CFD study on Spark Assisted Compression Ignition Combustion

for a Natural Gas Powered Heavy Duty Engine

Federico Pessah

Abstract

Nowadays to have clean engines able to reduce bothcarbon dioxide and pollutants emissions is of funda-mental importance for safeguarding and protectingthe environment. For this reason, natural gas is avery interesting kind of fuel, since it does not formany particle matter and the low carbon-hydrogenratio (C/H) leads to a reduced CO2 productionwith respect to the combustion performed with thesame amount of gasoline or diesel. Due to its highoctane number, the thermodynamic cycle chosenas a reference is the ”Otto cycle”, but natural gasflame front velocity is not high enough to considerthe combustion as instantaneous (which is the onetheoretically assumed), especially in heavy duty en-gines where the displacement per cylinder can reachand overcome 2 liters. To do so, different and in-novative combustion modes have been investigatedusing computational fluid dynamics: they all aimto exploit compression ignition, which in homoge-neous charge context is faster with respect to flamepropagation one, but far way more difficult to con-trol. In this work, a spark assisted compressionignition engine has been simulated, highlighting ad-vantages and drawbacks of this combustion modeand deriving some design criteria to develop thisinnovative kind of engine. At the end of the anal-ysis, a pollutants emission computation has beenperformed, comparing carbon monoxide formationand unburnt fuel mass with values obtained by anentirely flame propagation combustion engine.

Introduction

Natural gas high octane number leads the enginemanufacturers to develop a traditional Otto cycle,with a spark-plug installation at the centre of thehead of the cylinder and to perform the combustion

process through flame propagation in the whole vol-ume.

Figure 1: Mesh of the investigated engine

In figure 1, the combustion chamber of a typi-cal heavy-duty natural gas powered engine is dis-played. Injection system is a port fuel one, andcharge can be assumed homogeneous. When deal-ing with heavy duty application (where displace-ments per cylinder can easily overcome the valueof 2 liters), path to be covered by the flame is con-siderable, and hence knock occurring probability isenhanced. This is the reason why even if octanenumber of natural gas is very high (in the rangebetween 120 and 130 even if not strictly defined),compression ratio adopted in the engine is limited,affecting performance and fuel consumption. Atthe same time, a well-exploited spontaneous igni-tion can lead to adopt higher compression ratiosand a faster combustion at the same time, but itis very difficult to control and it can induce verysevere loads to the mechanical structure of the en-gine. Therefore, as first choice HCCI (homoge-neous charge compression ignition) engines weredeveloped, working with huge air excess (equiva-lence ratio φ between 0.5-0.6) and gasoline or nat-ural gas used as fuels, ignited entirely by compres-sion. This combustion mode theoretically was lead-ing to enormous efficiency improvements, but inpractice it generated very violent pressure waves

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inside the cylinder and combustion was very dif-ficult to control. These are the reasons why re-searchers switched to spark assisted compression ig-nition (SACI) combustion, where a part of the mix-ture was supposed to burn due to flame propagationand the remaining one to autoignite spontaneously.This last combustion mode can not permit such ahigh compression ratio as in HCCI engines, but thepresence of the spark-plug leads to have a bettercontrol to the combustion process itself, especiallyto not induce too high loads to the engine. In ad-dition, this consideration is fundamental especiallywhen natural gas is chosen as a fuel: its octanenumber is not commercially defined, and autoigni-tion strength can vary regions by regions dependingon the kind of natural gas composition used. There-fore, it is fundamental to detect and clearly definethe difference between knock and a well-exploitedspontaneous ignition, and then to use the obtainedresults to develop a spark assisted compression ig-nition engine.

Objectives

Objective of this investigation is to verify that thiscombustion mode can take some efficiency advan-tages to the engine: ideal Otto cycle combustion isinstantaneous, but flame propagation one can lasteven 50 CAD of engine time. In addition, a well ex-ploited spontaneous ignition leads to the possibilityto adopt higher compression ratio, increasing evenmore the theoretical thermal efficiency of the cycle.Simulation setup chosen is a 2D mesh, with axialsymmetry imposed for the whole combustion cham-ber. Some important design criteria to have a goodexploitation of the SACI (spark assisted compres-sion ignition) combustion mode have been derivedinvestigating piston shape and air-fuel ratio.

Solver setup

This investigation has been carried out by build-ing a solver in OpenFOAM R© environment, in or-der to use all the utilities available from the open-source software and to build a code suitable forthe illustrated objectives at the same time. Thealgorithm must hence consider both kinds of com-bustion modes present in the engine: flame prop-

agation has been computed exploiting the Wellercombustion model, while spontaneous ignition hasbeen simulated using tabulated kinetics with thehelp of Cantera R©. In both cases, two differentprogress variables are introduced: c and cfresh. Aprogress variable is a specific quantity implementedin the CFD solver which is initialized at 0 having amonotonous trend up to 1:

Progress variable0 → 1

As the name suggests, they are introduced to in-dicate the combustion progress: they assume thevalue of 0 if no reaction occurred, while they reachthe value of 1 whenever combustion can be con-sidered entirely completed. Therefore, a thermo-dynamic quantity which has this behavior must bechosen: during this work, c is the species formationenthalpy at a reference temperature of 298 K, as ithas a monotonous trend with the combustion pro-cess from a minimum up to a maximum value. Tomake it from 0 to 1, the progress variable has beennormalized:

cNorm =h− hmin

hmax − hmin(1)

In this way, consistency between combustion chem-ical reactions and heat released by them is guaran-teed. Therefore, the following implementation hasbeen chosen:

- c for flame propagation combustion

- cfresh for compression ignition combustion

The two phenomena by a solver point of view areclearly detached, and they can not interact one withthe other. As explained in “The Development of aNew Flame Area Combustion Model Using Con-ditional Averaging” published by H.G. Weller in1993, flame propagation combustion model is writ-ten with respect to the regress variable b, to haveconsiderable numerical advantages:

b = 1− cNormalized (2)

∇b = −∇cNormalized (3)

There is no strong conceptual difference between band c, but the former is initialized to 1 and proceedsto 0 as combustion takes place: for this reason, b

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is named ”regress variable”. To have a good rep-resentation of the flame, its velocity is computedcycle per cycle, hence also Ξ is initialized:

Ξ =AfAs

=SturbulentSlaminar

(4)

This Ξ is derived by the ratio of vortex normal vec-tor with respect to surface front one, but as illus-trated in equation 4 it can be easily re-conduced tothe ratio of the flame front turbulent velocity withrespect to the laminar one. With this settings thetwo equations model can be implemented:

∂ρb

∂t+∇ · (ρ~ub)−∇ · (µt∇b) = ρ~uSuΞ|∇b| (5)

An algebraic expression was used to compute Ξ.Spontaneous ignition instead must be consideredin a completely different way, because there is nota reaction front propagating inside the combustionchamber; at the same time, the number of pos-sible oxidation reactions for every kind of hydro-carbon are in the order of magnitude of hundredsor thousands, so in terms of computational timeit would result too expensive to implement all ofthem. Therefore, the concept of autoignition delayhas been exploited: for every pressure, tempera-ture and mixture fraction conditions, a quantity oftime is necessary for spontaneous ignition to occur.This delay time can be evaluated using an Arrhe-nius correlation:

τa = Ap−nexp(EaRT

) (6)

In the equation 6 it is possible to appreciate howhigh cylinder temperatures enhance spontaneousignition massively, as autoignition delay trend withtemperature is exponential. Even for this reason,fresh unburnt specific quantities must be numer-ically detached by flame propagation ones, as incombustion chambers at the same time unburntgases temperatures are about 1000 K while burntones have a typical temperature value of 2600 K instoichiometric mixture conditions. Therefore, evenbefore the simulation a table of kinetic chemistryis generated, as illustrated in figure 2.

Figure 2: Tabulated kinetic computation for everyfuel

Before the simulation, a set of different pressures(p), temperatures of unburnt charge (Tu), mixturefractions (Z) and exhaust gas recirculation (EGR)must be provided to the mechanism, which com-putes the chemical reaction velocity rate for everycombination of them. Since no exhaust gas recir-culation is considered in this work, the size of thetable is going to be:

Ntable = np · nTu · nZ (7)

Where n stands for the number of input valuesfor any specific quantity. In natural gas poweredengines compression ratios are high with respectto the other spark ignition thermal machines, andalso due to the presence of the temperature gra-dient generated by the flame, Tu maximum to beassumed must be high. As this table is generatedbefore the simulation, it is not possible to have aprecise knowledge of maximum unburnt tempera-ture value, but if it is underestimated the wholecomputed results are compromised; otherwise, ifTu is overestimated, there will be just an increaseof computational time needed for the table gen-eration. Once provided the described inputs, themechanism returns as output different values ofcfresh, the compression ignition progress variablesource term. In the end, the following combustion

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transport equation can be written:

∂ρcfresh∂t

+∇ · (ρ~ucfresh) + µ∇cfresh = ρcfresh · b(8)

Equation 8 reports on the left side time-derivative,convection and diffusion term as every conservationequation; on the right side the source term relatedof progress variable increasing, composed by:

- a term proportional to chemical reactions veloc-ity (cfresh)

- a term connecting spark ignition combustion tospontaneous one (b)

”b” is added to equation source term because burntmixture (having b = 0) has already reacted, and itcan not give any further heat release contribution.Once progress variable equations have been solved,the ”Update composition” script computes speciesformed by combustion from obtained cell pressure,temperature, c or cfresh and mixture fraction, andchemical products of combustion process can even-tually be obtained.

Engine presentation

Engine analysed is a natural gas powered one forheavy duty applications. To reduce computationaltime required, just the power cycle (compression,combustion and expansion) is simulated. Figure 3compares the whole CFD domain at the startingpoint of the simulation (at the intake valve closing,IVC) with respect to the top dead centre pistonposition. To further reduce computational time,axial symmetry has been imposed: in figure 4 aname is attributed for every mesh boundary butthe left one as it is the axis of symmetry, and figure5 shows how the whole region modeled is a small2◦ arc of the whole cylinder.

Figure 3: CFD domain at intake valve closing (left-wards) and at top dead centre (rightwards)

Figure 4: Cylinder walls names

Figure 5: OpenFOAM mesh top view

Displacement of the modeled cylinder is 2.14liters. Squish region designed is important to in-crease turbulence presence inside the combustionchamber, and then to have a flame propagatingquickly in the whole domain considered. Walls clos-ing the crevice regions are called “Adiabatic Pis-ton” and “Adiabatic Liner” because that region isso small that considering no heat transfer is influ-encing flame propagation simulation results. In any

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case, this part of the whole domain is going to beanalysed when dealing with knock and spontaneousignition, as it is located very far from the spark lightand the assumption made can influence simulationsresults. The most important geometrical and tur-bulence parameters are listed in table 1. As justone power cycle is simulated, turbulence presencein the engine must be initialized a priori at thebeginning of the simulation, also because no gasexchange across inlet or outlet valve is considered.As already described, intake valve closing coincideswith the beginning of the simulation while the ex-haust valve opening with the ending of it, hencea period of 299◦ is computed with respect to the720◦ of the whole Otto cycle. Compression ratioadopted is 11.7,which leads to have no importantknock problems when natural gas is used.

stroke 0.135 mbore 0.15 m

cRodLength 0.230IVC -175◦

EVO 124◦

swirl/RPM 1.5swirlAxis (0 0 1)

swirlProfile 10−5

uprime/Up 0.7lint/Bore 0.017

Table 1: Engine main geometrical parameters

This engine design has been chosen because amap of experimental data is available: as illustratedin figure 6, some working points have been providedin “Torque-Engine speed” map and their availabil-ity is essential to tune CFD solver parameters.

Figure 6: Engine map of experimental workingpoints available

As illustrated in table 2, for every working pointdrawn in figure 6 spark timing, cylinder pressure,

cylinder temperature and all wall temperatures areknown at the intake valve closing condition. In ad-dition, for all of them not just the final gross indi-cated work or heat release value are available, buteven experimental mixture pressure, heat releasedand heat transfer ongoing along the whole cycle areall plotted with respect to crank angle degrees.

1000-1710 1200-1670Spark time -9.75◦ -11◦

IVC pressure 1.72 bar 1.72 barIVC temperature 358 K 362 KTwall head 508 K 510 KTwall liner 420 K 420 KTwall piston 570 K 563 K

Table 2: Initial available pressure, temperaturesand spark timing for every experimental workingpoint represented in CFD map

Working points name is derived by the combi-nation of engine speed and torque: for instance,1000-1710 means 1710 Nm of torque at 1000 rpm.A spark ignition solver validation needs to be per-formed, in order to verify that it is able to repro-duce correctly flame propagation combustion, bothin terms of flame velocity and heat release. Sparkplug strength is kept constant (value of 3), while itsenergy release duration is kept fixed at 1.25 ms, butchanging in terms of engine time. To control sparkignition combustion flame velocity, Ξcoeff equilib-rium coefficient has been modified.

Ξ = 1 + Ξcoeff

√u′

SuRη (9)

The higher Ξcoeff , the higher flame velocity is go-ing to be. By the expression above, it is possible toappreciate also turbulence influence in increasingflame combustion velocity. These data are henceessential to validate the solver: for every Ξcoeffvalue in the range 0.4-0.7, the obtained pressurecurve has been compared with the provided exper-imental one, in order to chose the coefficient valueable to provide a good representation of the heatreleased by the flame itself.

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Figure 7: Xi equilibrium sweep

This Ξcoeff analysis has been performed in 1000-1710 working point, as it presents the lowest enginespeed available, hence the turbulence influence isnot so relevant (as the ratio of its intensity andengine rotational speeds is almost constant alongthe whole map). The Ξcoeff value finally chosenis 0.5 as it is associated the most realistic pressurecurve representation as illustrated in figure 8, andthis obtained value has been kept constant duringthis work for the whole engine map. Simulationparameters are the one illustrated in the first col-umn of table 2, and as a consequence spark timingis imposed at -9.75◦ with respect to top dead cen-tre. Represented ignition delay is slightly underes-timated, as the calculated curve drawn in red in fig-ure 9 detaches a little bit too early from the dottedexperimental ones, but this difference is not so im-portant as curve trends are very similar. The onlyappreciable difference concerns wall heat transfercurve illustrated in figure 10: up to the minimumvalue, curves are almost coincident and heat lossesrepresentation can be considered realistic; but thenthe two curves clearly detaches leading to a hugeunderestimation of the computed wall heat transferterm.Very similar results are obtained even if the otherworking point of the engine map are considered.Computed curves trends remain consisted with theavailable experimental ones, just sometimes an un-derestimation of the ignition delay is present butit is adjustable regulating spark timing value. Asanticipated, spark plug quantity of energy releasedhas been kept fixed.

Figure 8: Experimental and Computed data com-parison: cylinder pressure curve

Figure 9: Experimental and Computed data com-parison: rate of heat release curve

Figure 10: Experimental and Computed data com-parison: wall heat transfer curve

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For this validation of the flame propagation sec-tion of the solver, a natural gas entirely composedby methane has been used as a fuel, leading to noknock issues and hence no heat released by com-pression ignition.Spark advance actually influences spark ignition en-gine performances in a considerable way, and thisworking points as they are not optimized can notbe taken as a reference to develop a spark assistedcompression ignition combustion mode. Therefore,some spark-advance sweeps have been performed inorder to find the ones maximizing gross indicatedwork value.

Spark Ignition Engine Opti-mization

Anticipating the spark timing, combustion startsearlier and hence the engine can better exploit heatreleased by the mixture to increase maximum cylin-der pressure value; but if ignition is too anticipated,both compression and heat transfer losses increase,and as a consequence thermal efficiency of the en-gine gets worse.

Figure 11: Spark timing sweep: pressure curves

As illustrated in figure 11, always higher maxi-mum values are reached when ignition is advanced.Every pressure increase before top dead centre (incorrespondence of 0◦) is enhancing compressionlosses, hence the piston spends a higher quantityof energy; but, it leads to a higher underlying areaduring the expansion stroke, as clearly visible inthe figure above.

Figure 12: Spark timing sweep: wall heat transfer

Also heat losses are influenced by start of igni-tion: the more it is anticipated, the more time avail-able for the mixture to loose a part of the heat pre-viously released. This effect influences especiallythe second part of expansion stroke, when combus-tion can be considered fully completed and henceno further heat released is present. Even if resultsare plotted in pressure-volume plane, it is possibleto appreciate how the cycle gets more ideal whenignition is anticipated. As illustrated in figure 13,combustion always gets similar to a constant vol-ume one, increasing cycle underlying area.

Figure 13: Spark timing sweep: pressure-volumeplane

In 1000 rpm full load case, the spark timingwhich optimizes useful work done by the thermalmachine is -19◦. In order to compare even differentload conditions, engine efficiency has been defined.

η =GIW

cumRoHR(10)

As in equation 10, thermal efficiency is computed

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through the ratio of gross indicated work obtainedby the power cycle and the whole energy releasedby the fuel. In spark ignition engines, the load ismodified increasing or decreasing both air and fuelmass injected in the cylinder. Figure 14 has beenobtained keeping constant the engine speed, butvarying the load: all conditions have similar tem-perature maximum values, but the higher thermalinertia present in high load working point leads to alower temperature reduction, hence gross indicatedwork obtained at the end of expansion stroke in-creases. This behavior explains why turbochargersare so important for internal combustion engines,as higher initial pressures lead at the same time tohave more useful work and lower losses.

Figure 14: Temperature curves at different load

Therefore, an efficiency map considering allpoints available of the whole engine map has beenobtained. Spark advance optimization as illus-trated so far has been performed in all conditionsavailable, and this values are taken as a referenceonce spark assisted compression ignition engine hasbeen developed. As previously illustrated, at con-stant rotational speed, high loads working pointspresent a higher efficiency with respect to low-loadones: if 1200 rpm working points efficiencies infigure 15 are observed, there is a difference of al-most 3% between lowest load (17%) to the full loadconditions. Then, due to the presence of higherturbulence intensity and lower heat transfer, ef-ficiency values slightly improves when engine be-comes faster.

Figure 15: Engine efficiency map

These are hence the reference values to overcomeof the spark ignition engine analysed. From nowon, a knock analysis has been performed, in orderto verify if some autoignition issues are present andif the solver behaves correctly. Then, a spark as-sisted compression ignition has been developed, tounderstand if it can lead to appreciable advantagesto engine performances.

Natural gas knock analysis

Dealing with natural gas autoignition is alwayscomplicated because this kind of fuel octane num-ber is not commercially defined, but it can varyregion by region. Octane number is an autoigni-tion indicator, assuming the value of 100 if the fuelbehaves as isooctane, higher if it is more knock re-sistant. Natural gas usually has an octane numberbetween 120-130, and it can change between onecomposition and another. In this section, a first fo-cus has been pointed on the fuel, then on how thesolver behaves whenever its hydrocarbons compo-sition changes.Natural gas is mainly composed by methane in allover the world. CH4 is the lightest hydrocarbonavailable, a very stable molecule with an octanenumber of 130, but the presence of heavier hydro-carbons molecules considerably influences fuel be-havior in combustion reactions. Therefore, in table3 most important fuel heat releasing molecules areillustrated, with respective octane number (ON).

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HC Ch. formula LHV[MJ/kg] ONmethane CH4 50,0 130ethane C2H6 47,622 108

propane C3H8 46,35 103butane C4H10 45,75 91

Table 3: Natural gas main hydrocarbons

To find hydrocarbon molecules heavier than bu-tane is very uncommon in natural gas. If values oftable 3 are observed, LHV of the molecules do notchange so much, hence when passing from a natu-ral gas composition to another (without consideringpossible nitrogen and/or carbon dioxide presence)heat released are almost coincident. This is notvalid for octane number column, as a huge presenceof propane or butane can lead to knock. There-fore, a natural gas composition investigation hasbeen performed comparing different sources: theyall agreed that in Italy (named as “Nazionale” infigure 16) there is a huge methane prevalence, whilein the rest of Europe fuel is quite heavier.

Figure 16: SNAM provided natural gas composi-tion in more regions

As illustrated in figure above, nitrogen and car-bon dioxide are the most common inerts presentin natural gas. Since they are not involved in anycombustion reaction, their presence increases oc-tane number of the whole fuel, as chemical activ-ity of oxygen is reduced. At the same time, lowerheating value of the fuel decreases as they can nottake part to any combustion reaction. Values dis-played in figure 16 are computed in mole percent-ages, hence when passing to mass ones heavier hy-drocarbons values becomes higher. In the end, evenafter converting these numbers, almost all naturalgases commercially available have a methane masspercentage higher than 85%, with ethane as the

second main hydrocarbon. Very low mass percent-ages of propane or butane are enough to let theengine knock, as just the 4% of them can start au-toignition chain reactions. As already described,two different kinds of combustion coexist in thesolver. Flame propagation velocity has been val-idated determining the most suitable Ξcoeff , whilean external experimental study is needed to per-form natural gas knock validation. This subsectionis fully dedicated to comparison between solver re-sults and a paper published in 2012 by Jiri Vavra,Michal Takats, Vojtech Klir and Marcel Skarohlidof the Czech Technical University, named [“Influ-ence of Natural Gas Composition on TurbochargedStoichiometric SI Engine Performance”].In this study, knock tendency of a compressed natu-ral gas engine has been investigated including someadditives to a reference fuel composition. Enginehas hence been tested with a multiplicity of differ-ent fuels, in order to detect if knock is occurring ornot. If the solver can reproduce results comparablewith experimental values, validation can be consid-ered completed and the solver able to be used fordesign purposes; but it is impossible to get coinci-dent results, as some important parameters such asexperimental study combustion chamber design arenot reported. In any case, if a result consistency isdetected, validation can be considered completed.In the CFD solver, knock occurs whenever the nor-malized progress variable cfresh reaches the valueof unity, starting the so-called compression ignitioncombustion; in the experimental study, an AKRsensor has been used, returning a voltage outputproportional to knock intensity. If the knock rec-ognizing sensor signal is higher than 2 V, autoigni-tion is considered heavy, otherwise it is light. Ex-ploiting pressure curves available in the publishedpaper, conversion value between the sensor outputand knock associated pressure rise is 2.16 knockingbar/V.

CH4 96.73%C2H6 0.82%C3H8 0.44%C4H10 0.38%N2 1.44%CO2 0.19%

Table 4: Reference natural gas composition to betested

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In table 4, reference fuel composition taken fromCzech grid has been described. Some ethane,propane or butane has been added to the fuel,obtaining heavier natural gases to let the engineknock. Another important thing to be highlightedis that the engine tested in the experimental publi-cation is very similar to the CNG-heavy duty one.As already anticipated, its combustion chamber de-sign is unknown, but as suggested by table 5 bothgeometrical parameters and intake maximum pres-sures are very similar.

paper engine CNG-heavy dutyCR 12:1 11.7:1Max pressure 2.4 bar 2.44 barBore 102 mm 135 mmStroke 120 mm 150 mmIVC -125◦ -175◦

EVO 123◦ 124◦

φ 1 1

Table 5: Comparison between paper engine andCNG-heavy duty one

By a solver point of view, a different Canteratable must be generated for every kind of fuel anal-ysed.

Figure 17: Simulated AKR output sensor in ethaneaddition

As it is possible to see in figure 17, there aresome differences between the computed knock in-tensity and the experimental one. In any case, amore realistic representation would be too preten-tious, also because combustion chambers comparedhave different sizes. What is important to remark is

that results consistency is maintained along differ-ent experiments shown in figures 18 and 19, whichcan lead to an appreciable consideration regardingnatural gas spontaneous ignition representation.

Figure 18: Simulated AKR output sensor inpropane addition

Figure 19: Simulated AKR output sensor in butaneaddition

Once the solver can be considered reliable, fo-cus is pointed on how knock is reproduced in theCFD domain. As already specified, autoignitionstarts whenever cfresh normalized reaches the valueof unity. As predicted by theory and displayed infigure 20, spontaneous ignition starts in most ex-ternal region, where squish area is located. It re-quires time to start, hence a slow flame or a bigcylinder bore enhance its probability. In the fig-ure below, leftwards the flame front is displayed,while rightwards just compression ignition associ-ated progress variable is represented in the CFDdomain. In any case, compression ratio adopted

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in CNG-heavy duty engine is too low to have anappreciable autoignition, hence a new natural gascomposition has been edited in order to let the en-gine knock.

SNAM heavy Natural GasMethane 81.88 % 84 %Ethane 7.65% 8%Propane 2.28% 6%Butane 1.23% 2%Nitrogen 3.77%Carbon Dioxide 3.19%

Table 6: Natural gas composition in mass percent-ages

Figure 20: Engine flame front and spontaneous ig-nition visualization

Figure 21: Autoignition progress variable increase

In table 6, the edited fuel named as “heavy Natu-ral Gas” is reported. This kind of fuel composition

is very heavy and hence unlike to be found in ordi-nary fuel station, but it guarantees the knock pres-ence once fed in CNG-heavy duty engine. Close toit, SNAM North Europe derived natural gas com-position is reported in mass percentages, as it isthe one used from the next section on to designspark assisted compression ignition engine. Chartsplotted in figure 21 shows how spontaneous ignitionprogress variable increases in a really fast way. Inthe illustrated case, it appears after the 80% of thewhole amount of fuel has already ignited by flamepropagation. As soon as the first mixture particleauto-ignites, it releases an appreciable amount ofheat needed to start a sort of compression ignitionflame, which is oriented from the most external partof the cylinder to the most internal one. In anycase, it can not be considered a proper flame, asusually it is very violent and associated to very se-vere and unwanted pressure waves; but in terms ofcompression ignition progress variable transport, itbehaves similarly to an incredibly fast flame prop-agation combustion.

Figure 22: Pressure curves for heavy Natural Gaspowered engine

As shown in figure 22, a rotational speed sweephas been performed to verify knock presence andintensity for any available constant torque condi-tion. Compression ignition can in fact be easily de-tected by the presence of a very vertical pressure in-crease when it takes place. As displayed by the fig-ure above, this increase is heavier at low rotationalspeeds, but final part of combustion results fasterand pressure curve underlying area gets wider. Thisexplains the potentiality of spontaneous combus-tion in homogeneous charge condition, as it resultsfaster and hence more ideal. Every simulation has

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been carried in optimized spark timing condition,as previously illustrated.

Figure 23: Pressure rise curves for heavy NaturalGas powered engine

In figure 23, pressure rise of obtained curves isplotted. Even if compression ignition contributionto cylinder pressure is not enormous, maximum risedisplayed in this curve is very close to the assumedmechanical limit of the engine. In addition, afterspontaneous ignition occurred some values oscilla-tions are present: they are due to the mechanicalwave originated by this huge amount of heat re-leased, but time step and turbulence model set inthe solver can not lead to a realistic representationof these kinds of waves.

Development of a spark assistedcompression ignition engine

In previous section CNG-heavy duty engine couldknock only if powered with very heavy natural gas,unlikely to be found in ordinary fuel stations. Todevelop a SACI (spark assisted compression igni-tion) combustion, spontaneous ignition must be re-liable and under control, not depending on kind ofnatural gas used. Therefore, compression ratio hasbeen increased to have higher pressures and tem-peratures and then enhance fresh air-fuel mixtureautoignition. This leads to a redesign of combus-tion chamber, performed in such a way to increasethe compression ratio adopted but to not introducesignificant changes on piston shape.

Figure 24: Combustion chamber shapes at top deadcentre

As it is possible to see in figure 24 where origi-nal design and the new one are compared, volumeat top dead centre of increased compression ratiochamber is far way lower. Parameters such as boreand stroke are not changed, so this kind of modifi-cation practically implies just a substitution of theoriginal piston with a new one.

Figure 25: Mass of fuel injected in different com-pression ratio engines

As it is evident by figure 25, considering thesame working points (characterized by initial cylin-der pressure and temperatures values) a lower massof fuel is injected inside the cylinder. This leads toa first advantage in terms of efficiency: as shownin figure above, for the same initial pressure andtemperature values, in the CR 15 case fuel amountis more than 2% lower with respect to the referenceone at CR 11.7. In addition, since a better thermalefficiency is predicted as explained by the theoryin section 1, even gross indicated work should in-crease, leading to another efficiency benefit. Alsoknock must be taken into account: in a spark as-

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sisted engine, it needs to be controlled to be ex-ploited for a more ideal constant volume combus-tion. At this purpose, two different solvers are usedin this section:

- Xi solver: including b equation

- SACI solver: including b and cfresh equations

As briefly described, the ”Xi” solver can not con-sider autoignition, and as a consequence it is possi-ble to simulate very extreme compression ratio en-gines without any knock presence. ”SACI” solverinstead includes both flame propagation combus-tion and chemical reaction delay time, computingalso heat released by a possible mixture autoigni-tion. A comparison between these two solver re-sults is very important to isolate combustion modeinfluence from all the other parameters that affectthe power cycle, and to have more clear advantagesand disadvantages concerning just spontaneous ig-nition. As already proved, autoignition can be veryrough and violent, stressing in a dangerous way me-chanical engine resistance. Due to the almost in-stantaneous heat release, it is possible to reach veryhigh cylinder pressure values and, if bad controlled,it can turn into heavy knock. For this reason, notall the working points in the engine map can workusing a SACI combustion mode, but just the onesat half load or lower. In figure 26 displayed be-low engine map is revisited, with SACI combustiontarget points drawn in green, while entirely flamepropagation combustion target points in red.

Figure 26: Revisited engine map, showing SACIand flame propagation points

To perform the illustrated solvers results com-parison, 1200 rpm - half load working point hasbeen considered. Natural gas composition used isthe SNAM source derived North Europe one illus-trated in table 6, and compression ratio adopted is15, which is a very high value for a spark ignitionengine.

Figure 27: Pressure curves obtained by both solversin the same initial conditions

Figure 28: Pressure rise curves obtained by bothsolvers in the same initial conditions

As illustrated by figure 27, spontaneous igni-tion occurs in correspondence of maximum pres-sure value, evidently increasing SACI solver curveunderlying area. A spark timing of -16◦ has beenimposed to both the cases, as well as SNAM NorthEurope derived natural gas composition (the samethat could not autoignite in previous section) hasbeen used. Because of spontaneous ignition, pres-sure maximum value increases considerably, overthan 20% in relative terms. Especially regard-ing cumulative heat release curve of figure 29, one

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can appreciate how fast compression ignition com-bustion is in homogeneous mixture conditions: inthe case where it is included in the solver, com-bustion ends between 13-14◦ crank angle, almost20◦ before entirely flame propagation combustionis completed. This of course leads to a very severeand dangerous pressure rise, that overcomes enor-mously the already high assumed limit. As shownin figure 28, pressure rise is close to 30 bar/deg, avery high value, especially for a half load condition.

Figure 29: Cumulative heat release curves obtainedby both solvers in the same initial conditions

η GIW cum Heat lossesSACI 40.78% 2759 J -1359 J/degXi 41.89% 2827 J -927 J/deg

Table 7: SACI and Xi solvers comparison

If cumulative gross indicated works are com-pared, result obtained seems in contradiction withpressure curves previously illustrated. As shown intable 7, both gross indicated work and hence effi-ciency values are lower whenever compression igni-tion is implemented in the solver. To have a betterknowledge of work during the cycle, instantaneousgross indicated work is computed as illustrated inequation below:

GIWInst =p ·∆V

∆θ=

pi+pi+1

2 · (Vi+1 − Vi)θi+1 − θi

(11)

Where in the equation 11 ”i” stands for time stepconsidered, ”p” for cylinder pressures, ”V” for com-bustion chamber volume and, in the end, ”∆θ forthe time step used computed in crank angles, inthe order of magnitude of thousands of crank angledegrees.

Figure 30: Instantaneous gross indicated workcurves obtained by both solvers in the same initialconditions

Figure 31: Wall heat transfer curves obtained byboth solvers in the same initial conditions

Observing figure 30, the two curves are coinci-dent up to spontaneous combustion presence, whichincreases blue curve pressure values hence instan-taneous work made by burnt gases. Even if thedetach is clear, in less than 20◦ the curves inter-twine, and expansion computed by ”Xi” solver isclearly more efficient with respect to the one com-puted by ”SACI” once both combustions are com-pleted. Wall heat transfer minimum value in knockpresence is more than four times lower than the onecomputed by ”Xi” solver reported in red. Effectson temperature curves are enormous since immedi-ately after compression ignition weighted averagemaximum values are higher of 300 - 400 K, butthen burnt mixture cools down in a very fast wayand blue curve negative slopes after compressionignition are considerable. Burnt mixture has henceless energy to be exploited for almost all the sec-ond part of expansion stroke, and cumulative grossindicated work gets lower at the end of the cycle.

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This leads to a sort of paradox, because with sparkassisted compression ignition a more ideal combus-tion mode has been reached, but losses are so im-portant to considerably reduce power cycle workand efficiency.

Figure 32: Temperature curves obtained by bothsolvers in the same initial conditions

Knock is defined by theory as an abnormal com-bustion in a spark ignition Otto cycle engine thatcan be recognized by the presence of metallic noisecoming from the engine. It origins pressure waveswhich, propagating in the combustion chamber andrebounding in correspondence of cylinder walls, re-duce power cycle efficiency and hence gross indi-cated work, even if the whole combustion processneeds less time to be completed. In this thesis work,knock caused performance loss is associated to awall heat transfer increase due to turbulence gen-erated by the wave itself.

Figure 33: Turbulence intensity comparison be-tween before (left) and after (right) autoignition

Therefore, in figure 33 it is possible to see twoturbulent kinetic energy time step values of thesame case: leftwards, during flame propagation butbefore autoignition appearance; rightwards oncespontaneous ignition is almost completed and huge

heat has been released. As evident, maximum κvalues differs of more than two orders of magnitude.In a entire spark ignition context, turbulent kineticenergy reaches its maximum values very close tosquish area, which is designed to increase fluid tur-bulence and hence flame velocity; after autoignitionoccurs, κ reaches its maximum values as soon as thefluid is exiting from the squish area to go towardsinner parts of the combustion chamber, as soon asthere is more space for the pressure wave to prop-agate.

Figure 34: Pressure gradient after wave has propa-gated inside the combustion chamber, 11 deg afterTDC

This behavior is strongly connected to pressurewaves presence, which move from the most exter-nal cylinder region to the inner of the combus-tion chamber, to then rebound in correspondenceof cylinder walls. This kind of behavior is actu-ally typical of all knocks met so far, and it providesa further explanation on performance decrease en-countered. Some crank angle degrees after sponta-neous ignition is completed, pressure gradient in-side the combustion chamber becomes maximum,with pressure differences of over than 30% in mostsevere knock condition. All of it is contributing intoincreasing kinetic energy of the working fluid, andas a consequence heat transfer coefficient betweenthe fluid and the wall, explaining also why heattransfer curve and turbulent kinetic energy oneshave such a similar behavior.As a consequence, to reduce the pressure waveintensity is fundamental for a good exploitationof spontaneous ignition combustion; two differ-ent strategies are hence investigated: piston shapechanges and leaner air fuel ratio. The first one canbe explained by these last considerations, showing

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that pressure waves are enhance by squish region;the second aims to have both a far way less violentcompression ignition and to have more thermal in-ertia of gases inside the cylinder, in order to reduceheat losses.

Piston shape analysis to exploitCI combustion

It has been demonstrated how squish region is im-portant for flame properties but also detrimentalfor a good exploitation of compression ignition.Heat released from homogeneous charge autoigni-tion is far way faster than the one coming fromflame propagation; hence combustion chamber hasbeen redesigned in the external region with the aimof taking benefits of this combustion mode, keepingthe same compression ratio. Most important geo-metrical parameters such as stroke, bore or cylinderhead shape have not been changed to not imposehuge modifications to the engine, but just pistonshape is investigated.

Figure 35: Piston geometry comparison at top deadcentre

In figure 35 geometry changes are illustrated,with both images taken in correspondence of topdead centre piston position. Some more space hasbeen added in correspondence of squish area, insuch a way to expand the pressure wave originatedby compression ignition. The extension of this re-gion forces the designer to reduce the distance be-tween cylinder head and piston in the central partof the combustion chamber (in correspondence ofthe axis of symmetry), and for flame propagationthis is a huge disadvantage because flame frontsurface will be lower and combustion is predicted

slower than before. Taking in consideration theseassumptions, also a third piston geometry has beensimulated, called ”Turb”. Just this last combustionchamber design has a slightly lower compression ra-tio (14.8), and it has been edited to find a goodexploitation of flame surface properties.

Figure 36: ”Turb” piston geometry at top deadcentre

In both of these two last edited geometries, im-portance and influence of squish area have been re-markably sacrificed to exploit homogeneous chargecompression ignition, which in CNG-heavy duty ge-ometry was too violent.

Figure 37: Pressure curves associated to differentpiston geometries

From figure 37 displayed above, it is evident howcombustion is far way slower in ”Eng” and ”Turb”geometries, which have almost coincident pressurecurves. Spark ignition combustion in CNG-heavyduty engine is considerably faster but pressure risesdue to autoignition of the three cases are compara-ble. As predicted, combustion velocity is reducedbecause of the lower flame front surface but also tothe lower turbulence intensity presence, as shownin the figure below.

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Figure 38: Flame front in CFD domain associatedto different piston geometries

Figure 39: Turbulence intensity weighted averageassociated to different piston geometries

Figure 38 reported above illustrates in a goodway how CNG-heavy duty geometry presents aflame front with a higher surface, leading to con-siderable advantages for combustion. Even if com-bustion development has been slowed down, powercycle efficiency values between the three differentcases are comparable: in table 8, efficiency valuesare reported, showing how even if flame propaga-tion is this slow both ”Eng” and ”Turb” geome-tries are more efficient with respect to the CNG-heavy duty one. Therefore, a spark advance sweephas been performed for both the geometries but,as they lead to very similar considerations, just the”Turb” one is illustrated.

solver ηCNG-heavy duty SACI 40.78%Eng SACI 41.69%Turb SACI 42.01%CNG-heavy duty Xi 41.89%

Table 8: Efficiencies comparison of different pis-ton geometries, with the same initial conditions andspark timing

Figure 40: ”Turb” pressure curves in spark timingsweep

Pressure curves behavior (shown in figure 40) isconsistent with the one previously analysed. Com-pression ignition is occurring in every cycle, with anincreasing intensity as spark timing is anticipated.Maximum efficiency is reached at -20◦ of spark tim-ing, hence compression ignition in this case is notsomething to be absolutely avoided. In figure 41 itis shown autoignition development from its origin(11◦) to its ending (14◦). Changing in squish area isalso important to give a shape or compression igni-tion development, as happens with flames. Whensquish area was adopted, as soon as autoignitedmixture has more space, it developed in a messyway, originating pressure waves. With this geome-try, cfresh has a more defined shape very similar tothe one of a flame coming from most external re-gions, hence wave generated is far way lower leadingto heat losses reduction.

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Figure 41: Compression ignition evolution between11 deg and 14 deg in ”Turb” geometry

Figure 42: Rate of heat released by compressionignition comparison between different geometries

If figure 42 is observed, it is evident how dif-ferent piston geometries can influence compressionignition heat release: in both the edited ones, max-imum rate of heat released value is lower and, atthe same time, bells associated to compression ig-nition are wider, meaning it is requiring more timeto be completed. These effects combined lead to asofter combustion mode, and intensity of pressurewaves generated is far way lower. Charts reportedin figure 43 are a direct consequence of this char-acteristic: as lower pressure gradients are presentin combustion chamber, fluid kinetic energy doesnot increase that much and a lower heat amountis lost. As obtained when different solver resultswere compared, heat losses penalize work done byburnt gases in the final part of expansion stroke,after 40 CAD. In this way, advantage of the pres-ence of spontaneous ignition is maintained, and atthe same time heat losses are reduced in such away to not penalize work done during the expan-

sion stroke.

Figure 43: Wall heat transfer comparison betweendifferent geometries

With the previous CNG-heavy duty combustionchamber geometry, compression ignition was some-thing to be avoided as it decreased useful workdone by the thermal machine. In this case, as il-lustrated by the efficiency curve in figure 44, peakis located in correspondence of a spark advance of20◦, with more than 30% autoignited mixture mass.Therefore, the value of 41.89% obtained by theXi solver in a fast combustion geometry has beendefinitely overtaken, and compression ignition canbe exploited even in stoichiometric mixture chargeconditions. As previously illustrated, not all theworking points represented in the engine map canrun with this combustion mode, because it wouldresult with very high pressure rise values dangerousfor the mechanical resistance of the engine.

Figure 44: Efficiency curve for ”Turb” piston ge-ometry running at stoichiometric mixture

Design a combustion chamber with such a low

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flame propagation velocity can lead to several prob-lems when full load condition is simulated, as sparktiming needs to be delayed to avoid autoignitionand, at the same time, the flame is slow and pres-sure peak is far from the top dead centre. For thisreasons, spark assisted compression ignition com-bustion in lean conditions is investigated, as it per-mits to easily pass from a lean partial load mixtureto a stoichiometric full load one.

Lean SACI combustion

Piston geometry taken as a reference is the startingCNG-heavy duty one, as it presents the squish re-gion that speeds up combustion. Equivalence ratiochosen for this analysis is 0.8, as it permits to haveconsiderable air excess but, at the same time, tocompletely avoid flame extinguish. Reference effi-ciency value is always the 41.89% obtained simu-lating 52% load high compression ratio engine withthe Xi solver. As thermal efficiency is the real tar-get of this investigation, amount of fuel injectedas been kept the same for all the illustrated cases:this leads to an IVC pressure increase to have thewanted air excess.

Figure 45: Efficiency curve in lean conditions

To feed the engine with lean mixture leadsto appreciable efficiency improvements: maximumreached one is 42.75% taken in correspondence of-24◦ as spark timing. This solution can lead to im-plement a squish region and a compression ignitiondesign at the same time, as stresses induced to theengine are far way lower.

Figure 46: Pressure rise comparison between leanand stoichiometric mixtures in CNG-heavy dutypiston design

In figure 46, lean spark advance optimized caseis compared with the stoichiometric mixture onecomputed with the SACI solver. In lean case pres-sure rise reaches lower maximum values and it ismore distributed in time: this leads to similar con-siderations derived by figure 42, as they can bothprove a softer and well exploited compression igni-tion. When SACI and Xi solver results were com-pared, temperature has a massive decrease aftercompression ignition as a consequence of the hugeheat transfer. In temperature curves of figure 47,no huge negative slope is present. This can be con-sidered a result of a combination of different fac-tors, such as the lower adiabatic flame temperaturedue to air excess, the lower wave intensity ad previ-ously illustrated but also the higher thermal inertiapresent in the cylinder due to the increased workingfluid mass.

Figure 47: Temperature curves in spark timingsweep in lean conditions

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Therefore, to develop a lean spark assisted com-pression ignition combustion can lead to the oppor-tunity to differentiate equivalence ratio dependingon the kind of combustion mode wanted for everyworking point. If also the two illustrated editedgeometries run with some air excess, spark timingneeds to be further anticipated as flame propaga-tion is very slow. At the same time, also compres-sion ignition is considerably slowed down, leadingto an overall better combustion exploitation. Asfigure 48 shows, when even ”Eng” and ”Turb” ge-ometries are simulated, spontaneous combustion infurther slowed down. From the first simulated caseto the lean ”Turb” optimized one (having a sparktiming of -30◦) maximum rate of heat released hasbeen more than halved, leading to an always highercontrol of spontaneous ignition combustion. At thesame time, the presence of also this second kindof ignition leads to an appreciable combustion ve-locity increase even in every lean conditions. Themaximum efficiency has been obtained in ”Turb”geometry fed with lean air-fuel mixtures, assumingthe value of 43.34%, (which is a massive increasefor a half load 1200 rpm case, as the starting onein an entirely methane flame propagation combus-tion context was 40.09%, as show in figure 15). Atthe same time, what has been derived by this studyis that criteria for a fast flame propagation and awell exploited spontaneous ignition combustion aresomehow opposite. The overall Otto cycle becomesfar way more ideal: in figure 49 the ”Turb” leancase has been compared with the CNG-heavy dutyone with 15 as compression ratio, evidencing howthe overall cycle becomes more ideal.

Figure 48: Maximum rate of heat released compar-ison between different configurations

Figure 49: Pressure curves in pV plane

Even by figure above, the point of weakness ofthis piston configuration is the huge ignition de-lay: -30◦ as optimized spark timing at 1200 rpmis huge, and when high loads or high regimes arereached this can become an issue. For this rea-son, to perform the load sweep CNG-heavy dutypiston shape has been chosen (with a compressionratio reduction up to 14.5), running with lean SACIcombustion at half load or lower and adopting justthe stoichiometric mixture flame propagation onewhen load is higher. All efficiency points displayedin figure 50 have been optimized with respect tospark advance. Therefore, the resulting curve is thematching of two different trends, but SACI pointshave always higher (or almost equal) efficiency withrespect to the others. In any case, spark timing injust flame propagation working points must be con-siderably delayed (up to 10◦), and because of thathigh load working points have no efficiency benefits.

Figure 50: Efficiency-load curve of CNG-heavyduty engine at increased compression ratio

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Pollutants analysis

Running the engine lean could be a problem for thethree-phase catalyst since its conversion efficiencyis optimized in stoichiometric air-fuel mixture con-dition, in which correspondence values about 95%can be reached and overtaken, consequently de-creasing pollutant emissions of the whole engine.In this CFD solver after-treatment systems are notconsidered, hence focus is pointed on the carbon-monoxide formation during the combustion processand on the amount of unburnt hydrocarbons at theend of the power cycle.

Figure 51: CO production in 1200-half load work-ing point comparison

As evident by the figure 51, air excess enhancescarbon oxidation both during and after the com-bustion process. Reducing the equivalence ratiofrom 1 to 0.8, less than the 25% of the CO isglobally present at the exhaust valve opening, andthis quantity is even predicted to decrease onceexhaust gases cross the catalyst. Carbon oxida-tion proceeds with an appreciable velocity until tothe end of the expansion stroke (with a slowdownat around 90◦ crank angle because of gases tem-perature reduction) and the pollutants is homoge-neously distributed within the whole combustionchamber. However, not all the fuel particles canburn and then release heat: as shown by figure 52,even if at higher compression ratios a lower fuelquantity is injected in the engine (for the sameinitial pressure), unburnt natural gas quantity in-creases, probably due to a wider flame extinguishregion near to cylinder walls. Always in the samefigure, respective percentages representing the ra-tio of unburnt hydrocarbons with respect to mass

of fuel injected are reported.

Figure 52: unburnt hydrocarbons after the wholecombustion process in 1200-half load working pointcomparison

For what regards nitrogen-oxides emission, how-ever the solver can not predict NOx formation. Forsure, the oxygen excess is enhancing their produc-tion, but adopting a leaner mixture permits to havea lower adiabatic flame temperature hence decreaseNOx production. Figure 53 reported above com-pares cell by cell temperatures reached in a SACIat high compression ratio with the ones of an en-tirely flame propagation combustion. Temperatureweighted average values are higher when sponta-neous combustion takes place, but this increase isgiven by the presence of two different combustionmodes, which speed up fuel oxidation reactions.Since cell temperature maximum values in lean con-ditions are almost 200 K lower with respect to thestoichiometric case, these effects combined shouldlimit the overall NOx production during the powercycle.

Figure 53: Cell temperatures in 1200-half loadworking point comparison

As already written, these calculations were per-formed ignoring the presence of any after-treatmentsystem at the turbine outlet. In any case, three-phase catalyst conversion efficiency is function of

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NOx, CO and hydrocarbons inlet mass, and thiscan be an interesting study case for future works.

Conclusions

In conclusion, spark assisted compression ignitioncan take to several efficiency benefits, but knockneeds to be avoided. To do so, spontaneous igni-tion flame needs to be slowed down, and this canbe done in two different ways: redesign the com-bustion chamber and/or using a lean air-fuel ratio.Adopting both the solutions can increase thermalefficiency considerably, passing from the 40.09% ofentirely flame propagation CR 11.7 engine to the43.34% obtained by a CR 14.8 edited on purposegeometry. The real issue of this solution is that de-sign criteria of a fast flame propagation and onesof a well-exploited compression ignition combus-tion are somehow opposite, as the second starts inmost external part of the cylinder where squish re-gion in commercial spark ignition engines usuallyis. Therefore, another investigated strategy to ex-ploit SACI combustion is adopting some air excess:compression ratio has been increased up to 14.5,but piston shape has not been changed. Equiva-lence ratio has been reduced up to 0.8, leading to aslower compression ignition flame and consequentefficiency improvements. The real weakness of leanmixture combustion is represented by three-phasecatalyst conversion efficiency, which is usually opti-mized for stoichiometric mixture exhausts. In anycase, adopting such an air excess carbon monoxideproduction is just the 25% of the one produced bya stoichiometric mixture in the same initial condi-tions, which can be an interesting object of studyfor possible future works.

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