. In this study, the in-cylinder pressure is measured using a piezoelectric transducer (Kistler
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H.G. Zhang et al. / Applied Energy 102 (2013) 1504–1513
6055Bsp) connected to an amplifier (Kistler 5011). 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 C12H23 [23]. Therefore, the combustion process of diesel fuel with air in the cylinders can be expressed simply as
xC12 H23 þ yðO2 þ 3:76N2 Þ ! 12xCO2 þ 11:5xH2 O þ 3:76yN2 þ ðy 17:75xÞO2
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 CO2, H2O, N2, and O2 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 CO2, H2O, N2, or O2) can be computed using the REFPROP software [24]. Typically, the exhaust temperature is between 200 °C and 600 °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
hexh ðTÞ ¼ mfCO2 hCO2 ðTÞ þ mfH2 O hH2 O ðTÞ þ mfN2 hN2 ðTÞ þ mfO2 hO2 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
Indicated power (kW)
300
300
30
0 1000
100
20 50
2000
3000
10
0 1000
4000
2000
Engine speed (r/min)
Waste heat of coolant (kW)
300
300
160
250
5 17
0 14
120
200
0 10
80 60
52
250
42 200
32
150
22
Engine Torque (N.m)
350
40
Engine Torque (N.m)
Enthalpy of exhaust (kW)
100
12
20 50
50
12 0 1000
4000
(b)
350
100
3000
Engine speed (r/min)
(a)
150
0 16
0 12
150
60
50
80
100
200
60
0 20
0 15
150
0 14
0 35
0 30 0 25
200
250
0 10
250
40
Engine Torque (N.m)
350
0 10
Engine Torque (N.m)
Fuel combustion energy (kW) 350
2000
3000
Engine speed (r/min)
4000
4
0 1000
2000
3000
Engine speed (r/min)
(c)
(d) Fig. 4. Waste heat characteristics of the R425 engine.
4000
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H.G. Zhang et al. / Applied Energy 102 (2013) 1504–1513
Pair,in (bar)
300
300
250
120
5 11
200
0 11
150
50
0 1000
2000
3000
0 1000
4000
1.5
100
2000
Engine speed (r/min)
Waste heat of intake air (kW)
300
300
250
200
14
8
30
16
200
250
11
150
5
26
30
2
Engine Torque (N.m)
350
22
Engine Torque (N.m)
T air,out (ºC)
100
1
18 2000
3000
4000
0.5
50
50 0 1000
4000
(b)
350
100
3000
Engine speed (r/min)
(a)
150
2
1.1
40
50
2.1
9 1.
0 10
100
200
1.7
80
150
2.2
250
1.3
Engine Torque (N.m)
350
60
Engine Torque (N.m)
T air,in (ºC) 350
0 1000
2000
3000
4000
Engine speed (r/min)
Engine speed (r/min)
(c)
(d) Fig. 5. Waste heat of the intake air.
Table 3 Thermodynamic properties of the working fluids at the rated engine power.
Table 4 Results of energy loads and exergy destruction rates at the rated engine power.
Cycles
State no.
Pressure (MPa)
Temperature (K)
Enthalpy (kJ/kg)
Entropy (kJ/kg K)
Subsystems
E_ (kW)
I_ (kW)
HT ORC
1 2 3 4
0.695 2.4 2.4 0.695
348.15 349.38 404.38 359.79
301.87 303.65 487.97 471.82
1.327 1.328 1.800 1.815
Pump 1 Evaporator 1 Expander 1 Pre-heater
0.8614 89.45 7.836 82.48
0.1325 19.39 2.001 2.104
LT ORC
1 2 3 a 4 b 5 6
0.572 2.117 2.117 2.117 2.117 2.117 2.117 0.572
293.15 294.16 312.69 343.15 343.15 343.15 348.15 299.75
227.47 229.02 255.53 304.28 383.12 428.65 436.26 416.28
1.096 1.097 1.185 1.333 1.563 1.696 1.718 1.740
Pump 2 Intercooler Evaporator 2 Expander 2 Condenser
1.002 17.14 34.35 12.92 122.05
0.1765 28.78 1.099 3.971 0.0305
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.
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 30.5–54.5% of the combustion energy and the coolant waste heat accounts for 5.9–30.8%. At the rated power point, the
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Net power of HT ORC (kW)
300
300
250
0.1
100 50
01 0. 0 1000
0. 03
0.1 5
0.0 6
6
150
0.2 5
250
5
0. 2
45 0.
200
35 0.
0. 3
Engine Torque (N.m)
350
0.4
Engine Torque (N.m)
Mass flow rate of HT ORC (kg/s) 350
3 150
3000
1. 5
100
0.6
0. 3
0 1000
4000
2000
3000
(a)
(b) Net power of LT ORC (kW)
300
300
250
0. 5
200
0.4 0.3 0.2
0 1000
0.0 4
250
3000
4000
11
7 150
5
3
100 50
2000
9
200
0.1
50
Engine Torque (N.m)
350
6 0.
Engine Torque (N.m)
Mass flow rate of LT ORC (kg/s) 350
100
4000
Engine speed (r/min)
Engine speed (r/min)
150
2
1 50
2000
4
200
1
0 1000
Engine speed (r/min)
2000
3000
4000
Engine speed (r/min)
(c)
(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. (10). 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 2600 r/min and the engine load is 100%. 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 30 °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 10 kW, the waste heat from the intake air exceeds 2 kW, and the intake air temperature is higher than 60 °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 8.0 [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-
H.G. Zhang et al. / Applied Energy 102 (2013) 1504–1513
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 0.485 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 6.975 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 0.646 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 11.913 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 89.45 kW, whereas the LT loop is 133.97 kW. In addition, the thermal efficiency of the HT loop is 7.80%, 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 18.89 kW at the rated power point, which improves the output power by 20%, 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 Improvement of engine effective power (%)
300
300
200
10
150
7
4
100 50 0 1000
1
2
200
15
250
18
16
16
13
17
250
150 100
19
14
20
50
2000
3000
2000
Engine speed (r/min)
3000
210
220
100
260
50 0 1000
2000
300
3000
13 15
13
200
14
19 0
200
250
14
300
Engine Torque (N.m)
300
0 20
4000
Improvement of bsfc (%) 350
18 4
Engine Torque (N.m)
bsfc of combined system (g/kW.h)
150
42
(b)
350
186
26 34
22
Engine speed (r/min)
(a)
250
30
38
0 1000
4000
17
350
Engine Torque (N.m)
Engine Torque (N.m)
Overall net power of ORCs (kW) 350
16
1512
.5
15
150 100
16
12
18 50
400 4000
0 1000
25 2000
3000
Engine speed (r/min)
Engine speed (r/min)
(c)
(d) Fig. 7. Performance of the combined system.
17
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H.G. Zhang et al. / Applied Energy 102 (2013) 1504–1513
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
_f m _ Wb
ð26Þ
and the bsfc of the combined system is defined as
bsfccs ¼
_f m _ _ _ n;LT W b þ W n;HT þ W
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 600 g/kW h to 400 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 25–30% 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 (30.5–54.5% of the total combustion energy) is slightly higher than the indicated power (22.4–44.4% 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 11.91 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.
1513
Acknowledgements This work was sponsored by the National Basic Research (973) Program of China (Grants #2011CB707202 and #2011CB710704), the National High-Tech Research and Development Program of China (863 Program) (Grant No. 2009AA05Z206), and the Funding Project for Academic Human Resources Development in Institutions of Higher Learning Under the Jurisdiction of Beijing Municipality (Grant No. PHR201008019). References [1] Fiaschi D, Manfrida G, Maraschiello F. Thermo-fluid dynamics preliminary design of turbo-expanders for ORC cycles. Appl Energy 2012;97:601–8. [2] Tempesti D, Manfrida G, Fiaschi D. Thermodynamic analysis of two micro CHP systems operating with geothermal and solar energy. Appl Energy 2012;97:609–17. [3] De Pascale A, Ferrari C, Melino F, Morini M, Pinelli M. Integration between a thermophotovoltaic generator and an organic Rankine cycle. Appl Energy 2012;97:695–703. [4] Clemente S, Micheli D, Reini M, Taccani R. Energy efficiency analysis of organic Rankine cycles with scroll expanders for cogenerative applications. Appl Energy 2012;97:792–801. [5] Roy JP, Mishra MK, Misra A. Performance analysis of an organic Rankine cycle with superheating under different heat source temperature conditions. Appl Energy 2011;88:2995–3004. [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 2011;36:3406–18. [7] Quoilin S, Lemort V, Lebrun J. Experimental study and modeling of an organic Rankine cycle using scroll expander. Appl Energy 2010;87:1260–8. [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 2011;88:2183–90. [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 2011;88:2740–54. [10] 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 2011;88:1500–9. [11] Qiu K, Hayden ACS. Integrated thermoelectric and organic Rankine cycles for micro-CHP systems. Appl Energy 2012;97:667–72. [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 2006-01-3522; 2006. [13] Freymann R, Strobl W, Obieglo A. The turbosteamer: a system introducing the principle of cogeneration in automotive applications. MTZ 2008;69:404–12. [14] Ringler J, Seifert M, Guyotot V, Hubner W. Rankine cycle for waste heat recovery of IC engines. SAE 2009-01-0174; 2009. [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 2012;43:385–95. [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 2010;53:2299–303. [17] Wang W, Wu YT, Ma CF, Liu LD, Yu J. Preliminary experimental study of single screw expander prototype. Appl Therm Eng 2011;31:3684–8. [18] Honyewell. GenertronÒ245fa applications development guide. Morristown, USA: Honeywell Fluorine Products; 2000. [19] Calm JM, Hourahan GC. Refrigerant data summary. Eng Syst 2001;18:74–88. [20] Bahadori A. Estimation of combustion flue gas acid dew point during heat recovery and efficiency gain. Appl Therm Eng 2011;31:1457–62. [21] National standard of the People’s Republic of China. GB/T 18297-2001; 2001. [22] Heywood JB. Internal combustion engine fundamentals. New York: McGrawHill; 1998. [23] On the web: