An advanced combustion mmechanism for D.I diesel e

Amin Maghbouli, Wenming Yangal Univer

reduced at the engine mid loadcondition.

condition, a consid-ess where salue of the fload at all

speeds producing higher amounts of unburned hydro carbons and carbon monoxide. By reducingload and speed, output power was decreased as well. However, same trend was not reported for tcated thermal efciency as the middle engine speed in considered engine loads, had slightlyefciency.

2013 Elsevier Ltd. All rights reserved.

Corresponding author. Tel.: +65 65166481.

Applied Energy 111 (2013) 758770

Contents lists available at SciVerse ScienceDirect

journaE-mail address: mpeywm@nus.edu.sg (W. Yang).alence ratio distribution inside the combustion chamber showed that under full loaderable amount of fuel was trapped in piston bowl after initiation of the injection procrich local regions provide the potential for production of higher soot emission. Mean vcentration history showed that the ignition delay was increased under mid engine0306-2619/$ - see front matter 2013 Elsevier Ltd. All rights reserved.http://dx.doi.org/10.1016/j.apenergy.2013.05.031uch fueluel con-engineenginehe indi-higherDiesel engineDirect injectionKIVA-4Multi-component combustion model

develop a multi-component fuel combustion model. The developed combustion model was capable ofmodeling combustion process of number of chemical species as the components of direct injected liquidfuel. Comparing to the single component fuel combustion model, new model is capable of comprehensivecombustion modeling of blend fuel and heavy hydro-carbon fuels. Furthermore, spray breakup and col-lision models were changed to more advanced KelvinHelmholtz and RayleighTaylor (KHRT) andORourke models, respectively. The model was used to simulate direct injected diesel engine under fulland mid engine loads at three engine speed conditions. Extracted temporal and spatial results for equiv-a r t i c l e i n f o

Article history:Received 21 September 2012Received in revised form 6 April 2013Accepted 13 May 2013Available online 15 June 2013

Keywords:a b s t r a c t

A multi-dimensional computational uid dynamics (CFD) modeling was conducted on a direct injectionturbo-charged diesel engine based on KIVA-4 code under full and mid engine loads. Multi-componentfuel evaporation model of KIVA-4 was used and coupled with advanced combustion chemistry to gener-ate a multi-component fuel combustion model by integrating CHEMKIN II into the KIVA-4 code. As thecoding schema of KIVA-4 in the case of data/parameter allocation, etc. was different compared to previ-ous version of KIVA-3V, a considerable amount of FORTRAN programming was performed in order tonumerical simulations. CO2, CO and UHC emissions wereDepartment of Mechanical Engineering, Nation

h i g h l i g h t s

An advanced multi-component fuelcombustion model was developed.

Numerical simulations wereconducted for a diesel engine at fulland mid loads/speeds.

Temporal/spatial variation of theemissions were extracted fromodel coupled with detailed chemical reactionngine simulation, Hui An, Jing Li, Siaw Kiang Chou, Kian Jon Chuasity of Singapore, 9 Engineering Drive 1, Singapore 117576, Singapore

g r a p h i c a l a b s t r a c t

l homepage: www.elsevier .com/ locate/apenergyApplied Energy

d En1. Introduction

Diesel engines have proved their reliability as a beating heartfor the mass transportation and power generation units. The mostimportant factor that has caused diesel engines to surpass gasoline

Nomenclature

A0 uid ow constantD diffusion coefcientJ heat uxM number of reactions in chemical kinetics mechanismP pressureNc number of cylinders_Qc source term due to chemistry_Qs source term owing to sprayt timeu velocity vectorW molecular weight

SubscriptsM number of species in chemical kinetics mechanismn number of the fuel components

Greek symbolsdml dirac delta functionDhf heat of formation at absolute zeroq density_qcm density change rate due to chemistry_qsm density change rate owing to Spraye dissipation rate of turbulence_xm molar production rate

A. Maghbouli et al. / Applieengines, is their lower fuel consumption associated with high spe-cic output power and torque. Despite the fact that diesel enginesfulll the daily increasing power demand and seem to be a prom-ising tool to achieve efcient fuel consumption, conventional dieselcycle suffers from high levels of NOx and Soot emissions at high en-gine loads, and low thermal efciency and high levels of UnburnedHydro Carbon (UHC) emissions at lower engine load conditions[1,2]. This is why there has been a signicant amount of researchesworldwide to optimize and introduce new technologies in order tominimize emissions from diesel combustion while providingacceptable power output. These modern technologies have mostlyconcentrated on direct injecting of the fuel into the combustionchamber with precise timing using common-rail fuel injectorsand the use of Exhaust Gas Recirculation (EGR) to reduce in-cylin-der peak temperature, avoiding massive production of NOx emis-sions [36].

As an alternative approach to reduce emission levels and fossilfuel usage, application of several environmentally friendly fuelssuch as: ether, alcohol, vegetable/fat based bio-diesel or gaseousfuel in diesel cycle was investigated by many researchers. Yinget al. studied behavior of the diesel and di-methyl ether (DME)blends on a diesel engine at low and high engine loads [7]. Theyconcluded that at the high load conditions, smoke emissions werereduced about 5868% for the blend fuel in comparison to the purediesel. At low load conditions, smoke emissions for the blend fuelwere almost comparable to the neat diesel. In addition, by increas-ing DME, NOx emission was gradually decreased, while CO andUHC emissions were increased at the most operating conditions.Papagiannakis and Hountalas conducted experiments on a dieselengine running under dual-fuel mode for various engine loads,where liquid diesel was partially replaced with natural gas [8].The results showed that using gaseous fuel reduced the peak cyl-inder pressure. In the case of pollutant emissions, they concludedthat addition of the gaseous fuel had a positive effect on reduc-ing NO emission. However, at lower engine load, the CO andUHC emissions were higher compared to the normal diesel

AbbreviationsATDC After Top Dead CenterCAD Crank Angle DegreeCFD computational uid dynamicsCO Carbon MonoxideCO2 Carbon DioxideD.I Direct InjectionDVODE double-precision variable-coefcient ordinary differen-

tial equation solverEGR Exhaust Gas RecirculationEVO Exhaust Valve OpeningHCCI Homogeneous Charge Compression IgnitionHRR Heat Release RateIC Internal CombustionIVC Inlet Valve ClosureITE indicated thermal efciencyNOx Nitrogen OxidesSOI Start of InjectionTDC Top Dead CenterUHC Unburned Hydro Carbon

ergy 111 (2013) 758770 759operation.Although experiments are essential tools to study, understand

and optimize the typical IC engines, the range of their applicabilityin terms of economy, time and available infrastructures is the mat-ter of debate. Numerical simulations of the IC engines through de-tailed mathematical framework have been shown to be a veryuseful tool for saving experimental resources. Nowadays, withthe advent of powerful computers with high computational capac-ities, numerical modeling has proved its exibility and low cost.Generally, engine modeling mathematical framework can be clas-sied into two main groups: thermodynamic and dimensionalmodels. Thermodynamic models assume energy equation as themain conservation equation and are subdivided into single andmulti-zone models [911]. Since physical processes such as turbu-lent ow effects are not directly modeled, some correlations withnumerous constants are used instead. Kouremenos et al. used amulti-zone thermodynamic model to study a D.I diesel engine atdifferent engine loads and speeds [12]. Results showed that fuelconsumption was increased by increasing the engine load for alloperating engine speeds. To simulate performance and emissionsof a D.I diesel engine, Rakopoulos and Hountalas developed a 3Dmulti-zone thermodynamic model [13]. In their multi-zone model,engine uid ow was not taken into account directly in simula-tions; however, they achieved to a good agreement between simu-lations and experimental measurements for in-cylinder pressure,heat release rate and emission data. Hountalas et al. developed amulti-zone thermodynamic model to investigate characteristicsof a diesel engine under full and partial loads [14]. They concludedthat higher injection pressure signicantly increased NO emissionespecially at low speed and partial load, whereas positive effectwas resulted on soot reduction.

By contrast, in dimensional models temporal and spatial varia-tions of the ow eld, pressure, composition, temperature andturbulence inside the combustion chamber are taken into account[15]. Various sub-models are applied in dimensional models tosimulate different aspects of in-cylinder processes. Besson et al.optimized HCCI dedicated diesel engine combustion chambergeometry under engine full load condition using KIVA II-Renaultcode [16]. They concluded that widest bowl geometry at full loadprovides the highest gross mean effective pressure and lowestsoot emission. Ehleskog et al. used KIVA-3 code and diesel surro-gate fuel chemical kinetics together with turbulence-chemistryinteractions in multi-dimensional simulation of D.I diesel engineat a single engine speed and just for 25% load condition [17]. Theystudied effects of split injection and concluded that split injectionusing a common rail piezo-electric common rail injector could beused to reduce NOx emission and specic fuel consumption with-out increasing soot emissions at lower loads. However, CO andUHC emissions tended to increase in all cases including split injec-tion. Helmantel and Golovitchev used 3D-CFD with chemical

emission characteristics of a typical diesel engine for variety of en-gine loads and speeds. In this paper, the conducted experimentswill be briey discussed rst and then the developed mathematicalsimulation framework will be presented. After validation of thenumerical results by the experiments variety of engine combustionand performance parameters will be analyzed.

Table 1Toyota 2KD-FTV engine specications.

Engine aspiration Turbo-chargedEngine cycle Diesel 4 StrokeNumber of cylinders 4 InlineInjector orices 6 CentricFuel injection system Common rail, densoBore stroke 92 93.8 (mm)Displacement volume 2.494 (lit)Compression ratio 18.5:1Max. output power 75 (kW) @ 3600 (rpm)Max. output torque 260 (N m) @ 1600 (rpm)

760 A. Maghbouli et al. / Applied Energy 111 (2013) 758770kinetics to simulate triple injection strategy to reduce soot andNOx emission levels in a single cylinder D.I. diesel engine at themedium load [18]. As they inferred, although split injection to-gether with high levels of EGR could reduce NOx emission, sootlevels would increase due to reduction in combustion chambertemperature. To overcome this, triple injection was introducedto oxidize the produced soot particles. Suggested amount of massand timing for the third injection or so called post injection was1730% of total fuel mass at 1012 crank angle ATDC, respec-tively. Shuai et al. used KIVA-3V coupled with chemical kineticsto simulate injection rate-shape effect on engine performanceand emission under low load operating condition [19]. They con-cluded that the early injection timing gave lower soot, UHC andCO emissions but higher NOx emissions compared to the lateinjection timing.

By the reference to extensive research on 3D-CFD modeling ofthe diesel engines, rare works have used KIVA-4 incorporated withdetailed chemical reaction model to investigate the combustionand emission formation process in the terms of temporal and spa-tial variation. Especially, if the purpose of the diesel engine devel-opment and optimization is in the vehicle application, detailedstudying of its behavior in different engine loads and speedsthrough numerical simulations is essential. The most importantaim of this study is to generate a promising numerical workstationwhich would be able to accurately predict combustion andFig. 1. Schematic diagram of t2. Experimental procedure

2.1. Engine setup and specications

The experiments were performed on a four cylinder Toyota2KD-FTV common rail fuel injection diesel engine in the enginelaboratory of National University of Singapore. A schematic dia-gram of the engine test bed is shown in Fig. 1 and the detailed en-gine specications are listed in Table 1. The engine was loadedwith an AVL DP 160 water-cooled passive eddy current dynamom-eter which is able to provide a peak brake power of 160 kW and amaximum torque of 400 N-m with an accuracy of 0.3%. The airow rate was measured using an AVL Sensyow P air ow meterwith a resolution of 100 ms for sampling frequency. An AVL733S.18 fuel balance was used to measure the fuel consumptionrate with a sampling frequency of 500 ms and an accuracy of1%. The cylinder pressure was measured at a resolution of 1CAD by an AVL GH13P water-cooled pressure transducer whichwas mounted on the cylinder head, and it can sustain a peak pres-sure of 250 bars.

2.2. Experimental test cases

Experimental cases including three engine speeds of 1200, 2400and 3600 rpm under mid and full engine loads were considered tohe diesel engine test bed.

In doing so, oxidation chemistry of the target heavy hydrocarbon

d Enperform numerical simulations. Input parameters including initialand boundary conditions for these cases are illustrated in Table 2.These parameters will be used on simulation process of the casesdiscussed above.

3. Multi-dimensional modeling

3.1. 3D-CFD modeling tool

Accurate and more realistic simulation of the liquid fuel spray,particle collision, oscillation, atomization and evaporation encoun-tered by highly turbulent reactive uid ow in the diesel enginesrequires utilization of a multi-dimensional CFD code. In this study,uid ow simulation was carried out using the latest version of LosAlamos National Laboratory (LANL) CFD code, KIVA-4 [20]. KIVA isa computer code for the numerical calculation of the transient, twoand three dimensional chemically reactive ows with sprays. Ituses a time-marching, nite-volume scheme which solves the con-servation equations of mass, momentum, energy and accounts forturbulence using an Arbitrary Lagrangian Eulerian (ALE) method inthree solution phases. Comparing to the earlier versions of theKIVA code, detailed evaporation model together with ability formodeling unstructured mesh were added to the KIVA-4. Nonethe-less, some basic sub-models in the default KIVA-4 require furthermodications to maximize code predictability. In doing so, Kel-vinHelmholtz and RayleighTaylor (KH-RT) spray break up mod-els [21] were implemented in KIVA-4 calculations for sprayprimary and secondary breakup stages, respectively. Moreover,ORourke collision model [22] was used for fuel particle collisionand the RNG ke model was used as turbulence model.

3.2. Multi-component fuel combustion model

3.2.1. Insights to the applied chemical kinetics mechanismDeveloping comprehensive reduced chemical kinetics mecha-

nisms for different fuels, researchers needed to include one ortwo global reactions at the beginning of the chemical kineticsmechanism of the heavy hydrocarbons such as diesel and biodiesel

Table 2Selected experimental cases (IVC = 149 CAD ATDC, EVO = 150 CAD ATDC).

Engine speed (rpm) 1200 2400 3600

Engine Load 50% 100% 50% 100% 50% 100%Pressure at IVC (bar) 1.161 1.27 1.686 1.81 1.87 1.89Temperature at IVC (K) 373 374 380 383 397 398Wall temperature (K) 523 525 529 530 531 539

Initial charge Air + 5% EGR Air + 7% EGR Air + 10% EGRDiesel mass (gr/cycle) 0.026 0.035 0.022 0.037 0.025 0.044Air mass (gr/cycle) 0.844 0.7 0.957 0.955 0.967 0.957Total equivalence ratio 2.16 1.33 1.72 1.72 2.58 1.45SOI (CAD ATDC) 7 5 9 10 16 17

A. Maghbouli et al. / Appliefuels [2325]. This is because either chemical pathways of molec-ular break down of the considered fuel surrogate was not availableor number of intermediate chemical species and kinetics pathwaysof their production/destruction would be so large that the nalmechanism cannot be used in 3D CFD calculations. For instance:Bergman and Golovitchev [23] developed a Diesel Oil Surrogate(DOS) chemical kinetics mechanism for conventional diesel fueloxidation chemistry and have been applied it in IC engine simula-tions. They have used C14H28 as the diesel fuel surrogate but chem-ical reaction pathways and number of intermediate species forC14H28 oxidation chemistry would be so high that the mechanismcannot be used in the engine 3D-CFD application. To remedy com-putational time issue, they introduced a global single step reactionwhich describes dissociation of 1.5 mol of C14H28 into 2 mol ofn-heptane (C7H16) and 1 mol of toluene (C7H8):fuel, in this case C14H28, was dened by two lighter hydrocarbonswhich rstly can represent components of the major fuel and sec-ondly have applicable chemical kinetics mechanism in terms ofnumber of the intermediate species and reactions. Ehleskog et al.[17] have used DOSmechanism in simulation of direct injected die-sel engine using single fuel component, C14H28, and the global singlestep reaction in its oxidation as discussed in Eq. (1). Nonetheless,this approach still has some deciencies such as; it is not suitablefor the multi-component fuel evaporation model, it considers lim-ited number of fuel components and its difculty in determinationof chemical kinetics rate coefcients of the introduced global reac-tion which can signicantly affect ignition delay time and combus-tion [17]. In this study, detailed reactionmechanism of the DOS [23]including 72 chemical species and 325 elementary reactions wasincorporated into KIVA-4 to simulate multi-component diesel fuelignition and combustion. However, the global reaction, Eq. (1),was removed from the chemical kinetics mechanism and insteadmulti-component evaporation and combustion were applied insimulations. Fuel components were considered to consist of n-hep-tane and toluene with mole basis proportions of two and one as ali-phatic and aromatic hydrocarbon compounds of the diesel fuel,respectively. Initial mole fractions of liquid fuel components wereintroduced to the KIVA-4 code in evaporation and subsequentlycombustion processes. It means that in the developed model, themulti-component fuel evaporation and combustionmodels operatetogether in a way that the evaporation model performs phasechange calculations and provides the combustion model with theamount of evaporated fuel for each specic fuel component. As itmentioned, n-heptane (C7H16) and toluene (C7H8) were used asthe fuel components in this study. Whenever the liquid fuel is in-jected into the combustion chamber, these two components willrepresent the diesel fuel both in evaporation and combustion mod-els. Phase change calculations of these species will be done by themulti-component evaporation model and then the multi-compo-nent fuel combustionmodelwill perform the combustion chemistrycalculations. Considering stochastic liquid fuel distribution and ap-plied multi-component evaporation model, it is possible that in atypical computational cell n-heptane chemistry would be moredominant compared to the toluene due to differences in heat ofvaporization of the applied fuel components. Nonetheless, multi-component fuel combustionmodelwill take into account chemistrycalculations of the both fuel components in a computational cellusing the dened chemical kinetics mechanism. This is becausereaction pathways of both n-heptane and toluene are available inthe chemical kinetics mechanism; however, oxidation and subse-quently combustion strength of the each fuel components are dueto their mixing with oxidant and forming of the mixture in amma-bility limit. Regarding to explanations above, Fig. 2 schematicallyillustrates comparison between developed multi-component fuelcombustionmodel in this study and conventional single componentfuel combustion model [17].

Using the DOS chemical kinetics mechanism, soot formationmechanism is dened within the reaction mechanism by a seriesof elementary reaction steps leading from acetylene to the rst aro-matic ring,A1. These reactionstepsare followedbysuccessive stagesof the HACAmechanism:H-abstraction and C2H2-addition resultingin formation of a chain of aromatic rings.Moreover, extended Zeldo-vich mechanism was used in NOx emission modeling [23].

3.2.2. Mathematical framework and programming details of the multi-1:5C14H28 0:5O2 ) 2C7H16 C7H8 H2O 1

ergy 111 (2013) 758770 761component fuel combustion modelFluid ow and fuel particle simulation procedure in the KIVA

code, the ALE method, decouples calculations of the diffusion and

convection terms from chemical source terms. Hence, each compu-tational cell can be treated as a homogeneously mixed reactor ateach time-step. Continuity equation for species m and energyequation in terms of specic internal energy are formulated inthe KIVA code as given in Eqs. (2) and (3), respectively [20].

@qm@t

r qmu r qDrqmq

_qcm _qsdml 2

@

@tqI r quI Pr u 1 A0r : rur J

A0qe _Qc _Qs 3

where _qcm in Eq. (2) and _Qc in Eq. (3) are the parameters that need tobe calculated by the combustion model. Mathematical descriptionsof these terms are as follows:

_qcm Wm _xm 4

_Qc XMm1

_xmDhf m 5

Combustion calculation in the default KIVA-4 code is based on aglobal single step reaction and cannot directly be used to includechemical kinetics reaction mechanism. To calculate the molarproduction rate of the chemical species, _xm; the gas phase kinetics

Fig. 2. Schematic gure of applied multi-component fuel evaporation and combustion model in this study (right) compared with single component evaporation andcombustion model [17] (left).

762 A. Maghbouli et al. / Applied Energy 111 (2013) 758770Fig. 3. Summarized ow chart of coupled KIVA4-CHEMKIN-DVODE model developed in this study.

d EnA. Maghbouli et al. / Applielibrary of CHEMKIN II [26] was integrated into KIVA-4 code. A sum-marized ow chart of this integration is presented in Fig 3. In thisprocedure, the CHEMKIN libraries have identied introducedchemical kinetics mechanism before coupling with KIVA-4 calcula-tions. Sample simulation starts with initialization tasks by theKIVA code such as: reading input thermodynamic, species concen-tration, initial and boundary conditions. KIVA continues the calcu-lation processes of the uid ow and when it reaches to the

Fig. 4. Flow chart of multi-component fuel combustion m

Table 3Needed programming modications and parameter denition of the developed multi-comCHEMKIN model [28,29].

Needed modication/parameter role Coupled KIVA-3V-CHEMKIN

Increase number of active chemical species tomatch the number of species in appliedchemical kinetics mechanism

nsp was increased from default 12 to

Include enthalpy data of the added chemicalspecies

hk enthalpy data array was updatedKIVA: kiva.f

For the fuel specie(s) First specie (nsp = 1) should be the a

Number of other chemical species (major/intermediate)

nsp-1

Species placement order in chemical kineticsmechanism

Fuel specie at the beginning of the spof the other species is optional, howmatch enthalpy array in kiva.f

Data allocation schema entire the code comkiva.i and comfuel.i external lesglobal code parameters throughoutof the KIVA-3Vcodeergy 111 (2013) 758770 763combustion calculations, if the temperature value for typical com-putational cell would be higher than 600 K as threshold value,combustion calculations will be activated for that cell. Otherwise,combustion will be bypassed by the code. Referring to Fig. 3, whenthe combustion is active for a computational cell an interfacenumerical unit will perform chemistry solutions by iterative call-ing of the coupled CHEMKIN-DVODE code [27] to solve the stiffsystem of ordinary differential equations.

odel in a particular time step developed in this study.

ponent fuel combustion model in KIVA-4-CHEMKIN model compared to KIVA-3V-

Coupled KIVA4-CHEMKIN

M nsp was increased from default 12 to M

in main driver of hk enthalpy data array was updated in an external le:datahk

pplied fuel From rst specie to n (n = nspl) should be the applied fuelcomponentsnsp-nspl

ecies. Placementever it should

Fuel components (n = nspl) at the beginning of the species.Placement of the other species is optional, however it shouldmatch enthalpy array in datahk

were used to linkthe sub-routines

allocatedata.f and emodule.f sub-routines were used to linkglobal code parameters throughout the sub-routines of theKIVA4 code

Chemical kinetics mechanisms developed for IC simulationsare mostly in gas phase; hence, code was developed in such away that it would also be able to simulate premixed combustionsystems such as HCCI engines. Fig. 4 is the detailed ow chart ofthe numerical calculations which would be done by the interfaceunit as a multi-component fuel combustion model. If there is aninjection process in the system, rstly the spray injection, breakupand multi-component evaporation routines will be called. In atime step of a simulation, evaporation model will perform liquidfuel phase change calculations till the end of that particular timestep and then the interface unit will deliver the concentrations ofthe gaseous fuel components to the multi-component fuel com-bustion model. Integrating by CHEMKIN chemistry solver, thisprocess in KIVA-4 is not same as the earlier version of KIVA-3V.In order to this, most important modications and parameterswhich should be accurately introduced to the coupled code are gi-ven in Table 3. The parameter named nsp is the number of activechemical species in the KIVA code which is only 12 in the defaultKIVA and while integrating with the CHEMKIN code it should beincreased to M number of species participated in chemical kinet-ics mechanism [28,29]. Moreover, there is a need to introduce en-thalpy of the species into the versions of the KIVA code as it wasstated in Table 3. The developed multi-component fuel combus-tion model in this study has a major difference with the coupledKIVA-3V-CHEMKIN in the case of number of fuel components. InKIVA4, loop counters of the chemical specie throughout the codeconsider nspl as the number of fuel components and nsp as thenumber of total species. Hence, in the Eqs. (4) and (5) domainof m can be dened as below:

m 1; 2; . . . ; n; n 1; n 2; . . . ; M 6

The model is capable of using any number of fuel componentsas long as their reaction pathways would be available in thechemical kinetics mechanism. It should also be noted that pro-gramming schema of the KIVA-4 code in the case of data/param-eter allocation is different compared to its previous version ofKIVA-3V. KIVA-3V uses an external le to link global parametersthroughout the code, whereas KIVA-4 has two added sub-routinesfor this purpose, see Table 3. Especially, if a new parameter isneeded to be introduced in couple of subroutines, this newparameter should be included in the mentioned internal data allo-cation subroutines of allocatedata.f and emodule.f. Otherwise, codecannot identify an applied parameter in a typical subroutinewhich is also used anywhere else in the code. Useful program-ming information can be found in [30].

Once the combustion in a particular computational cell is acti-vated, all processes discussed above will be done and then theinterface unit will transfer the newly calculated molar produc-tion/destruction rates of the species including fuel componentsand other species to the KIVA code to update the combustionsource terms in Eqs. (2) and (3). It should be noted that automatictime step sub-cycling of the DVODE code [27] was also activated toensure accurate chemistry calculations when the stiffness of someof ordinary differential equations would be high.

3.3. Engine geometry and computational mesh

Bowl geometry of the Toyota 2KD-FTV diesel engine is in omegashape to increase mixing of the liquid fuel and air by initiatedcirculation. In this study, the common rail fuel injector had 6injector orices. To reduce the computational time, a 60 sectormesh was created based on the symmetric combustion chamber

764 A. Maghbouli et al. / Applied Energy 111 (2013) 758770Fig. 5. Toyota 2KD-FTV 60 (a) medium and (b) ne sector meshes (at TDC) used in 32400 rpm and 100% load condition.D-CFD simulations and calculated pressure results using mesh a and b for engine

geometry of the Toyota 2KD-FTV car engine as shown in Fig. 5. Themedium mesh consisted of 3540 cells at TDC with the piston cre-vice region included. Although grids size at wall boundaries seemsto be coarse, advantage of using polar mesh over comes this issueas very ne cells (much less than 1 mm) are in liquid fuel spraypath and in vicinity of injector nozzle. This ensures accurate break-up and evaporation predictions for liquid fuel droplets. Using thismesh, run times for a closed cycle simulation took about 35 h ona single processor at the High Performance Computing system(HPC) in National University of Singapore. To validate the meshindependence, a ner mesh with the same engine geometry wascreated as shown in Fig. 5. The cylinder pressure curves were com-pared at the condition of 2400 rpm and 100% load (see Fig. 5). Itcan be seen that with further renements on the mesh, no

signicant difference was observed on the predicted cylinder pres-sure. By considering the computational time, the medium meshwas used for all the simulations.

4. Results and discussion

For considered experimental cases relevant numerical resultsare extracted to get better insight about engine operation detailsat engine full and mid loads and in three engine speeds. In doingso, results are illustrating engine performance characteristics suchas work, accumulative heat release and indicated thermal ef-ciency (ITE) and engine emissions characteristics for most impor-tant emissions such as NOx, Soot, CO, CO2 and UHC.

Fig. 6. Comparison of cylinder pressure and HRR results of experimental and numerical multi and single component fuel combustion models.

A. Maghbouli et al. / Applied Energy 111 (2013) 758770 765Fig. 7. Numerical results of performance characteristics of 100% load: (a) accumulativeefciency.heat release, (b) total work and (c) numerical and experimental indicated thermal

4.1. Model validation

After the model development, simulations are performed basedon the input parameters of the experimental cases for enginespeeds of 1200, 2400 and 3600 rpm and engine loads of 100%and 50%. In-cylinder pressure and HRR diagrams for consideredcases are compared in Fig. 6. It can be seen that the simulation re-sults of the multi-component fuel combustion model are in goodagreement with the experimental measurements. In order to com-pare results of the developed model, simulations were also carriedout using conventional single component combustion model.Using the same chemical kinetics mechanism, the global oxidativepyrolysis reaction, Eq. (1), is considered in single component fuel,C14H28, combustion model. In-cylinder pressure and HRR resultsshow the single component combustion model predicts longerignition delay time compared to the multi-component fuelcombustion model. Comparing single and multi-component fuelcombustion models, latter model shows better agreement withexperimental pressure and HRR data.

By reference to Fig. 6 higher peak pressure and HRR are resultedby increasing engine load at each specic engine speed. In contrastto the higher engine speeds, at the lowest engine speed of1200 rpm the location of the maximum HRR is delayed in the con-sidered engine loads. This is mainly due to the timing for SOI andthe amount of injected fuel in each case where injecting more fuelat higher engine speeds and loads has been resulted in formation ofmore combustible air fuel mixtures. This trend can be seen in thepressure diagrams that the fuel tends to burn late under the lowestengine speed of 1200 rpm. Pressure curves for this engine speedhave two peaks where the pressure is dipped due to starting ofexpansion stroke exactly after TDC and created the rst peak.Although combustion was initiated in cranks before TDC in HRRdiagrams, amount of released heat was not sufcient enough toovercome cylinder pressure reduction due to expansion. However,as combustion has got stronger and higher heat was produced,pressure curve was started to go higher and created the secondpeak. It should be noted that second pressure peak tends to reduceby decreasing the engine load which is mainly due to reduced fuel

766 A. Maghbouli et al. / Applied Energy 111 (2013) 758770Fig. 8. Mean value CAD history of (a) fuel concentration history, (b) temperature, (c) sootload and 3600 rpm.and (d) NOx and their temporal contours at 7, 8, 14 and 50 CAD ATDC at engine full

amount and weak air fuel mixtures and subsequently weakercombustion.

4.2. Performance and emission characteristics of the engine under fullload

Fig. 7a shows engine accumulative heat release for three enginespeeds at full load condition. At higher speeds it can be seen moreheat is released due to higher fuel input in that particular case.Moreover, fuel tends to burn rapidly in higher speeds. Closed cyclesimulation results for the work done by the engine show that lowerengine speeds are associated with reduced output work, seeFig. 7b. Nonetheless, this trend is not observed for ITE in Fig. 7c.Both simulation and experiment results indicate that a higher ITEcan be observed for the engine speed of 2400 rpm compared tothe other engine speeds.

Using the detailed chemical kinetics mechanism incorporatedwith the uid dynamics code, in-cylinder thermodynamic proper-ties and concentration change of any chemical species and theirtemporal and spatial variation can be tracked. Fig. 8 shows men-tioned advantage of the multi-dimensional models for in-cylinderfuel concentration history, temperature and soot and NOxemissions for the engine condition of full load and 3600 rpm.Moreover, their contours are also presented in four crank angle de-grees of 7, 8, 14 and 50 ATDC. Equivalence ratio contour showsthat a considerable amount of fuel (lumped value including n-hep-tane and toluene) is trapped in piston bowl after initiation of the

injection process; where mean value of fuel concentration historyshows that massive amount of fuel is burned around 10 CAD ATDC.This is why high local temperature regions can be observed in thetemperature contour at 14 CAD ATDC. It also can be seen that com-bustion is initiated when the mean cylinder temperature reachesto 1000 K and high temperature local regions can be distinguishedat the 7 CAD ATDC contour. Mean value results of soot and NOx inFig. 8 show that soot tends to oxidize while it is encountered byhigh temperature, whereas NOx emission tends to increase and sta-bilized after 30 CAD ATDC. Soot contours illustrate fuel rich regionswith moderate local temperature have the potential of producingmore soot emission. It can be seen high local concentration of sootis available in piston bowl which coincides with equivalence ratiocontours. Nonetheless, high local concentrations of NOx are ob-served where the local temperature inside the combustion cham-ber has higher values. Figs. 9 and 10 represent contour levels oftemperature, NOx, soot, CO, CO2 and O2 for full load and three en-gine speeds at 20 CAD ATDC. For this specic crank angle in Fig. 10higher NOx levels are produced by increasing the engine speed,whereas opposite trend is predicted for the soot emission. Refer-ring to Figs. 9 and 10, it can be seen that higher local temperatureexpedites NOx formation and CO and Soot oxidation. This is alsocoincided with O2 consumption, where in high temperature localregions CO and soot are oxidized with O2 and CO2 is produced. Thistrend is pronounced at higher engine speed because considerablehigh temperature local regions are formed and are shown by thecontours.

A. Maghbouli et al. / Applied Energy 111 (2013) 758770 767Fig. 9. Temperature (K), NOx and soot (gr) contours for engine full load and speeds of 1200, 2400 and 3600 rpm at 20 CAD ATDC.

d En768 A. Maghbouli et al. / ApplieFig. 11 illustrates mean concentration-CAD history of NOx, soot,CO, CO2 and UHC emissions compared to measured experimentaldata at 100% engine load and three considered engine speeds. Asit can be seen acceptable agreements are achieved. Tangible pro-duction/destruction of the mentioned chemical species is observedfrom 10 CAD to 30 CAD ATDC; however, after 60 CAD ATDC con-centrations are stabilized. By referring to Fig. 11, nal UHC emis-sion levels for three engine speeds are almost zero both insimulation and experiments. Moreover, under full load conditionNOx tends to have slightly higher levels by increasing engine speed,whereas soot shows opposite trend. It also shows that there isaccumulative trend in CO2 production, whereas CO starts to oxidizearound 20 CAD ATDC at 3600 and 2400 rpm. This trend is not ob-served at the engine speed of 1200 rpm where for this enginespeed CO oxidation is initiated after 35 CAD ATDC. Mentionedtrend at the lowest engine speed coincides with cylinder pressureand HRR results and mainly is because of lower in-cylinder tem-perature and heat release.

4.3. Performance and emission characteristics of the engine under midload

Fig. 12 shows simulation results of accumulative heat releaseand work done at mid load with ITE at three engine speeds. Com-paring to engine full load results in Fig. 7a, it can be seen that bydecreasing engine load, heat release and output work are reduced.For instance: at engine speed of 3600 rpm accumulative heat

Fig. 10. CO, CO2 and O2 (gr) contours for engine full load anergy 111 (2013) 758770release reaches 1020 J under 50% load comparing to 1680 J of fullload operating condition. Same as the engine full load condition,speed of 1200 rpm has the longer ignition delay which is capturedby accumulative heat release diagrams. The reason for this is thelean air/fuel mixtures applied for 1200 rpm at the mid engine load.Furthermore, ITE diagrams for 50% load show that the engine tendsto have higher efciency at 2400 rpm. It should be noted that athigher engine speed of 3600 rpm, ITE tends to drop by 5% fromfull load condition to the mid engine load, compare Fig. 7c withFig. 12c.

Bar chart in the Fig. 13 illustrates the normalized emission datafor 50% load condition. In order to make comparison between sim-ulated engine emissions data at 50% load to the engine full loadcondition, bar chart in Fig. 13 normalized by the emission valuesat their corresponding engine speeds at 100% load condition. Con-sidering this normalization, under mid load condition, most ofemissions are less than 100% load except NOx at 2400 and1200 rpm, and soot only at 1200 rpm. In addition, at higher enginespeeds of 2400 and 3600 rpm considerable reduction on CO2, COand UHS emissions are observed where the amount of mentionedemissions is halved at mid engine load.

5. Conclusions

(1) Multi-dimensional CFD calculations were performed bydeveloping a comprehensive multi-component fuel combus-tion model through integrating CHEMKIN II chemistry solverinto the KIVA-4 code in order to study combustion and

d speeds of 1200, 2400 and 3600 rpm at 20 CAD ATDC.

Fig. 11. Concentration-CAD history of considered emissions and comparison withthe measured experimental emission data under full load condition.

Fig. 12. Numerical simulation results of performance characteristics for 50% load: (a) accthermal efciency.

50% Load

Nor

mal

ized

con

cent

ratio

n [-]

Engine speed [rpm]

Fig. 13. Normalized concentration of considered emissions for three engine speedat 50% load condition.

A. Maghbouli et al. / Applied Energy 111 (2013) 758770 769emission characteristics of a D.I diesel engine. Application ofmulti-component combustion model had led to accuratesimulation methodology and better validity of the resultscompared to the experimental data.

(2) The developed model was capable of modeling variety ofoperating conditions where diverse range of equivalenceratio was applied in the simulated cases. A good agree-ment was achieved for predicted cylinder pressure, HRRand emission results comparing to the conductedexperiments.

(3) It has been observed that under low speed and mid loadoperating conditions, engine incurs weak combustion dueto the reduction in the amount of injected fuel and reducedin-cylinder temperature. This was shown in pressure andHRR diagrams where fuel was burnt late after the injectionprocess.

(4) Closed cycle simulation results show that if engine load andspeed were reduced; output power was decreased as well.However, same trend was not reported for the ITE as themiddle engine speed in considered engine loads had slightlyhigher efciency.

(5) Comparing to the full load operating condition, CO, CO2 andUHC emissions had lower concentrations at mid engine load.However, further reduction in the engine speed at midengine load has resulted in production of higher NO andxsoot emissions.

umulative heat release, (b) total work and (c) numerical and experimental indicated

References

[1] Heywood JB. Internal combustion engine fundamentals. New York: McGrawHill Book Company; 1988.

[2] Khair M, McKinnon DL. Performance evaluation of advanced emission controltechnologies for diesel heavy-duty engines. SAE paper 1999-01-3564; 1999.

[3] Yang F, Minggao O. Experimental research on EGR in a diesel engine equippedwith common-rail injection system. SAE paper 2003-01-0351; 2003.

[4] Kouremenos DA, Hountalas DT. The effect of EGR on the performance andpollutant emissions of heavy duty diesel engines using constant and variableAFR. SAE paper 2001-01-0198; 2001.

[5] Roy MM. Effect of exhaust gas recirculation on combustion and odorousemissions in direct injection diesel engines. SAE paper 2008-01-2482; 2008.

[6] Cho K, Han M, Wagner RM, Sluder CS. Mixed-source EGR for enabling high-efciency clean combustion modes in a light-duty diesel engine. SAE paper2008-01-0645; 2008.

[7] Wang Y, Li G, Zhu W, Zhou L. Study on the application of DME/diesel blends ina diesel engine. Fuel Process Technol 2008;2008(89):127280.

[8] Papagiannakis RG, Hountalas DT. Experimental investigation concerning theeffect of natural gas percentage on performance and emissions of a DI dual fueldiesel engine. Appl Therm Eng 2003;2003(23):35365.

[9] Soylu S. Examination of combustion characteristics and phasing strategies of anatural gas HCCI engine. Energy Convers Manage 2005. http://dx.doi.org/10.1016/j.enconman.2004.02.013.

[10] Scott B, Dennis F. Development and validation of a quasi-dimensional modelfor HCCI engine performance and emissions studies under turbochargedconditions. SAE paper 2002-01-1757; 2002.

[11] Komninos NP, Hountalas DH, Kouremenos DA. Development of a new multi-zone model for the description of physical processes in HCCI engines. SAEpaper 2004-01-0562; 2004.

[12] Kouremenos DA, Rakopoulos CD, Hountalas DT. Multi-zone combustionmodeling for the prediction of pollutants emissions and performance of DIdiesel engines. SAE paper 970635; 1997.

[13] Rakopoulos CD, Hountalas DT. Development and validation of a 3-D multi-zone combustion model for the prediction of DI diesel engines performanceand pollutants emissions. SAE paper 981021; 1998.

[14] Hountalas DT, Kouremenos DA, Mavropoulos GC, Binder KB, Schwarz V. Multi-zone combustion modeling as a tool for DI diesel engine development andApplication for the effect of injection pressure. SAE paper 2004-01-0115; 2004.

[17] Ehleskog R, Golovitchev VI, Denbratt I, Andersson S, Rinaldini CA.Experimental and numerical investigation of split injections at low load inan HDDI diesel engine equipped with a piezo injector. SAE paper 2006-01-3433; 2006.

[18] Helmantel A, Golovitchev VI. Injection strategy optimization for a light duty DIdiesel engine in medium load conditions with high EGR rates. SAE paper 2009-01-1441; 2009.

[19] Shuai S, Abani N, Yoshikawa T, Reitz RD, Park SW. Evaluation of the effects ofinjection timing and rate-shape on diesel low temperature combustion usingadvanced CFD modeling. Fuel 2009;2009(88):123544.

[20] Amsden AA, Ramshaw JD, ORourke PJ, Dukowicz JK. KIVA: a computerprogram for two- and three-dimensional uid ows with chemical reactionsand fuel spray. Technical report LA-10245-MS. NM (USA): Los Alamos NationalLab.; 1985.

[21] Beale JC, Reitz RD. Atomization Spray 1999;9:62350.[22] ORourke PJ, Bracco FV. Modeling drop interactions in thick sprays and a

comparison with experiments. In: Stratied charge automotive engines. I.Mech E. Conference Publications 19809; 1980. p. 10116.

[23] Bergman M, Golovitchev VI. Chemical mechanisms for modeling HCCI withgasoline and diesel oil surrogates. In: Proceedings of the VII. congress onengine combustion processes; 2005. p. 32334 [Munchen, 1516 March].

[24] Chen W, Shuai S, Wang J. A soot formation embedded reduced reactionmechanism for diesel surrogate fuel. Fuel 2009;2009(88):192736.

[25] Ismail HM, Ng HK, Gan S, Lucchini T, Onorati A. Development of a reducedbiodiesel combustion kinetics mechanism for CFD modelling of a light-dutydiesel engine. Fuel 2012. .

[26] Kee RJ, Rupley FM, Miller JA. CHEMKIN-II: a FORTRAN chemical kineticspackage for the analyses of gas phase chemical kinetics. Sandia Report SAND89-8009; 1989.

[27] Brown PN, Byrne GD, Hindmarsh AC. VODE: a variable coefcient ODE solver. JSci Stat Comput 1989;10:103851.

[28] Maghbouli A, Saray RK, Shafee S, Ghafouri J. Numerical study of combustionand emission characteristics of dual-fuel engines using 3D-CFD modelscoupled with chemical kinetics. Fuel 2012. .

[29] Maghbouli A, Shafee S, Saray RK, Yang WM, Hosseini V, An H. A multi-dimensional CFD-chemical kinetics approach in detection and reduction ofknocking combustion in diesel-natural gas dual-fuel engines using local heatrelease analysis. SAE Int. J. Engines 2013;6(2). .[30] Dr. Randy P. Hessel, senior Scientist. ERC home page: .

An advanced combustion model coupled with detailed chemical reaction mechanism for D.I diesel engine simulation1 Introduction2 Experimental procedure2.1 Engine setup and specifications2.2 Experimental test cases

3 Multi-dimensional modeling3.1 3D-CFD modeling tool3.2 Multi-component fuel combustion model3.2.1 Insights to the applied chemical kinetics mechanism3.2.2 Mathematical framework and programming details of the multi-component fuel combustion model

3.3 Engine geometry and computational mesh

4 Results and discussion4.1 Model validation4.2 Performance and emission characteristics of the engine under full load4.3 Performance and emission characteristics of the engine under mid load

5 ConclusionsReferences