multidimensional modeling of the effect of egr (exhaust gas recirculation) mass fraction on exergy...
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Energy 66 (2014) 305e313
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Energy
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Multidimensional modeling of the effect of EGR (exhaust gasrecirculation) mass fraction on exergy terms in an indirect injectiondiesel engine
Samad Jafarmadar*
Mechanical Engineering Department, Technical Education Faculty, Urmia University, Urmia, West Azerbaijan 57561-15311, Iran
a r t i c l e i n f o
Article history:Received 30 April 2013Received in revised form8 January 2014Accepted 10 January 2014Available online 22 February 2014
Keywords:Exergy analysesIndirect injectionEnergyThree dimensional modelingEGR (exhaust gas recirculation)
* Tel.: þ98 441 2972000; fax: þ98 441 2773591.E-mail addresses: [email protected], s_jafar
0360-5442/$ e see front matter � 2014 Elsevier Ltd.http://dx.doi.org/10.1016/j.energy.2014.01.040
a b s t r a c t
In this investigation, the energy and exergy analyses are carried out for a Lister 8.1 IDI (indirect injection)diesel engine at four different EGR (exhaust gas recirculation) mass fractions (0%, 10%, 20% and 30%) andat 50% load operation. The energy analysis is performed during a closed cycle by using a three-dimensional CFD (Computational Fluid Dynamics) code. For the exergy analysis, an in-house computa-tional code is developed, which uses the results of the energy analysis at different EGR mass fractions.The cylinder pressure results for baseline engine are compared with the corresponding experimentaldata that shows a good agreement. With crank position at different EGR mass fractions, various exergycomponents and the cumulative exergy are identified and calculated separately. It is found that at 50%load operation, as EGR mass fraction increases from 0% to 30% (in 10% increments), exergy efficiencydecreases from 31.74% to 25.38%. Also, the cumulative irreversibility related to the combustion chamberdecreases from 29.8% of the injected fuel exergy to 25.5%. This work demonstrates that multidimensionalmodeling can be used to simulate the effect of various EGR mass fractions and gain more insight into theimpact of flow field on combustion process in IDI engines from the second law perspective.
� 2014 Elsevier Ltd. All rights reserved.
1. Introduction
DI (direct injection) or IDI (indirect injection) diesel engines arewidely used for transportation, automotive, agricultural and in-dustrial applications because of their high thermal, fuel conversionefficiencies, and easy operation. In a very competitive world, theimprovement of engine performance has become an importantissue for automotive manufacturers. In order to improve engineperformance, the combustion and emission processes are studiedmore thoroughly these days by simultaneously applying the firstand second laws of thermodynamics. Exergy is the key concept inthe second law analysis; it is a special case of themore fundamentalconcept, the available energy, which has been introduced in Ref. [1].For analyzing the performance of engine subsystems, exergy anal-ysis can be a useful alternative to energy analysis, because it is ableto reveal more information about engine processes [2]. Dunbaret al. [3] reported that almost 33% of the energy of fossil fuels aredemolished during the combustion process in power generation.Dunbar and Lior [4] applied the second law of thermodynamics
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analyses to combustion reactions of methane and hydrogen fuels.They have explored four paths for irreversibility. These paths aresimilar to alternative chamber designs and arrangements toexplore their effect on combustion irreversibilities. They haveshowed that there are four primary sources of combustion irre-versibility and that these vary depending on numerous factors. Theresults show that combustion irreversibility increases withincreasing excess air. Furthermore, a percentage breakdown of eachsources contribution to the overall combustion irreversibility wascalculated for each of the various paths.
The result of thework conducted by Rakopoulos and Andritsakis[5] in a turbocharged diesel engine shows that the value of com-bustion irreversibilities can be correlated to the differential changein mixture composition, and notably nothing else. They calculatedthe rate of combustion irreversibility as a function of fuel reactionrate only.
Rakopoulos and Giakoumis [6] have performed exergy analysesin diesel engines at transient engine operations. They haveexplored interesting aspects of the transient engine operation fromsecond law of thermodynamics viewpoint. Further studying aboutirreversibility in internal combustion engines were conducted inRefs. [7]and [8]. A summary of other studies on the subject isprovided below:
S. Jafarmadar / Energy 66 (2014) 305e313306
Rakopoulos and Kyritsis [9] carried out a numerical studyingabout the effects of cylinder wall temperature on the exergy tran-sient performance of an IDI turbocharged multi-cylinder dieselengine at various loads, which had a special emphasis on the case oflow heat rejection. They have showed that the transient first-lawproperties are almost unaffected by the applied wall temperaturescheme, while the second-law terms of engine and turbochargerare greatly affected, especially when a low heat rejection cylinderwall is chosen.
Rakopoulos and Giakoumis [10] carried out an exergy analysisfor a multi-cylinder turbocharged diesel engine and all its com-ponents by using a single-zone thermodynamic model. They havedemonstrated that exergy analysis offers a comprehensive insightinto the processes occurring in a diesel engine than its traditionalfirst-law counterpart.
Ghazikhani et al. [11] carried out an experimental study about theeffect of EGR (exhaust gas recirculation) onvarious exergy termsof anIDI diesel engine cylinder. Their results indicated that the applicationof EGR to engine mainly increases the total in-cylinder irreversibilitydue to the extension of the flame region, thereby raising the com-bustion temperature. Also, the results revealed that the behaviors ofthe total in-cylinder irreversibility and engine BSFC (Brake SpecificFuel Consumption) are the same, especially at high load conditions.
Rakopoulos and Giakoumis [12] carried out an exergy analysis inan IDI engine and all of its subsystems including the compressor,after-cooler, inlet manifold, cylinders (for both the closed and openparts of the cycle), exhaust manifold and the turbine. Variousexergy terms were calculated during a transient event such aswork, heat transfer, exhaust gas and irreversibility. In particular, theirreversibility components of every transient cycle were calculatedfor the diesel engine and its subsystems. Also, for the sake ofcomparison, the rate and the cumulative value of all the importantexergy components were given for the first and last cycles of thetransient event. The importance of combustion irreversibility aswell as exhaust manifold irreversibility was revealed.
Amjad et al. [13] used a single-zone model to perform a nu-merical availability analysis of the combustion of n-heptane andnatural gas blends in HCCI engines. They showed that as the masspercentage of natural gas in the fuel blend increases, irreversibilitydecreases and the second-law efficiency increases. Adding the EGRto the intake charge of the dual-fuel HCCI engine, up to an optimumvalue, enhances the exergy efficiency. EGR values above this pointcould deteriorate engine performance.
Hosseinzadeh et al. [14] carried out a numerical study bycomparing the chemical, thermal and radical effects of EGR gasesusing a single-zone model to analyze exergy in dual-fuel enginesoperating at part loads. They showed that the chemical effect ofEGR causes an increase in the unburned chemical exergy and adecrease in the work exergy, in comparison with the baseline en-gine (without EGR); while the thermal and radical effects havepositive effects on the exergy terms, especially on the unburnedchemical and work exergies. They also demonstrated that by usinglow values for the radical and thermal components of EGR, theexergy efficiency increases.
Turan [15] studied exergetic effects of some design parameterson the small turbojet engine for unmanned air vehicle applications.In this article the effect of some design parameters such ascompressor pressure ratio and turbine inlet temperature wasanalyzed on the exergetic and energetic performances for a smallturbojet engine for unmanned vehicles. Studying these parametersindicates showed that significant improvement is possible for thesmall turbojet engine to achieve better energy and exergy con-sumption by optimization of these parameters.
Rakopoulos and Giakoumis [6] carried out an exergy analysis oftransient diesel-engine operation. They showed that the exergy
properties of the diesel-engine subsystems vary according to theengine cycles for various speed and load changes. Also, the effect ofoperation parameters such as intensity of heat loss to wall aredescribed from first- and second-law perspectives.
Jafarmadar [16] carried out a numerical exergy analysis in thechambers of an IDI engine by three-dimensional code. Thecomputational results show that at 50% and full load operations,77.4% and 55.7% of total irreversibility concern to combustion inmain chamber, respectively. Also, 89% and 86% of burned fuelexergies and also, 65.7% and 55% of total heat loss exergies concernto burn fuel and heat loss in the main chamber at 50% and full loadoperations.
Jafarmadar et al. [17] carried out a numerical analysis about theexergy analysis in a LHR (low heat rejection) IDI diesel engine bythree-dimensional modeling. The comparison of the results forbaseline and LHR cases shows that when the load increases from25% to 100% (in 25% increments), heat loss exergy decreases by68.73%, 80.24%,91.38% and 74.97% in LHR engine in comparison tobaseline engine. Also, exergy efficiency increases by 17.2%, 12.4%,6.07% and 11.81% in LHR engine.
Jafarmadar and Zehni [18] carried out a numerical analysisabout the effect of dwell duration in a split injection scheme on theexergy terms of an IDI diesel engine. The results show that the valueof exergy efficiency decreases when the dwell duration is changedfrom 5�CA to 30�CA. Also, there is a sharp variation in the exergyparameters when the dwell time reaches 25�CA.
Balli and Hepbasli [19] investigated the performance of T56turboprop engine from exergoeconomic, sustainability and envi-ronmental damage cost viewpoints at different power loadings. Theresults of optimizationwere presentedwithmore details for energyand exegy analyses.
The research of Motasemi et al. [20] on the transportation sectorof Canada that which is the second largest energy consuming sectorwith 30% of the total energy consumption of the country in2009,showed that Energy and exergy efficiencies may reach 20.95%and 20.97% in the year 2035 respectively based on the forecasteddata. The purpose of this work was to analyze the energy, exergy,and emission performance for four different modes of transport(road, air, rail, and marine) from the year 1990e2035.
As can be seen in relevant literature, no attempt has been madeso far to three-dimensionally study the effect of EGR mass fractionon various exergy terms in IDI diesel engines from the second-lawviewpoint. In the present work, a CFD (Computational Fluid Dy-namics) code for combustion modeling along with an in-housecode for exergy analyses have been used to predict various exergyterms at 0%, 10%, 20% and 30% EGR mass fractions in the Lister 8.1indirect injection diesel engine operating at 50% load. At the pre-sent work, charge heating (hot EGR) and heat capacity effects ofEGR were considered at only 50% load.
2. Initial and boundary conditions
Calculations are carried out on the closed system from IVC(intake valve closing) at 165�CA BTDC (Before Top Dead Center) toEVO (exhaust valve opening) at 180�CA ATDC. Fig. 1-a shows thenumerical grid, which is planned to model the geometry of com-bustion chamber and contains a maximum of 42,200 cells at BTDC.As can be seen from the Fig. 1-b, grid dependency is based on thein-cylinder pressure and present resolution is found to giveadequately grid independent results. Initial pressure in the com-bustion chamber according to experimental data is set to 86 kPaand initial temperature is calculated to be 384 K. Boundary tem-peratures in the combustion chamber are as following:
Head temperature: 550 K.Piston temperature: 590 K.
Fig. 1. a Mesh of the Lister 8.1 indirect injection diesel engine and b Grid dependency based on the in-cylinder pressure.
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Cylinder temperature: 450 K.The present work is performed for four EGR mass fractions (0%
as baseline, 10%, 20% and 30%) at 50% load and engine speed of730 rpm. All the boundary temperatures are assumed to be con-stant throughout the simulation, but allowed to vary with thetemperature of the combustion chamber’s surface regions.
3. Energy analysis
The numerical model was established for a Lister 8.1 indirectinjection diesel engine with the specifications listed in Table 1.
The governing equations for unsteady, compressible,turbulently-reacting multi-component gas mixture flows andthermal fields were solved numerically from IVC to EVO. The tur-bulent flow within the combustion chamber was simulated usingthe RNG (Reynolds Normalized Group) k�ε turbulence model [21].Also, for the primary and secondary atomization modeling of theresulting droplets, the standard WAVE model was employed. TheDukowicz model, described in Ref. [22], was applied to for themodeling of heating up and evaporation of the droplets. A Sto-chastic dispersion model was employed to take the effect ofinteraction between the particles and the turbulent eddies intoaccount by adding a fluctuating velocity to the mean gas velocity[23]. This model assumes that the fluctuating velocity has arandomly Gaussian distribution. The Shell auto-ignition model wasused for modeling the auto ignition [24]. The Eddy EBU EBU (break-
Table 1Specifications of Lister 8.1 IDI diesel engine.
Cycle type Four strokeNumber of cylinders 1Injection type IDICylinder bore*stroke 0.1141*0.1397 (m̂2)L/R 4Displacement volume 1.43e�3 (m̂3)Compression ratio 17.5:1Vpre-chamber/VTDC 0.750% load injected mass 3.2009e�5 kg per cycleInjection pressure 88.8 (bar)Start injection timing 20� BTDCNuzzle diameter at hole center 0.003 mNumber of nuzzle holes 1Nuzzle outer diameter 0.0003 mSpray cone angle 10�
Valve timing IVO ¼ 5� BTDCIVC ¼ 15� ABDCEVO ¼ 55� BBDCEVC ¼ 15� ATDC
up model) based on the turbulent mixing was used for modeling ofthe combustion in the combustion chamber [25] as follows:
r_rfu ¼ CfusR
rmin�yfu;
yoxS;Cpr:ypr1þ S
�(1)
where this model assumes that in premixed turbulent flames, thereactants are contained in the same eddies and are separated fromeddies containing hot combustion products. The rate of dissipationof these eddies determines the rate of combustion. In other words,chemical reaction occurs fast and the combustion is mixingcontrolled. The first two terms of the “minimum value of” operatordetermine whether fuel or oxygen is present in limiting quantity,and the third term is a reaction probability which ensures that theflame is not spread in the absence of hot products. In above equa-tion, Cfu and Cpr are constant coefficients and sR is the turbulentmixing time scale for reaction. The value of Cfu varies from 3 to 25 indiesel engines and optimum value selected according to experi-mental data [26]. At the present work, the numerically solving ofspecies transport module for gas phase chemical species (O2, H2O,N2, fuel, CO2) provides more accurate mass fractions in thecomputational domain. Also, the numerically solving of transportmodules for momentum and energy give temperature and veloc-ities in the computational domain. Above transport equations withcombustion, spray and appropriate turbulence model modules aresolved simultaneously in each cell of the computational domain byfinite volume method. Final values for species mass fractions,pressure, and temperature are averaged over all cells. Therefore,because of the high space resolution of 3-D modeling, the moreaccurate results are achieved in comparison to the other thermo-dynamics models. Exergy analyses are carried out by using thecalculated values of species mass fractions, pressure, temperatureand burning rate. For this reason, exergy analyses by 3-D modelsresults higher accurately than the thermodynamics models.
4. Exergy analysis
The exergy of a system is defined as themaximumvalue of workthat can be obtained from that systemwhen it reaches mechanical,thermal and chemical equilibriums with its environment. This stateof equilibrium is defined as the dead state of the system and it isdependent on the temperature, pressure and composition of theenvironment. When there is no heat exchange between the systemand environment, thermal equilibrium is achieved. Similarly, me-chanical equilibrium is achieved when no work exchange occursbetween the system and its environment.
Fig. 2. Comparison of measured [26] and calculated pressure for baseline at 50% loadand engine speed 730 rpm.
S. Jafarmadar / Energy 66 (2014) 305e313308
Also, chemical equilibrium is achieved when no systemcomponent can react with the component in environment. For thepresent case, this means that in the dead state, all the species of theworking medium were either reduced or oxidized to N2, O2, CO2,and H2O. According to [27], the total exergy of a system is equal to:
Ex ¼ Exch þ Extm ¼ E � P0V � T0S�Xkki¼1
m0i mi (2)
where is the chemical potential of species i at the true dead state,and mi is the mass of species i.
The exergy balance equation for the inside chamber of the IDIengine, on crank angle basis, is expressed as follows [27]:
dExdq
¼ dExwdq
� dExqdq
� dIdq
þ dExfdq
(3)
Exw is the exergy associated with work done by the system inthe chamber and it is defined as:
dExwdq
¼ ðP � P0ÞdVdq
(4)
Also, Exq is expressed as the exergy associated with heat lossacross the boundary of chamber. Its variation with crank angle isdefined as:
dExqdq
¼�1� T0
T
�dQdq
(5)
I term is the destruction exergy associated with the combustionprocess in the chamber and it can be defined as:
dIdq
¼ T0T
Xkki¼1
midmi
dq(6)
where index i includes all the reactants and products. For perfectgases, mi ¼ gi is valid.
The exergy of liquid fuels (CzHy), which are used in compressionignition engines, is approximated by Ref. [27]:
dExfdq
¼ afdmfdq
(7)
where af is denoted by:
af ¼ LHV�1:04224� 0:011925
yzþ 0:042
z
�(8)
The exergy efficiency can be defined as the ratio of indicatedwork over total input chemical exergy. For the closed part of thecycle in an engine, the exergy efficiency is as follows:
hII ¼ Wnet
Exfuel(9)
Fig. 3. Variation of in cylinder pressure with crank angle position at various EGR massfractions in 50% load and engine speed 730 rpm.
5. Results and discussions
The calculations of exergy terms are carried out for a single-cylinder chamber of Lister 8.1 IDI diesel engine operating at 50%load, while applying 0%, 10%, 20% and 30% EGR mass fractions, at aconstant speed of 730 rpm. Fig. 2 shows the computed andmeasured [26] mean in-cylinder pressures for the baseline engine.
The presented results in these figures are global quantities(cylinder-averaged) and were shown as a function of time (crankangle). The parameters of fuel injection timing and amount of
injected mass were adjusted according to experimental data. Fig. 2demonstrates that computational and experimental pressure his-tories obtained during the compression, combustion and expansionstrokes are in excellent agreement. The discrepancy between peakpressures of experiment and computation is less than 0.2%. Thepeak pressure and ignition delay are 42.3 bar and 10.7 CAD. Thediscrepancy between the SOC (start of combustion) or ID (ignitiondelay) values of computation and experiment is 0.1 CAD. Fig. 3
Fig. 5. The variation of accumulative heat release rate with crank angle position atvarious EGR mass fractions in 50% load and engine speed 730 rpm.
S. Jafarmadar / Energy 66 (2014) 305e313 309
shows the variation of pressure with crank angle at various EGRmass fractions. The peak pressure values in the chamber are 42.4,40.01, 38.64 and 36.97 bar at 0%, 10%, 20% and 30% EGR massfractions, respectively. The main reason for lower peak pressures athigher EGR mass fractions is mainly due to ignition delay andcombustion rate reduction through premixed combustion phase.Also, the lower combustion rates at higher EGR mass fractionsstems from lower oxygen availability, which plays a significant rolehere.
The mean in-cylinder temperature histories for various EGRmass fractions are presented in Fig. 4. It is apparent from Fig. 4 thatwith the increase of EGRmass fraction, the in-cylinder temperaturealso increases during the compression stroke. The peak tempera-ture values at 0%, 10%, 20% and 30% EGR mass fractions are 1391,1409, 1411 and 1406 K. The initial temperature increases at EGRmass fractions induce the peak values of in-cylinder temperature toincrease slightly. As the EGRmass fraction increases, the in-cylindertemperature increases considerably during the power stroke due tothe longer combustion duration and the retardation of the com-bustion process.
Fig. 5 shows the changes of cumulative heat release rate withcrank angle position at various EGR mass fractions. The cumulativeterms are obtained after the integration of the corresponding ratesover the crank angle positions. It is clear from this figure that as theEGR mass fraction increases, the cumulative energy and its finalvalue decrease because of incomplete combustion. The accumula-tive energy values at 0%, 10%, 20% and 30% EGR mass fractions are1355.97, 1343.32, 1322.21 and 1287.81 J. At the 30% EGR massfraction, the final value of cumulative heat release rate decreases by5% relative to that of the 0% EGR. Higher EGR mass fraction canresult in more incomplete combustion and thus in a significantincrease of UHC (Unburned Hydrocarbon) emissions. Therefore,accumulative heat release rate decreases with an increment of EGRmass fraction. Also as shown in this figure, the amount of ignitiondelay period is the same for all cases. At high EGR, the negativeeffect compensate by positive effect due to increasing of initialtemperature.
Figs. 6 and 7 illustrate the trends of work exergy rate and cu-mulative work exergy in the chamber at various EGR mass
Fig. 4. The variation of temperature in cylinder with crank angle position at variousEGR mass fractions in 50% load and engine speed 730 rpm.
fractions, during the engine’s closed cycle. At crank positions afterthe top dead center, the work exergy rate decreases with the in-crease of EGR mass fraction percentage. The higher rates of workexergy can be observed at lower EGR mass fractions because offaster combustion and the higher cylinder pressure gradients. Themaximum rates of work exergy at 0%, 10%, 20% and 30% EGR massfractions are 16.82, 15.82, 14.82 and 13.85 J/deg, respectively. It canbe seen that when the EGR mass fraction increased from 0% to 30%,the maximum rate of work exergy decreased by 17.7%. The cumu-lative work exergies for 0%, 10%, 20% and 30% EGR mass fractionsare 459.89, 433.27, 403.23 and 367.78 J. It is evident that when theEGR mass fraction increases from 0% to 30%, the cumulative workexergy decreases by 20%. Moreover, the cumulative work exergydecreased from 31.7% of injected fuel exergy to 25.4%. The value forwithout EGR case is confirmed by the data in the literature [7,28e
Fig. 6. The variation of the rate of work exergy in chamber with crank angle position atvarious EGR mass fractions for 50% load at 730 rpm.
Fig. 7. The variation of the accumulative work exergy in chamber with crank angleposition at various EGR mass fractions for 50% load at 730 rpm.
Fig. 9. The variation of the accumulative heat loss exergy in chamber with crank angleposition at various EGR mass fractions for 50% load at 730 rpm.
S. Jafarmadar / Energy 66 (2014) 305e313310
32] with considering exergy efficiency in the diesel engines. Themore exergy efficiency decrease at the higher EGR mass fraction isdue to lower accumulative heat release rate and incompletecombustion.
Figs. 8 and 9 show the trends of heat loss exergy rate and cu-mulative heat loss exergy in the chamber at various EGR massfractions during the engine’s closed cycle. At crank angles beyondthe start of combustion, the rates of heat loss exergy for EGR casesare lower than that of the baseline engine, because during thebaseline operation, the combustion process was improved and ahigher pressure existed in the chamber. The peak rates of heat lossexergy are 7, 6.86, 6.5 and 6.07 J/deg for EGRmass fractions from 0%to 30%. It is clear that when the EGRmass fraction increases from 0%
Fig. 8. The variation of the rate of heat loss exergy in chamber with crank angle po-sition at various EGR mass fractions for 50% load at 730 rpm.
to 30%, the heat loss exergy rate decreases by 11.5%. However, thiseffect can be completely compensated by higher temperatures inthe exhaust stroke and longer combustion durations at EGR cases.As is shown in Fig. 9, the amounts of cumulative heat loss exergyare the same for all cases, constituting 24% of the injected fuelexergy. The value for baseline engine was confirmed by the work ofPrimus RJ, Flynn PF [31] considering heat loss exergy in the dieselengines. Also, the results for EGR cases were validated by the workof Ghazikhani et al. [32].
Figs. 10e12 show the changes of thermo-mechanical, chemicaland total exergies in the cylinder with crank angle positions atdifferent EGR mass fractions. In the compression stroke, and beforethe fuel injection time, the thermo-mechanical exergy in thechamber increases due to (a): the work produced by the piston and(b): increase of the initial temperature associated with EGR appli-cation. Also, the chemical exergy increases because of the EGRmassfractions addition and the complete combustion products in the
Fig. 10. The variation of thermo mechanical exergy with crank angle position atvarious EGR mass fractions for 50% load and 730 rpm.
Fig. 11. The variation of chemical exergy with crank angle position at various EGR massfractions for 50% load and 730 rpm.
Fig. 13. The variation of rate of irreversibility with crank angle position at various EGRmass fractions for 50% load and 730 rpm.
S. Jafarmadar / Energy 66 (2014) 305e313 311
cylinder. As shown in Fig. 11, with the increase of the EGR massfraction from 0% to 30%, the chemical exergy increases considerablyduring the compression stroke. The thermo-mechanical andchemical exergies in the chamber increased with the start of thecombustion process due to the rise of temperature and pressureand the concentration of complete combustion products. At the endof the combustion process, the amount of chemical exergyremained constant because of the constant concentration of com-bustion products, while the thermo-mechanical exergy diminisheddue to the decrease of gas temperature during the expansionstroke. As Fig. 10 illustrates, when the EGR mass fraction increasesfrom 0% to 30%, the exhaust thermo-mechanical exergy increasesby 33.22%. This occurs from higher pressure and temperature at the
Fig. 12. The variation of total exergy with crank angle position at various EGR massfractions for 50% load and 730 rpm.
exhaust valve opening time in the EGR case (as shown in Fig. 4). Ateach EGR mass fraction, chemical exergy increases during thecombustion due to the formation of complete combustion prod-ucts. Chemical exergy at exhaust valve opening increases from39.35 J to 71.38 J, when the EGR mass fraction goes up from 0% to30%. Fig. 12 indicates that the total exhaust loss exergy for 0%, 10%,20% and 30% EGR mass fractions are 262.9, 301.6, 338 and 369.6 J,which correspond to 18.1%, 20.8%, 23.3% and 25.51% of total fuelexergy, respectively. These values show a good agreement with thecorresponding data in Refs. [9,10,31,32]. The increasing of exhaustloss exergy in higher EGR rate is due to higher exhaust temperature.
The combustion process makes the highest contribution to thetotal in-cylinder irreversibility in a diesel engine, which according
Fig. 14. The variation of burn fuel exergy with crank angle position at various EGRmass fractions for 50% load and 730 rpm.
Fig. 15. The variation of accumulative irreversibility with crank angle position atvarious EGR mass fractions for 50% load and 730 rpm.
S. Jafarmadar / Energy 66 (2014) 305e313312
to the research by Primus and Flynn [31], is more than 90%. Figs. 13and 14 illustrate the trends for the rate and cumulative irrevers-ibilities due to in-cylinder combustion during an engine’s closedcycle at various EGR mass fractions. As shown in Figs. 5 and 15, thesame behavior can be seen for the cumulative heat release rate andcumulative irreversibility due to combustion. This similarity is alsoobserved in Figs. 13 and 14 between the rates of these parameters.For lower EGR mass fraction operations, with more oxygen avail-ability, which results in very rapid burning rates, combustion startswith higher rates of pressure rise and higher peaks of pressure.Therefore, because of this higher burning rate in the premixedphase, the rate of irreversibility increases more at lower EGR massfractions than at higher ones and then it decreases at the diffusionphase because of a lower combustion rate. The peak value of therate of irreversibility drops from 19.44 J/deg to 13.0 J/deg, whenEGR mass fraction increases from 0% to 30%.
The variation of cumulative irreversibility with crank angleposition is shown in Fig. 15 for different EGR mass fractions. It isclear from this figure that when EGR mass fraction increases, therate of cumulative irreversibility decreases due to the deteriorationof combustion. When EGR mass fraction increases from 0% to 30%,cumulative irreversibility diminishes by 14.3%. This result isconfirmed by the experimental work of Ghazikhani et al. [11] wherethey showed that at lower engine load and speed, there is aconsiderable decrease in total in-cylinder irreversibility whenhigher EGR is employed to an IDI diesel engine. Also, this trend wasconfirmed in the result of the reference [27]. They showed that anincreasing combustion temperature decreases the percentage ofcombustion irreversibilities. According to this study, an increasingcylinder charge gas temperature decreases the relative amount ofheat transfer from the reacting gas to the yet unburnedmixture. Forthe baseline engine, irreversibility constitutes 29.76% of injectedfuel exergy at 50% load operation. This result is confirmed by thework of Primus and Flynn [31] who discovered that the in-cylinderirreversibility values are in the range of 20e25% for full-load op-erations and that higher values are expected for lower-loadoperations.
6. Conclusions
In the present work, a three-dimensional CFD code has beenused to study the combustion processes within the chamber of a
Lister 8.1 IDI diesel engine, from the perspective of the second lawof thermodynamics, for various EGR mass fractions and 50% loadoperation. The calculated pressure results for the baseline enginewere compared with the corresponding experimental data, andshowed good agreement. Such correlations between the experi-mental and computed results make the model reliable for theprediction of exergy terms at various EGR mass fractions. Variousexergy terms including the fuel, heat loss, irreversibility, work,exhaust loss, chemical and thermo-mechanical exergies are pre-sented for different EGR mass fractions. Based on this study, thefollowing conclusions are made, as the EGR mass fraction increasesfrom 0% to 30% by 10% increments:
1 The in-cylinder peak pressure decreases by 1.5%.2Because of the more incomplete combustion in higher EGRmass fraction, the cumulative heat release rate decreases by 5%.
3 Because of the more incomplete combustion in higher EGRmassfraction, the burned fuel exergy decreases by 5%.
4 Because of the more decreasing of pressure in cylinder forhigher EGRmass fraction, the cumulativework exergy decreasesby 20%.
5 The cumulative heat loss exergy increases by 1.1%.6 Because of themore increasing of exhaust temperature in higher
EGRmass fraction, the exhaust losses exergy increases by 40.6%.7 Because of themore increasing of initial charge temperature and
peak temperature in cylinder for higher EGR mass fraction, thecumulative irreversibility decreases by 14.2%.
8 Because of the more decreasing of work exergy for higher EGRmass fraction, the exergy efficiency decreases by 20%.
The total irreversibility for the baseline engine operating at 50%load is 29.8% of the injected fuel exergy. This result correlates wellwith the corresponding data in Ref. [27]. The present study dem-onstrates that CFD simulations can be used to significantly improveand facilitate the understanding and analysis of combustion, fromthe second-law viewpoint, in multi-chamber diesel engines oper-ating under different EGR scenarios.
It should be acknowledged that this study considers only 50%load. Additional studies are needed to address the effects of EGRmass fraction at various loads on the exergy terms and exergy ef-ficiency. More differences in exergy terms should be observed athigh load because of higher temperature in cylinder. At this con-dition, EGR using is a very effective method for the reduction ofpeak temperature in cylinder.
NomenclatureE internal energy (J)G Gibbs function (J)Ex exergy (J)S entropy (J/K)T temperature (K)kk number of speciesI irreversibility (J/K)x mole fraction of speciesy mass fraction of species
Greek lettersm chemical potential (J/kg)q crank angle (degree)z number of carbon atom
AbbreviationsBTDC before top dead center (degree)ATDC after top dead center (degree)EVO exhaust valve opening (degree)
S. Jafarmadar / Energy 66 (2014) 305e313 313
IDI indirect injection engineCA crank angle (degree)SOC start of combustion (degree)EBU eddy break upID ignition delay (degree)N engine revolution (rpm)
Subscriptch relating to chemical exergytm relating to thermo-mechanical exergyf relating to fuelw associated with work transferQ associated with heat transfer0 dead state, or environment statepr relating to combustion productsox relating to oxidantsfuel relating to fuel
Superscript0 restricted dead state
References
[1] Gibbs JW. The scientific papers, vol. 1. New York, USA: Dover; 1961.[2] Haywood RW. Equilibrium thermodynamics for engineers and scientists. New
York, USA: John Wiley and Sons; 1980.[3] Dunbar WR, Lior N, Gaggioli RA. The component equations of energy and
exergy. Trans ASME J Energy Res Technol 1992;114:75e83.[4] Dunbar WR, Lior N. Sources of combustion irreversibility. Combust Sci Technol
1994;103:41e61.[5] Rakopoulos CD, Andritsakis EC. DI and IDI diesel engines combustion irre-
versibility analysis. In: ASME-WA Meeting, New Orleans LA, Proc. AES, vol. 30;1993. pp. 17e32.
[6] Rakopoulos CD, Giakoumis EG. Simulation and exergy analysis of transientdiesel engine operation. Energy 1997;22:875e85.
[7] Caton JA. On the destruction of availability (exergy) due to combustionprocessesdwith specific application to internal combustion engines. Energy2000;25:1097e117.
[8] Kyritsis DC, Rakopoulos CD. Parametric study of the availability balance in aninternal combustion engine cylinder; 2001. SAE paper 2001-01-1263.
[9] Rakopoulos CD, Kyritsis DC. The influence of cylinder wall temperature profileon the second-law diesel engine transient response. Appl Therm Eng 2005;25:1779e95.
[10] Rakopoulos CD, Giakoumis EG. Development of cumulative and availabilityrate balances in a multi-cylinder turbocharged IDI diesel engine. EnergyConvers Manage 1997;38:347e69.
[11] Ghazikhani M, Feyz ME, Joharchi A. Experimental investigation of the exhaustgas recirculation effects on irreversibility and brake specific fuel consumptionof indirect injection diesel engines. Appl Therm Eng 2010;30:1711e8.
[12] Rakopoulos CD, Giakoumis EG. Availability analysis of a turbocharged dieselengine operating under transient load conditions. Energy 2004;29:1085e104.
[13] Amjad AK, Khoshbakhi Saray R, Mahmoudi SMS, Rahimi A. Availability anal-ysis of n-heptane and natural gas blends combustion in HCCI engines. Energy2011;36:6900e9.
[14] Hosseinzadeh A, Khoshbakhti Saray R, Seyed Mahmoudi SMS. Comparison ofthermal, radical and chemical effects of EGR gases using availability analysisin dual-fuel engines at 50% loads. Energy Convers Manage 2010;51:2321e9.
[15] Turan Onder. Exergetic effects of some design parameters on the smallturbojet engine for unmanned air vehicle application. Energy 2012;46:51e61.
[16] Jafarmadar S. Three-dimensional modeling and exergy analysis in combustionchambers of an indirect injection diesel engine. Fuel 2013;107:439e47.
[17] Jafarmadar S, Tasoujiazar R, Jalelpor B. Exergy analysis in a low heat rejectionIDI diesel engine by three dimensional modeling. Int J Energ Res; 2013:1e13.http://dx.doi.org/10.1002/er.3100.
[18] Jafarmadar S, Zehni A. Numerical investigation of the effects of dwell timeduration in a two-stage injection scheme on exergy terms in an IDI dieselengine by three-dimensional modeling. Energ Sci Eng; 2013:1e14. http://dx.doi.org/10.1002/ese3.25.
[19] Balli O, Hepbasli A. Exergoeconomic, sustainability and environmental dam-age cost analyses of T56 turboprop engine. Energy 2014;64:582e600.
[20] Motasemi F, Afzal Muhammad T, Salema Arshad A, Moghavvemi M,Shekarchian M, Zarifi F. Energy and exergy utilization efficiencies and emis-sion performance of Canadian transportation sector, 1990e2035. Energy2014;64:355e66.
[21] Han Z, Reitz RD. Turbulence modeling of internal combustion engines usingRNG K-ε models. Combust Sci Technol 1995;106:267e95.
[22] Liu AB, Mather D, Reitz RD. Modelling the effects of drop drag and breakup onfuel sprays. SAE Technical Paper 930072; 1993.
[23] Dukowicz JK. Quasi-steady droplet change in the presence of convection.Informal report Los Alamos Scientific Laboratory, LA7997-MS.
[24] Halstead M, Kirsch L, Quinn C. The auto ignition of hydrocarbon fueled at hightemperatures and pressures - fitting of a mathematical model. Combust Flame1977;30:45e60.
[25] Magnussen BF, Hjertager BH. On mathematical modeling of turbulent com-bustion with special emphasis on soot formation and combustion. In: 16 thSymp. (Int.) on Combustion (1976). Pittsburg, Pennsylvania: Comb. Inst.;1976. pp. 719e29.
[26] Jafarmadar S, Zehni A. Multi-dimensional modeling of split injection scheme onperformance and emissions of IDI diesel engines. Int J Eng 2012;25:135e45.
[27] Rakopoulos CD, Giakoumis EG. Diesel engine transient operation. London, UK:Springer-Verlag; 2009.
[28] Van Gerpen JH, Shapiro HN. Second-law analysis of diesel engine combustion.Trans ASME J Eng Gas Turb Power 1990;112:129e37.
[29] Rakopoulos CD, Giakoumis EG. Second-law analyses applied to internalcombustion engines operation. Prog Energy Combust Sci 2006;32:2e47.
[30] Rodriguez L. Calculation of available-energy quantities. In: Gaggioli RA, editor.Thermodynamics: second law analysis. Washington DC: American ChemicalSociety Symposium Series, No. 122; 1980. pp. 39e59.
[31] Primus RJ, Flynn PF. The assessment of losses in diesel engines using secondlaw analysis. In: ASME WA-Meeting, Anaheim CA, Proc. AES; 1986. pp. 61e8.
[32] Ghazikhani M, Sharifzadeh M, Arjangnia A, Sharifzadeh A. Experimentalinvestigation of the EGR temperature effects on the destruction of the fuel’savailability due to combustion processes in IDI diesel engine cylinder. J EngineRes 2011;22:62e8.