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Energy xxx (2013) 1e10

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Energy

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Simulation and thermodynamic analysis of a bottoming Organic Rankine Cycle(ORC) of diesel engine (DE)

Guopeng Yu, Gequn Shu*, Hua Tian, Haiqiao Wei, Lina LiuState Key Laboratory of Engines, Tianjin University, No. 92, Weijin Road, Nankai Region, Tianjin 300072, People’s Republic of China

a r t i c l e i n f o

Article history:Received 25 August 2012Received in revised form28 October 2012Accepted 29 October 2012Available online xxx

Keywords:Bottoming Organic Rankine Cycle (ORC)Diesel engine (DE)Waste heat recoverySimulationThermodynamic analysis

* Corresponding author. Tel.: þ86 022 27409558E-mail addresses: [email protected], [email protected]

0360-5442/$ e see front matter � 2012 Elsevier Ltd.http://dx.doi.org/10.1016/j.energy.2012.10.054

Please cite this article in press as: Yu G, et aengine (DE), Energy (2013), http://dx.doi.or

a b s t r a c t

This paper presents a simulation model based on an actual Organic Rankine Cycle (ORC) bottomingsystem of a diesel engine. The ORC system is built to recover waste heat both from engine exhaust gasand jacket water using R245fa as working fluid. Simulations and thermodynamic analyses are conductedto observe the influence of evaporating pressure and diesel engine (DE) conditions on system perfor-mance. Comprehensive evaluations are carried out on waste heat absorbing, expansion power, systemefficiency, exergy loss and exergy efficiency. The combined system of diesel engine with bottoming ORC(DE-ORC) is finally investigated. Results indicate that, approximately 75% and 9.5% of waste heat fromexhaust gas and from jacket water respectively can be recovered under the engine conditions rangingfrom high load to low load. The ORC system performances well under the rated engine condition withexpansion power up to 14.5 kW, recovery efficiency up to 9.2% and exergy efficiency up to 21.7%.Combined with bottoming ORC system, thermal efficiency of diesel engine can be improved up to 6.1%.

� 2012 Elsevier Ltd. All rights reserved.

1. Introduction

In a typical diesel engine, less than 45% of fuel energy might beconverted into useful work output from crankshaft, and theremaining energy is mainly lost through exhaust gas and jacketwater [1]. If the waste heat contained in exhaust gas and in jacketwater could be efficiently recovered and utilized, the efficiency ofthe original diesel engine would be significantly improved withoutadding any fuel. Among all the existed waste heat recovery tech-nologies, the Organic Rankine Cycle (ORC) is getting increasingattention with high efficiency, reliability and flexibility [14].

As one of the promising technologies of converting low-gradewaste heat into electricity, the ORC system has been studied fromdifferent aspects. Working fluids researches [2e4] mainly focus ondiverse screening and assessment criteria for dozens of organicfluid on performance of ORC system; performance analysis [5e8]focus on usable percentage of heat, output expansion power,recovery efficiency and exergy efficiency et al; System designs [9e11,30] based on scroll expanders, vapor injectors and dual loopORCs; optimizations [12e15] on parameters of turbine inlet pres-sure, evaporating temperature, pinch point temperature, heattransfer area et al. And the ORC technique has also been explored in

m (G. Shu).

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l., Simulation and thermodyng/10.1016/j.energy.2012.10.05

the engine field: Bianchi M and Pascale [16] revealed that the ORCtechnology results as the most performing and well proven solu-tion, in order to exploit low/medium temperature heat sourcessuch as ICEs. Tchanche et al. [17] presented existing applications ofORC and analyzes their maturity, including in car and engine field.Velez et al. [18] did a technical, economical and market review ofORC for waste heat recovery in ICEs as a part of their work. Bom-barda et al. [19] simulated and compared performances of ORC andKalina cycle on recovery of waste heat in exhaust gases from twodiesel engines, although obtained useful powers were actuallyequal, the Kalina cycle was less suitable than ORC because itscomplicated plant scheme, large surface heat exchangers, highpressure resistant and no-corrosion materials. Srinivasan et al. [20]examined the exhaust waste heat recovery potential of a dual fuellow temperature combustion engine using an ORC. With hot EGRand ORC turbocompounding, the fuel conversion efficiencyimproved by an average of 7 percentage points for all injectiontimings and loads while NOx and CO2 emissions recorded an 18percent (average) decrease. Teng et al. [21,22] simulated an ORC-WHR system to recover heat in exhaust gas, charge air cooler andEGR cooler. The case study showed up to 20% increase in enginepower. They also proposed a system recovering waste heat fromonly EGR (EGR-WHR system). The composite fuel savings over theESC 13-mode test is up to 5%. Vaja and Gambarotta [23] studiedthree configurations of ORC system for a 12-cylinder natural gasengine. Best fluid and configuration were selected and the

amic analysis of a bottoming Organic Rankine Cycle (ORC) of diesel4

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Table 1Five typical conditions of the diesel engine.

Parameters Condition1

Condition2

Condition3

Condition4

Condition5

Power output [kW] 258.3 235.8 211.6 176.2 117.7Fuel consumption

rate [kg/h]53.22 47.79 42.81 35.43 23.91

Temperature ofexhaust gas [K]

808.15 792.15 771.15 747.15 693.15

Mass flow ofexhaustgas [kg/s]

0.2981 0.2752 0.2586 0.2235 0.1697

Output temperatureof jacket water [K]

366.15 365.65 365.25 364.65 363.75

Mass flow of jacketwater [kg/s]

2.1 2.1 2.1 2.1 2.1

Fig. 1. Schematic diagram of the bottoming ORC.

G. Yu et al. / Energy xxx (2013) 1e102

maximum efficiency increase was about 12.5%. Boretti [24,25]explored the recovery of waste heat from exhaust gases and thecoolant of a H2ICE and a naturally aspirated gasoline engine insuccession. Hountalas and co-workers [26,27] researched onheavy-duty truck diesel engines using Rankine bottoming cycles.When exhaust heat and EGR heat were recovered, the improve-ment of brake specific fuel consumption (bsfc) ranged between 6%and 7.5% depending on engine loads. When CAC (Charge Air cooler)heat was also taken into account, maximum improvement in bsfcwas 11.3% with Organic Rankine Cycle and 9% with steam rankinecycle. Kane et al. [28] integrated solar energy with a combined cycleof a biodiesel enginewith two ORCs in cascade in a small pilot plant(10e25 kW), achieving overall efficiency of 41%.

So far, most researches on waste heat recovery of the engineusing ORC are theoretical researches and simulations [16e23], andonly a few experiments are developed, especially in engine field[24,25]. What’s more, most simulations are based on simple ther-modynamic models, i.e., ignoring detailed structures and operatingcharacteristics of system components including heat exchangersand expanders. So models like that are too ideal to precisely reflectthe real performance of the ORC. The system model proposed inthis paper is built up based on an actual experimental ORC set builtin our lab which serves as the bottoming system of a typical dieselengine. Main components are detailed modeled according to realproducts. The experimental ORC set contains an expander, twopumps, seven heat exchangers and several accessory equipments. Itis built to investigate the feasibility of waste-heat-recovery froma high duty diesel engine (DE). If proven in high practicability, it canbe adopted by generation sets in power plants and ICEs in ships,where the requirement for compactness is low, and waste heat arein high grade and quantity. After optimizations and downsizing, itis also expected to be used in vehicles in the future.

After validating from different angles, the model is rather closeto the real system, and the simulation results are more reliable topredict the system performance. Besides, five typical DE conditionsare investigated to reveal how the variable condition of the engineaffects the bottoming ORC system. Simulations and thermodynamicanalyses are conducted to evaluate waste heat absorbing, expan-sion power, system efficiency, exergy loss and exergy efficiency ofthe ORC. The combined system of DEwith bottoming ORC (DE-ORC)is finally analyzed to show its maximal working potential.

2. Description of systems

2.1. Description of diesel engine as the topping system

The diesel engine selected as topping system is an inline six-cylinder turbocharged engine used in a generator set. As a powerplant used engine, its speed is constant (1500 rpm) while its loadvaries under different conditions. Five important conditions of theengine are picked out in Table 1, and it generally runs under ratedcondition (condition 2). The thermal balance of the diesel engine isfirstly analyzed according to data from engine tests, as listed inTable 1. Under these five conditions of Table 1, the temperatures ofthe exhaust gas are within 690e810 K, and that of jacket water arewithin 363e366 K. Approximately 30e46% of the fuel energy iscontained in exhaust gas and 15% in jacket water. It is furtherconfirmed that, heat recovery of the two resources will significantlyreduce fuel consumption and improve engine efficiency.

It has been calculated that the air fuel ratio is 19.7 under the ratedcondition. Under the hypothesis of perfect combustion of diesel fuel,the composition of the exhaust gas onmass basis has been calculatedat: CO2¼15.1%,H2O¼ 5.5%,N2¼71.6%,O2¼7.8%. This composition isused to evaluate gas properties.Without considering trace additives,the jacketwater is simulatedaspurewater. It shouldbenoted that the

Please cite this article in press as: Yu G, et al., Simulation and thermodyengine (DE), Energy (2013), http://dx.doi.org/10.1016/j.energy.2012.10.05

water pump is driven by the engine crankshaft, whose speed isconstant. So the pump rotates uniformly and the flow of the jacketwater is basically unchanged at about 2.1 kg/s.

2.2. Description of ORC as the bottoming system

Figs. 1 and 2 show the principle schematic diagram and the Tesdiagram of the bottoming ORC for the diesel engine. As shown inthese two figures, the deep red line represents the exhaust gas flowfrom behind the engine to reject heat to the ORC system. The lightred loop circuit besides the engine is the cooling water circuit. Thegreen loop represents the ORC circuit. A thermal-oil circuit (in blackline) is used between exhaust gas and ORC circuit to preventdecomposition of working fluid R245fa. The whole system operatesas follows: hot exhaust gas from engine rejects heat to thermal-oilcircuit and then is discharged to atmosphere; working fluid R245faunder low pressure in liquid state (point 5) is firstly pumped intohigh pressure state (point 1), and then is preheated by engine jacketwater (point 2); after that, working fluid evaporates and is super-heated by hot thermal oil, thus working vapor under high pressureand temperature is generated (point 3); the vapor (point 3) flowsinto the turbine, and its enthalpy is converted into expansionpower; low pressure vapor (point 4) exits from turbine and flowsinto condenser where it is liquefied and condensed into saturatedliquid (point 5) by cooling water. Thus a whole cycle completes.Cycles run in this way to generate continuative power.

The actual ORC system contains an expander, an oil pump anda working fluid pump, a pre-heater, a gaseoil heat exchanger,

namic analysis of a bottoming Organic Rankine Cycle (ORC) of diesel4

Fig. 2. Tes diagram of the bottoming ORC.

G. Yu et al. / Energy xxx (2013) 1e10 3

condensing devices and oil-working fluid heat exchangers. Asshown in the two diagrams, oil-working fluid heat exchangers arethree exchangers connected in series (HE1, HE2, HE3). HE1 heatsworking fluid to saturated state, i.e. from point 2 to (2e3)a; HE2evaporates the working fluid from point (2e3)a to (2e3)b; HE3superheats working fluid from point (2e3)b to 3. Similarly,condensing process is also completed by two connected exchangers(CON1 and CON2). All the seven heat exchangers are plate counter-flow exchangers modeled in detail according to actual exchangercomponents. Main practical parameters of the exchangers are listedin Table 2. The exchangers are selected under the rated condition ofdiesel engine, and the designed evaporating pressure is supposedto be 30 bar. The expander is selected based on its actual perfor-mance curves, including the compression ratioevolume flow curve(Rpen) and the efficiencyevolume flow (hexpen) curve. Mainperformance parameters of the expander are listed in Table 3.R245fa is the working medium of the ORC system, and it is a non-chlorinated hydro fluorocarbon, non-ozone depleting liquid withlow global warming potential, good heat transfer ability, excellentthermal stability and low viscosity, and it is rather an appropriateworking fluid for the ORC system [15].

3. Modeling

3.1. Modeling of the ORC system

The systemmodel proposed in this paper is built up based uponan actual experimental ORC set which is now under construction in

Table 2Main parameters of plate exchangers in the ORC system.

Parameters Unit Pre-heater Gaseoilheat exc

Total heat transferarea

m2 1.06 35

Overall H T C. W/m2 K 1380 30Number of plates e 40 38Dimension of plates Length mm 287 2000

Width mm 117 500Thickness mm 2.3 2

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our lab. The experimental ORC set is built to investigate the feasi-bility of waste-heat-recovery from a high duty DE described above.According to the actual ORC bottoming system, the ORC model isconstructed by integration of ASPEN PLUS and ASPEN EDR(Exchanger Design & Rating) software from Aspen Technology, Inc.[29]. Modeling of heat exchangers is based on detailed structuralparameters listed in Table 2 which are provided by manufactures.By importing these reliable data and inlet values of working fluid,parameters like outlet values of working fluid, heat loads, heattransfer coefficients (HTCs) of heat exchangers are calculated out byASPEN EDR software. After validations, models of heat exchangerare quite reliable.

Modeling of expander is based on its performance curves aslisted in Table 3, ignoring the internal structure of this component.The expansion power (Pexp) can be calculated based on Eqs. (1) and(2) below, combing with isentropic efficiency (hexp) and compres-sion ratio (Rp) values in Table 3 tested by manufactures:

Pexp ¼ _m*ðh3 � h4Þ (1)

hexp ¼ ðh3 � h4Þ=ðh3 � h4sÞ (2)

wherein, _m is mass flow of the working fluid. The numbersubscripts used above are all illustrated in Fig. 2. By importing theRpen curve and hexpen of the actual expander, the model of theexpander is supposed to run as the real expander does. As for thetwo pumps contained e the oil pump and working fluid pump e

they are simply modeled, because their power consumptions are sosmall that they have little effect on system performance, and theirefficiencies are set the same, at 0.8. The components modeledabove are connected by thermal-oil stream and R245fa streamcorrectly and the ORC system model is then completed.

3.2. Modeling for ORC thermodynamic analysis

Firstly, a proper systemmodel should be in good energy balance.Namely, energy output should match well with energy input, asdemonstrated in Eq. (3).

jQin � Qoutj ¼ 0 (3)

Qin is the total energy that enters ORC system, while Qout is thatleaves the system. Neglecting the energy loss in actual components,pipes and oil pump, energy balance can be extended to be Eq. (4).

��Qexh þ Qjw þ Pwf � Qcond � Pexp�� ¼ 0 (4)

Qexh and Qjw are the waste heats absorbed from exhaust gas andfrom jacket water; Pwf is the electric energy consumed by workingfluid pump. Qcond is the heat released from working vapor tocooling water in condensation. Pexp is the output expansion power.

hangerHE1 HE2 HE3 CON1 CON2

1.45 1.76 5.28 1.10 4.08

1280 520 180 882 243025 30 50 40 70

526 526 316 287 526119 119 316 117 119

2.2 2.2 1.2 2.3 2.2


Table 3Main performance parameters of the expander in the ORC system.

n e Volume flow [10�3 cum/s] 4.5 3.9 3.7 3.5 3.3 2.9 2.7 1.9 1.5 1.0 0.4hexp e isentropic efficiency [/] 0.45 0.5 0.54 0.55 0.60 0.51 0.50 0.48 0.48 0.47 0.46Rp e compression ratio [/] 10.30 10.90 11.40 11.60 11.70 11.68 11.67 11.63 11.63 11.62 11.60


The recovery efficiency is generally regarded as the 1st lawefficiency, i.e. the ratio of expansion power to the heat absorbed bythe ORC system from the heat resources.

h1st ¼ PexpQexh þ Qjw

(5)

Efficiency of the DE-ORC combined system (ha) is the ratio oftotal output mechanical power to total fuel energy. Total outputmechanical power is the sum of the original diesel engine power(Po) and the ORC expansion power (Pexp).

ha ¼ Pexp þ PoQfuel

(6)

The 2nd law analysis is then carried out to explore the exergyutilization and exergy loss within the ORC system [5]. Thesubscripts used below are all illustrated in Fig. 2. Eq. (7) shows howto calculate the exergy flow change of a certain material stream(stream i) when it flows through a certain component. Amongthem, _Ei is the exergy rate of stream i under a certain state and _mi isthe mass flow of this stream. In the following analysis, the deadstate temperature is assumed to be 290 K (T0).

_Ei ¼ _Ei�IN � _Ei�OUT ¼ ½ðhi�IN � hi�OUTÞ � T0ðsi�IN � si�OUTÞ�$ _mi

(7)

The net exergy rate of the exhaust gas entering the system is _EEXH

_EEXH ¼ _EEXH�IN � _EEXH�OUT

¼ ½ðhEXH�IN � hEXH�OUTÞ � T0ðsEXH�IN � sEXH�OUTÞ�$ _mEXH

(8)

The net exergy rate of the jacket water entering the system is _EJW

_EJW ¼ _EJW�IN � _EJW�OUT

¼ ��hJW�IN � hJW�OUT

�� T0�sJW�IN � sJW�OUT

��$ _mJW (9)

The total exergy rate entering the system is _EA

_EA ¼ _EEXH þ _EJW þ Pwf (10)

Neglecting small amount of exergy loss around the oil pump, theexergy loss in the heat exchanging process between exhaust gasand working fluid ð_IEXH�FÞ is considered to be the sum of exergylosses in the following four heat exchangers: Gaseoil heat exh-changer, HE1, HE2 and HE3.

_IEXH�F ¼�_EEXH�IN � _EEXH�OUT

��_E3 � _E2

�(11)

Similarly, exergy loss in the condensing process betweenworking fluid and cooling water ð_ICONDÞ is considered to be the sumof exergy losses in the following two heat exchangers: COND1 andCOND2.

_ICOND ¼�_E4 � _E5

��_ECW�OUT � _ECW�IN

�(12)


The exergy loss in the preheat process between jacket water andworking fluid _IPRE�HE:

_IPRE�HE ¼�_EJW�IN � _EJW�OUT

��_E2 � _E1

�(13)

Exergy losses that are carried out through output cooling water_ICW:

_ICW ¼ _ECW¼ ½ðhCW�OUT � hCW�INÞ � T0ðsCW�OUT � sCW�INÞ�$ _mCW

(14)

Exergy losses due to non-isentropic vapor expansion in theexpander _IEXP:

_IEXP ¼�_E3 � _E4

��_H3 � _H4

�¼ T0ðs4 � s3Þ$ _m (15)

Exergy losses due to non-isentropic compression of workingfluid pump _IP:

_IP ¼ T0ðs1 � s5Þ$ _m (16)

In summary, total exergy that leaves the ORC system ð_IÞ is definedas Eq. (17):

_I ¼ _IEXH�F þ _IPRE�HE þ _IEXP þ _ICOND þ _ICW þ _IP (17)

The second-law efficiency of the power cycle, also referred to asexergy efficiency (h2nd), can be defined as follows:

h2nd ¼ Pexp

E�A

(18)

4. Model validations

4.1. Mass-balance and energy-balance validation

No leakage is considered in the systemmodel. All themass flowsof materials contained remain unchanged, and the mass of thesystem is thus in good balance.

The energy-balance validation is reflected in the Eq. (4). Underthe rated engine condition, system energy-balance calculationswithin 25e33 bar evaporation pressure ranges are listed in Table 4.Energy budget under the 9 pressures all fit Eq. (4) quite well, andmaximum deviation is only 0.23 kW. Additionally, energy budget ofthe other four engine conditions are calculated, revealing that theenergy of the system is well balanced.

4.2. Model validated with heat exchanger products

The heat exchanger products in the real ORC system are selectedand bought under the hypothesis that, the diesel engine runs underthe rated condition, and evaporating pressure is 30 bar, the massflow of working fluid is 0.34 kg/s. With the same assumption,simulation is conducted on the basis of detailed modeling, andsimulating results (temperatures and pressures) are comparedwithproduct parameters supplied by heat exchanger manufacturers.These product parameters are design data of the exchangers whichare enough reliable. As shown in Fig.3, the comparison revels that


Table 4Validation of energy balance of the WHR system under rated engine condition.

25 bar 26 bar 27 bar 28 bar 29 bar 30 bar 31 bar 32 bar 33 bar

Qexhaust [kW] 118.23 118.12 118.09 118.06 118.04 118.02 118.02 118.03 118.03Qjacket-water [kW] 28.17 28.59 28.57 28.93 29.14 29.36 29.47 29.60 29.61Ppump [kW] 0.72 0.75 0.78 0.81 0.84 0.87 0.89 0.92 0.95Qcondens [kW] 135.94 135.52 134.42 134.60 134.43 133.70 133.96 134.67 135.53Pexpansion [kW] 11.07 11.80 12.80 13.25 13.47 14.45 14.48 13.75 13.07jEin � Eoutj [kW] 0.11 0.16 0.23 0.03 0.13 0.11 0.05 0.14 0.01


two sets of data are in good agreement at these main points. Thetick labels in X-axis of Fig. 3 are inlet or outlet points of exchangersin the ORC system, as pointed out in Figs. 1 and 2. Maximumtemperature deviation is in point 1, where the simulated temper-ature is 11.7 K lower than that of the actual pre-heater exchanger,and the absolute deviation is about 3.8%. Maximum pressuredeviation is at point CW-OUT, where the simulated pressure is0.01 bar lower than that of COND1 exchanger, and the absolutedeviation is about 3.1%. The exchangers are thus well validated. Theentire ORC system model is now proven reliable and accurate.

5. Results and discussion

5.1. Waste heat absorbing

In the proposed ORC system, waste heat is absorbed from boththe exhaust gas and the jacket water of the diesel engine. As is

1 2 3 4 5 CW-IN CW-MID CW-OUT JW-IN JW-OUT EXH-IN EXH-OUT

300

400

500

600

700

800

Tem

pera

ture

[K]

Main points in the ORC sytem

Simulating resultsParameters of heat exchangers from manufacturers

a

b

1 2 3 4 5 CW-IN CW-MID CW-OUT JW-IN JW-OUT EXH-IN EXH-OUT

0

5

10

15

20

25

30

Main points in the ORC sytem

Simulating resultsParameters of heat exchangers from manufacturers

Pre

ssur

e[ba

r]

Fig. 3. Comparison between simulating results and heat exchangers’ real parameterson (a) temperatures and (b) pressures.


shown in Fig. 4, the total heat absorbed by the ORC, and the heatrecovered from exhaust gas and from jacket water all vary notmuch with changing of evaporating pressure. The highest fluctua-tion is less than 2 kW. So it is sensible to replace these three items ofheat with their average values within the variation of evaporatingpressure. Then it is calculated that, under the rated engine condi-tion, the average amount of total heat absorbed is about 147 kW,the heats absorbed from exhaust gas and from jacket water are118.1 kWand 29 kW, accounting for 4/5 and 1/5 of the total amountrespectively.

In order to further clarify heat recovery capability of the ORCsystem, the concept of waste heat utilization rate (UR) is proposed.The waste heat UR is the ratio of heat absorbed from a certainsource by the ORC system to maximum heat available in this heatsource. As for the two heat sources involved in this paper, two URsare concerned, i.e. waste heat URs of exhaust gas and of jacketwater. It should be noted that, the maximum heat available inexhaust gas is the heat rejected under the hypothesis that the gas iscooled to 298 K, and 333 K is set for jacket water. Then the URs ofexhaust gas and of jacket water are 73.7% and 9.3% under the ratedengine condition.

Similar calculations are done for all five engine conditions. Theresults are listed in Table 5. The reduction of the engine load (fromcondition 1 to condition 5) surely results in the reduction intemperature and in mass flow of the heat resources e exhaust gasand jacket water. With these reductions, total heat absorbed by theORC decreases significantly, as shown in the 1st column in Table 5.From the 2nd and the 3rd columns, the average ratio of heatrecovered from exhaust gas keeps falling, while that of jacket waterkeeps rising. The reason can be found from Table 1. It shows thatwith the reduction of engine load, the temperature and mass flowof exhaust gas decrease greatly, but those of jacket water decreaseonly slightly. That is to say, the heat of exhaust gas decreases muchfaster than the heat in jacket water does. It then causes similarchanges in the heat absorbed. But in general, heat absorbed from

24 26 28 30 32 3426

27

28

29

30

31

Evaporating Pressure[bar]

Hea

t rec

over

ed f

rom

jack

et w

ater

[KW

]

117.0

117.5

118.0

118.5

119.0

Heat recovered from

exhaust gas[K

W]

145

146

147

148

149

Tot

al a

mou

nt o

f he

at r

ecov

ered

by

the

OR

C

sys

tem

[K

W]

Fig. 4. Variation of heat absorbed from exhaust gas and from jacket water, and totalheat absorbed by the ORC system with evaporating pressure under the rated DEcondition.


Table 5Waste heat recovery of the ORC system under different engine conditions.

Total heat recoveredof the ORCsystem [kW]

The average ratioof heat recoveredfrom jacket water [%]

The average ratio ofheat recovered fromexhaust gas [%]

The average UR forwaste heat in jacketwater [%]

The average UR forwaste heat inexhaust gas [%]

Condition 1 159.38 18.40 81.60 9.25 72.29Condition 2 147.12 19.74 80.26 9.29 73.68Condition 3 136.51 21.77 78.23 9.61 74.46Condition 4 117.42 25.66 74.34 9.92 74.64Condition 5 88.00 33.29 66.71 9.90 76.19


exhaust gas is all the time much more than that from jacket water.Because high grade heat in exhaust gas is definitely easier to berecovered than low grade heat in jacket water. From the last twocolumns in Table 5, URs of waste heat in exhaust gas and in jacketwater all change slightly, the value of the former is around 75%, andthat of the later is about 9.5%. It can be concluded that, the ORCsystem can recover heat in exhaust gas effectively, but can’t behavewell in absorbing waste heat in jacket water.

5.2. Expansion power and recovery efficiency

As shown in Fig. 5, the expansion power and recovery efficiencychange in the same way, because heat absorbed by the ORC systemis almost constant under certain engine condition, thus thechanging regularity of expansion power determines that of thesystem efficiency. Both of them appear to increase first and thendecrease with increasing evaporating pressure. The maximumexpansion power and recovery efficiency are gained under thepressure of around 31 bar and are about 14.5 kW and 9.2%. Thisregularity is explained as follows: when the evaporating pressureincreases from 25 bar to 33 bar, the volume flow of inlet stream(stream3) of expander reduces gradually from 4.5e�3 cum/s to4.0e�4 cum/s. When the evaporating pressure is around 31 bar, thevolume flow is about 3.3e�3 cum/s. From Table 3, the pressure ratioof expander (the ratio of inlet pressure to outlet pressure) reachesits maximum value of about 11.7, meanwhile the isentropic effi-ciency is as high as 0.6. That means, enthalpy difference betweeninlet stream and outlet stream of the expander is maximal, andexergy loss is relatively low, so maximum expansion power isoutput. When evaporating pressure is below 30 bar, pressure ratioof expander is smaller, the enthalpy difference is smaller and thusexpansion power is reduced. When evaporating pressure is above

24 26 28 30 32 3410.0

10.5

11.0

11.5

12.0

12.5

13.0

13.5

14.0

14.5

15.0

Recovery E

fficiency[%]

Exp

ansi

on P

ower

[KW

]


7.0

7.5

8.0

8.5

9.0

9.5

10.0

Fig. 5. Variation of expansion power and recovery efficiency with evaporating pres-sure under the rated DE condition.


31 bar, though expander’s pressure ratio is still high, its isentropicefficiency is as low as only 0.46, which means the exergy lossincreases and the expansion power is also reduced.

It can be concluded from Fig. 6 that peak value of expansionpower under each engine condition moves to the left side (lowpressure direction) gradually with decreasing of engine load (fromcondition 1 to condition 5). Under the condition 5 and condition 4,expansion power decreases constantly with the increase of evap-orating pressure, and themaximum power they get are only 7.2 kWand 11.3 kW. The reason is that the quantity and quality of the heatwithin heat resources are quite low under these two conditions;under condition 3 and condition 2, expansion power appears toincrease firstly and then decrease with the increase of evaporatingpressure. The maximum power of condition 3 is gained whenevaporating pressure is 28 bar, which is about 13.7 kW. While forcondition 2 with a higher load, maximum power is obtained whenevaporating pressure is about 30 bar, which is about 14.5 kW; whendiesel engine runs under condition 1 with high load, the expansionpower increases constantly with the increase of evaporating pres-sure and it is up to 15.5 kW when the evaporating pressure is33 bar. The reasons of these regularities are the same with that isexplained for Fig. 5. Namely, they are all decided by the combinedeffect of the pressure ratio and isentropic efficiency of theexpander.

As for the recovery efficiencies in Fig. 7, they change similarlywith corresponding expansion power. With the decreasing ofengine load from condition 1 to condition 5, maximum values are9.1%, 9.2%, 9.4%, 9.1% and 7.4% successively. They do not reducegradually as the peaks of expansion power do. Take condition 2 andcondition 3 for instance, their peak expansion power are lower thanthat of condition 1, but their peak recovery efficiencies are higher

24 26 28 30 32

7

8

9

10

11

12

13

14

15

16

Exp

ansi

on P

ower

[KW

]

Evaporating Pressure[KW]

Condition 1 Condition 2 Condition 3 Condition 4 Condition 5

Fig. 6. Variation of expansion power with evaporating pressure under five DEconditions.


Fig. 7. Variation of recovery efficiency with evaporating pressure under five DEconditions.

Fig. 9. Influence of DE conditions on exergy flow losses of the ORC system.


than that of condition 1, since recovered heat under these twoconditions are much less. Overall, the ORC system gets relative highexpansion power and recovery efficiencies under conditions 1, 2and 3.

5.3. Exergy flow losses and exergy efficiency

An appropriate exergy analysis will give a better understandingof the irreversibility of the ORC system, and is indispensable fora complete evaluation of the whole ORC system.

Six items of exergy flow losses of the ORC system under therated engine condition is plotted in Fig. 8. Orderly, they are (1)exergy flow losses in the heat exchanging process between exhaustgas and working fluid _IEXH�F; (2) exergy flow losses in the pre-heating process between jacket water and working fluid _IPRE�HE;(3) exergy flow losses due to non-isentropic vapor expansion in theexpander _IEXP; (4) exergy flow losses in the condensing processbetween working fluid and cooling water _ICOND; (5) exergy flowlosses that are carried out through output cooling water _ICW; (6)exergy flow losses due to non-isentropic compression of workingfluid pump _IP. Among the six items of exergy flow losses, _IEXH�F is

24 26 28 30 32 340

5

10

15

20

25

30

35

40

Exe

rgy

Flo

w L

osse

s un

der

The

R

ated

Con

diti

on [

KW

] EXH-F PRE-HE EXPANDER CONDENSE CW PUMP


Fig. 8. Exergy flow losses of the ORC system under the rated DE condition.


the highest. One reason is that four heat exchangers (gaseoil heatexchanger, HE1, HE2 and HE3) are included in the heat exchangingprocess between exhaust gas and working fluid, and the sum oftheir exergy flow losses is surely high. On the other hand,temperature differences in this process are rather great, resulting insevere irreversibility. The second and the third highest exergy flowlosses are _ICOND and _IEXP. The other three items of exergy flow lossesare as low as below 10 kW. Besides, under the other four engineconditions, the rankings are the same as under the rated condition.So it is critical to reduce these first three items of exergy losses tolower the irreversibility of the ORC system.

After exergy calculations of all the five engine conditions, itshows that, each itemof exergyflowloss does not changemuchwithincreasing of evaporating pressure, just as shown in Fig. 8 under therated condition. So the average values are calculated to replace theitems of exergy flow losses correspondingly. For example, the abovesix items of exergy flow losses of the ORC system under the ratedengine condition are replaced with the average value 20.6 kW,2.7 kW, 6.6 kW, 19.0 kW, 4.1 kW and 0.2 kW respectively.

Then average values under each engine condition are gatheredin Fig. 9, showing the influence of engine conditions on averageexergy flow losses of the ORC system. Overall, the reduction of

Fig. 10. Variation of exergy efficiency with evaporating pressure under five DEconditions.


Fig. 11. Increment of thermal efficiency by integrating the bottoming ORC system toDE.


engine load (from condition 1 to condition 5) results in down-grading of waste heat of exhaust gas and of jacket water (especiallyof the exhaust gas), and thus exergy from the heat resources falls,and total exergy flow loss of the ORC system drops consequently. Asis shown in Fig. 9, obvious drops occur in _IEXH�F, _ICOND and _ICW.Because of the reduction of engine load, the temperature differencebetween the exhaust gas and the working fluid decreases, andtherefore _IEXH�F decreases. On the other hand, with downgrading ofwaste heat, the highest temperature e the inlet temperature infront of the expander e of the working vapor falls and then thetemperature after expansion falls. Then the temperature differencebetween the working vapor and the cooling water decreases,therefore the _ICOND decreases. Decreases of _IEXH�F and _ICOND are themain reasons for the drops of the total exergy flow loss. Besides,since less heat is condensed and the outlet temperature of the

Fig. 12. Energy distribution of the DE a


cooling water is lower, the _ICW also drops a lot. As for _IPRE�HE, italmost keeps constant, since the temperature of outlet jacket waterdoes not change much, and heat exchanging within the pre-heateris stable. Also, _IEXP and _IP do not vary much, because their entropydifferences between inlet and outlet steams are very small as seenfrom the calculating results.

As explained above, heat absorbed by the ORC system is almostconstant under certain engine condition, so is exergy. Therefore, thechanging regularity of expansion power determines the changingtrends of both the recovery efficiency and the exergy efficiency, sothe regularities of the expansion power, recovery efficiency andexergy efficiency are quite similar under certain engine conditionby comparing Figs. 6, 7 and 10. But the difference occurs at the peakvalues, and with decreasing of engine load from condition 1 tocondition 5, peak values of exergy efficiencies are 20.9%, 21.7%,22.9%, 23.2% and 22.3% successively, as shown in Fig. 10. Undercondition 4 and condition 5, the expansion power and recoveryefficiencies are not high, but their exergy efficiencies are quite high,because waste heats under these conditions are not much and theexergy entered the system are low, which are only 49 kW and33 kW, thus high exergy efficiencies are obtained. Oppositely, incondition 1, the expansion power and recovery efficiency are ratherhigh, but its exergy efficiency is low, because large amount ofexergy (about 74 kW) is absorbed into the system. Besides, exergyefficiencies of condition 2 and condition 3 are still high. In a word,by analyzing Figs. 6e8, the overall performance of the ORC systemincluding its expansion power, recovery efficiency and exergyefficiency are relatively higher in condition 2 and condition 3 thanthat in other three conditions.

5.4. Performance of the DE-ORC combined system

Performance of the combined DE-ORC system is firstly showedin Fig. 11 by variation of increments in thermal efficiency of thecombined system to the original diesel engine. With the increase of

nd the DE-ORC combined system.



evaporating pressure, increments appear to rise constantly incondition 1, increase firstly and then decrease in condition 2 and incondition 3, and decrease constantly in condition 4 and condition 5.Maximal efficiencies under each condition are pointed out andconnected to show the greatest working potential of the ORCsystem. They are, 5.7% of the increment in condition 1 under 33 bar,5.8% in condition 2 under 30 bar and 31 bar, 6.1% in condition 3under 28 bar, 6.0% in condition 4 under 25 bar and 5.5% in condition5 under 25 bar. Besides, average values of increment under eachcondition are calculated and noted, they are 4.8%, 5.2%, 5.5%, 5.3%,5.2% from conditions 1 to 5.

Fig. 12 reveals the energy distribution before and after theintegration of the bottoming ORC to the DE by gathering theaverage value of each itemwithin the range 25e33 bar. The pumppower and other secondary energy losses are ignored. As markedin the left side, the DE power accounts for only around 40% of thefuel energy in the five conditions. Most fuel energy is lost throughexhaust gas and jacket water, especially in low load condition. Asmarked in the right side, the integration of the bottoming ORCmodifies the original energy distribution. Expansion powergained by the ORC increase the useful power without adding fuel,and the thermal efficiency is thus uplifted. It is very meaningfulin this bottleneck period of the engine efficiency lifting. As shownin the figure, heat recovery of exhaust gas is much more efficientthan that of the jacket water. More suitable system should beexplored in the future work to recover more low-grade heat fromjacket water, for example the dual loop ORC [30]. It also showsthat only a little energy of the heat absorbed converts intoexpansion power, the remaining is lost mainly throughcondensing. So the ORC system should be optimized to be moreefficient in the future.

6. Conclusions

A detailed ORC system is modeled and validated for a dieselengine to recover waste heat in exhaust gas and in jacket water.Comprehensive thermodynamic analyses are conducted and mainconclusions can be drawn as follows:

(1) The ORC system can recover heat in exhaust gas effectively, butbehaves badly in recovering heat in jacket water. URs for wasteheat in exhaust gas and in jacket water are about 75% and 9.5%.

(2) Expansion power and recovery efficiency change similarlyunder each engine condition. The ORC system gets relativelyhigh power (15.5 kW, 14.5 kW and 13.7 kW) and efficiencies(9.1%, 9.2% and 9.4%) under conditions 1, 2 and 3.

(3) The first three exergy losses occur in the heat exchangingprocess between exhaust gas and working fluid, in thecondensing process between working fluid and cooling waterand in the non-isentropic vapor expansion in the expander.

(4) The overall performance of the ORC system including itsexpansion power, recovery efficiency and the exergy efficiencyare relatively higher in condition 2 and condition 3 than that inthe other three conditions left.

(5) Up to 6.1% of increment in thermal efficiency can be obtainedby the combined DE-ORC system. If given the rated condition(condition 2), up to 5.8% of increment can be acquired bysetting the evaporating pressure at 30 bar or 31 bar.

Acknowledgements

This work was supported by a grant from the National NaturalScience Foundation of China (No. 51206117), and the National BasicResearch Program of China (973 Program) (No. 2011CB707201).


Nomenclature

AbbreviationsORC Organic Rankine CycleDE diesel engineICE internal combustion engineWHR waste heat recoveryEGR exhaust gas recirculationCAC charge air coolerbsfc brake specific fuel consumptionEDR exchanger design and ratingUR utilization rate (of waste heat)CR compression ratioHTC heat transfer coefficientcum cubic meter

SymbolsR ratio [/]P power [kW]Q exchanging heat [kW]h efficiency [%]_E exergy flow [kW]_I exergy (flow) loss [kW]T temperature [K]s specific entropy [kJ/kg]h specific enthalpy [kJ/kg]_m mass flow [kg/s]

Subscriptsp compressionexp expander/expansionin inletout outletexh exhaust gasjw jacket watercw cooling waterwf working fluidcond condenser/condensation

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