Energy Conversion and Management 50 (2009) 1902–1907

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Energy Conversion and Management

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A parametric study on the performance parameters of a twin-spark SI engine_Ismail Altın a, Atilla Bilgin b,*

a Trabzon Vocational School, Karadeniz Technical University, 61300 Trabzon, Turkeyb Department of Mechanical Engineering, Karadeniz Technical University, 61080 Trabzon, Turkey

a r t i c l e i n f o

Article history:Received 30 August 2007Received in revised form 19 September 2008Accepted 27 April 2009Available online 23 May 2009

Keywords:SI engine combustion modelingTurbulent flame propagationTwin-spark enginesEngine performance

0196-8904/$ - see front matter � 2009 Elsevier Ltd. Adoi:10.1016/j.enconman.2009.04.025

* Corresponding author. Tel.: +90 462 3772910; faxE-mail address: bilgin@ktu.edu.tr (A. Bilgin).

a b s t r a c t

A spark-ignition (SI) engine cycle model was used to study the effects of spark plug location on a twin-spark plug SI engine performance. A two-zone quasi-dimensional combustion model with a sphericallydeveloping flame propagation assumption was applied. Constructed simulation can be used for eithersingle- or twin-spark plug configuration. For the twin-spark arrangement, spark plugs were consideredto be located diametrically opposite to each other on cylinder head axisymmetrically. According todimensionless distance from the cylinder center to spark plug location on cylinder head, rsd = rs/R, fivelocations (rsd = 0, 0.25, 0.50, 0.75, and 1.0) were considered. Inevitably rsd = 0 corresponds to the single-spark arrangement that the plug is located at the center. To comparison, single-spark plug configurationswere also considered for other selected spark plug locations. From the result of the study it was foundthat centrally located single-spark plug arrangement gives the best engine performance and fuel econ-omy, while for the all the other spark-plug locations away from the center twin-spark arrangement favor-able to the single-spark plug configuration.

� 2009 Elsevier Ltd. All rights reserved.

1. Introduction

Optimum combustion chamber design is one of the mostlystudied objects of internal combustion engines. There have beennumerous studies in the existing literature performed on optimiza-tion of the combustion chamber. However, considerable differ-ences in today’s production engines show that this problem isstill unresolved [1]. This is because there are many options forcombustion chamber shape, number and location of spark plugs,size and number of intake and exhaust valves, etc., and, there isno single solution to this complex multigoal problem. During pastfew decades, however, a consensus has developed on engine com-bustion chamber design that satisfying ‘‘rapid combustion” crite-ria. A chamber design in which the ‘‘fuel burning process” takesplace faster, i.e., it occupies a shorter crank angle interval at a givenengine speed, results directly in efficiency gain due to theapproaching of the theoretical Otto-cycle. Rapid combustion pro-duces a more robust and repeatable combustion pattern that per-mits operation with a significantly large amount of exhaust gasrecirculation (EGR) or with a very lean mixture, without deteriorat-ing engine operation or stability [2]. Rapid combustion in com-bined with heavy EGR or very lean mixtures has a potential toachieve greater emission control in combustion chamber alongwith some improvement in fuel economy due to reduced pumpingwork, reduced gas temperature and reduced heat loss [3,4]. Rapid

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combustion also increases resistance to knock that can allow thefuel economy associated with higher compression ratios to be real-ized [5].

There are a number of different combustion chamber designtechniques for accomplishing rapid combustion, and they gener-ally can be categorized into two groups: (1) affecting the fluid mo-tions in the chamber to increase burning rate via increasingturbulence level, and (2) altering the combustion chamber geome-try to increase burning rate via decreasing the flame propagationdistance and increasing the frontal area of the flame [5,6]. The firstone involves the generation of more turbulent charge motion byusing intake flow restrictions, or combustion chambers with largeswirl or squish regions. Increased turbulence level results in highertransport rates within the flame, and increased distortion of theflame front, both of which increase the burning rate. Increases inthe pumping work and heat losses can be considered as negativesides of the first approach. The second approach, on the other hand,requires optimization of the chamber shape and location of thespark plug and/or the adoption of multipoint ignition [3].

The history of using multipoint ignition, especially twin-sparkplugs per cylinder, was longer than three decades. Most of thestudies available in the literature are experimental [3,4,7–15],and therefore, have limited opportunities for parametric investiga-tions. Quader [7], for example, used a specially modified combus-tion chamber installed between the flat head and cylinder block.Spark plugs were mounted to the side access ports of this chamber.He found that fast burn was achieved by dual-spark plug ignition,and lean limits of stable operation were extended to leaner

Nomenclature

A area wetted by burned and unburned gases (m2)Cp specific heat coefficient at constant pressure

(kJ kg�1 K�1)Fs stoichiometric fuel-air ratio for gasoline (dimensionless)HL total enthalpy loss (kJ)Hu lower heating value (kJ kg�1)m in-cylinder charge mass (kg)p pressure (bar)QL total heat loss (kJ)rs distance from the center to the spark plug location (m)rsd dimensionless spark-plug location (rsd = rs/R)R cylinder radius (m), gas constant (kJ kg�1 K�1)S stroke length (m)SL laminar flame speed (cm s�1)SL laminar flame speed reference value at (To = 273 K and

po = 1 atm) (cm s�1)ST turbulent flame speed (cm s�1)T temperature (K)u specific internal energy (kJ kg�1 K�1)

v specific volume (m3 kg�1)V instantaneous cylinder volume (m3)Vh stroke volume (m3)W work done (kJ)xb burned mass fraction (dimensionless)

Greek symbols/ equivalence ratio (dimensionless)x angular speed (s�1)h crank angle (degree or deg)q density (kg m�3)

Subscriptsb burnedi indicatedu unburnedw wall

_I. Altın, A. Bilgin / Energy Conversion and Management 50 (2009) 1902–1907 1903

mixtures with dual-spark plugs. Kuroda et al. [3] experimentallyoptimized combustion chamber shape and spark plug locationsto equalize the flame propagation from two spark plugs. Theyshowed that fast burn overcomes the slow burn limitation of con-ventional engines and greatly extends the stable combustion rangeunder heavy EGR conditions, with resulting marked reduction inNOx emission, and improved fuel economy. Hillyer and Wade [8],and Scussel et al. [4] carried out an experimental test programon the Ford PROCO stratified charge engine having a combustionbowl in piston, and dual-ignition system. They found that dual-ignition system produces reliable, misfire-free operation with thedilute mixtures and high EGR rates which result in NOx controlcapability coupled with limited HC and CO control. Witze [9] per-formed an experimental study to investigate the trade-off that ex-ists between spark locations and swirl rate. The goal of his studywas optimization for the fastest burn.

There are limited numbers of theoretical simulation studies onusing twin-spark plugs in SI engines [16,17], and only a few ofthem are based on a quasi-dimensional thermodynamic combus-tion models [17]. Bozza et al. [17] carried out both experimentaland theoretical analyses on a twin-spark SI engine equipped witha variable valve-timing device. Their aim is investigating propercombination of variable valve timing device position (and henceEGR level) and spark advance for different engine operating condi-tions. Therefore they didn’t change the locations of spark plugs intheir study.

As seen above, there is a scarcity in the literature on using ofthermodynamic based cycle simulations of twin-spark SI enginesfor investigating the effects of spark plug locations on engine per-formance parameters. In order to make contribution in this field,a quasi-dimensional thermodynamic cycle model was constructedin this study, to investigate the effects of varying of spark plug loca-tion on the performance parameters of an SI engine. The model canalso be used to study on single-spark SI engines to make compari-sons between twin-spark and single-spark configurations. Detailsof the thermodynamic model were given in the following section.

Table 1Constants in laminar flame speed relationship.

/ SLo (cm/s) a b

0.8 0.192 2.27 �0.17

2. Thermodynamic cycle model

The two-zone thermodynamic model used in this study followsthe development of zero dimensional model of Ferguson [18]. In

zero dimensional models, such as Ferguson’s model, the heat-re-lease rate was determined by some empirical functional relationssuch as Wiebe-function or cosine burn-rate formula [19]. In thisstudy, however, the model of Ferguson was reconstructed to takeinto account the flame development process in the cylinder. Todo this, it is needed for calculation of laminar and turbulent flamespeeds, and geometric features of the twin flame fronts. The lami-nar flame speed (SL) is calculated based on the correlation given byMetghalchi and Keck [20] in the form of a power law.

SL ¼ SLoðTu=ToÞaðp=poÞbð4:7f 2 � 4:1f þ 1Þ ð1Þ

The turbulent flame speed (ST) was modeled by the followingequation [21]:

ST ¼ SLfðqu=qbÞ

½ðqu=qbÞ � 1�xb þ 1ð2Þ

Values for SLo, a, and b, in Eq. (1), for a typical gasoline fuel aregiven in Table 1. f in Eq. (2) is a turbulent flame factor, and definedwith the empirical formula f = 1 + 0.0018n depending on the en-gine speed, n. Geometric features of the flame front (such as flamefront surface area, enflamed volume (the volume behind the flamefront), and wetted combustion chamber surface area by the burnedgases) were adopted from the relationships originally developed byBilgin [2] for twin-spark SI engines.

The model assumes that, during the compression stroke, cylin-der charge is made of a non-reactive mixture of fuel, air, and resid-ual gases, homogenous in composition with uniform pressure andtemperature. The properties of this mixture were calculated byusing the subroutine FARG, originally developed by Ferguson[18]. After initiation of combustion at a specified crank angle(CA) before top dead center, the second and twin burned-gas zonesare created at the positions of twin-spark plugs. While the combus-tion is continuing, the burned-gas zones grow spherically through-out the combustion chamber. The burned- and unburned-gas

Fig. 1. Schematic of the flame geometry for a twin-spark SI engine, Ref. [2].

-140 -105 -70 -35 0 35 70 105 1400

10

20

30

40

50

60

PredictedRef. 23

Gasolinen=3000 rpmφ=0.986ε=8.5θs=-25 deg.rsd=0.3

Crank angle, deg.

Cyl

inde

r pr

essu

re, b

ar

Fig. 2. Comparison of predicted cylinder pressure variation with experimental data.

1904 _I. Altın, A. Bilgin / Energy Conversion and Management 50 (2009) 1902–1907

zones are assumed to be separated by infinitesimally thin flamefronts. At any instant during the flame propagation process, thetwin spherical flame fronts can be in contact with cylinder head,cylinder wall, piston crown, and each other, as shown schemati-cally in Fig. 1 [2]. It was assumed that the burned-gas zones arechemically equilibrium in composition with uniform temperature.The composition and properties of the combustion products in theburned-gas zones were determined by using ECP subroutine of Fer-guson [18]. The pressure was assumed uniform throughout thecombustion chamber during the combustion period.

The following equations of the model, mainly based on the firstlaw of thermodynamics and ideal gas equation of state, were inte-grated from the beginning of the compression stroke (�180 �CA) tothe end of expansion stroke (180 �CA), by using 4th order Runge–Kutta method with 1 �CA increment of the time step [18,22]:

_p ¼ Aþ Bþ CDþ E

ð3Þ

_Tb ¼�hgAbðTb � TwÞ

xmCpbxb

þ vb

Cpb

@ ln vb

@ ln Tb_pþ hu � hb

xbCpb

_xb � ðxb � x2bÞ

CL

x

� �ð4Þ

_Tu ¼�hgAuðTu � TwÞxmCpu

ð1� xbÞþ vu

Cpu

@ ln vu

@ ln Tu_p ð5Þ

_W ¼ p _V ð6Þ

_Q L ¼hg

x½AbðTb � TwÞ þ AuðTu � TwÞ� ð7Þ

_HL ¼CLmx½ð1� x2

bÞhu þ x2bhb� ð8Þ

where

A ¼ 1m

_V þ VCx

� �ð9Þ

B ¼ hg

xmvb

Cpb

@ ln vb

@ ln Tb1� Tw

Tb

� �Ab þ

vu

Cpu

@ ln vu

@ ln Tu1� Tw

Tu

� �Au

� �ð10Þ

C ¼ �ðvb � vuÞ _xb � vb@ ln vb

@ ln Tb

hb � hu

CpbTb

_xb �ðxb � x2

bÞCL

x

� �ð11Þ

D ¼ xbv2

b

CpbTb

@ ln vb

@ ln Tb

� �2

þ vb

p@ ln vb

@ ln p

" #ð12Þ

E ¼ 1� xbð Þ v2u

CpuTu

@ ln vu

@ ln Tu

� �2

þ vu

p@ ln vu

@ ln p

" #ð13Þ

Table 2SI engine parameters.

Cylinder diameter (D), m 0.1Stroke (S), m 0.08Compression ratio (e), dimensionless 10Connecting rod length (lb), m 0.0125

3. Model validation

It is necessary validation by showing the predicted values are inagreement with available experimental data, before use of a simu-lation for a parametric investigation. In a quasi-dimensional enginecycle simulation, the comparisons of predicted values are made,

generally, with pressure-crank angle data or burned mass frac-tion-crank angle data. In this study, the predicted cylinder pres-sure-crank angle variation was compared with the availableexperimental data given by Benson and Baruah [23]. It can be seenfrom Fig. 2 that, the predictions are in good agreement with theexperimental data of Benson and Baruah. The error between thepredictions and experimental data is about 8% at the peak pressureregion.

4. Calculation of the engine performance parameters

The constructed simulation code needs inputs such as enginespeed n, equivalence ratio /, distance of ignition point from thecylinder axis rs, spark plug number, spark advance angle hs, proper-ties of fuel, ambient pressure and temperature. After the comple-tion of the cycle, indicated performance parameters, such asmean indicated pressure (Pmi), indicated thermal efficiency (gi),indicated specific fuel consumption (ISFC or bi), and indicatedpower (Ni) can be determined by using the following equations[22,24]:

pmi ¼W i

Vhð14Þ

gi ¼pmiRTo

Fs/Hupogvð15Þ

bi ¼3600Hugi

ð16Þ

Ni ¼pmiVhzn

k60ð17Þ

In the above equations, volumetric efficiency (gv) was taken as85% [23]. As known, the coefficient k in Eq. (17) depends on thestroke number of the engine cycle and the value is 2 for the four-stroke engine.

0 0.25 0.5 0.75 1

rsd

0.2

0.24

0.28

0.32

0.36

0.4

Indi

cate

d th

erm

al e

ffic

ienc

y

SSTS

n=2000 rpmθs=MBT

Fig. 4. Variations of indicated thermal efficiency versus spark plug locations for2000 rpm.

_I. Altın, A. Bilgin / Energy Conversion and Management 50 (2009) 1902–1907 1905

5. Results and discussion

Geometric features of the engine used in the numerical applica-tions are given in Table 2. In this study, all applications were per-formed at wide-open throttle condition. Mean indicated pressure,indicated thermal efficiency, indicated specific fuel consumption,and indicated power were predicted for both single- and twin-spark configurations at various spark-plug locations. The compari-sons were carried out at five spark-plug locations. Because of thedimensionless distance of the spark plugs from cylinder center isdefined as rsd = rs/R, the examined spark-plug locations becomersd = 0.0, 0.25, 0.50, 0.75, and 1.0. Here rsd = 0.0 corresponds, inevi-tably, to a centered location of single-spark plug while all the otherspark plug locations were examined for single- and twin-sparkplugs. Equivalence ratio, and compression ratio were chosen as /= 0.8 and e = 10, respectively. MBT (Maximum Brake Torque) tim-ing was selected for spark advance.

The effects of engine speed and spark location on the indicatedthermal efficiency are shown in Fig. 3, for single- and twin-sparkconfigurations. The best results for gi were obtained for centrallylocated single-spark plug and moving the plug away from the cen-tral location results in decreases in gi values depend upon the pluglocation, for both single- and twin-spark engines. Although thevariations are nearly linear with respect to plug location for bothsingle and dual plug cases, the dependence of thermal efficiencyon plug location is considerably stronger for single-plug case thantwin-spark case, as given in Fig. 4. This means that if the engine hastwin-spark plugs, then the more centrally locations of these plugsare not as important as the case where an engine having only sin-gle-spark plug. As shown in the figure, even sidewall twin-sparkconfiguration (rsd = 1.0) gives closer indicated thermal efficiencyvalue (32.97%) to single-spark configuration (33.8%) that locatedat rsd = 0.25. The decrement ratios of indicated thermal efficiencyvalues, because of moving spark plugs from center to sidewall, be-comes 10.33% and 31.19%, for twin- and single-spark configura-tions at an engine speed of 2000 rpm, respectively. Thesedecreases in efficiency values with increasing distance betweencylinder center and spark location result from the increases inthe combustion duration dependent on increase in flame travelinglength with moving the plug location away from center to sidewall.

1000 1500 2000 2500 3000Engine speed, rpm

0.22

0.24

0.26

0.28

0.30

0.32

0.34

0.36

0.38

0.40

0.42

Indi

cate

d th

erm

al e

ffic

ienc

y

SS@0.0SS@0.25SS@0.50SS@0.75SS@1.0

TS@0.25TS@0.50TS@0.75TS@1.0

θS=MBT

Fig. 3. Variations of indicated thermal efficiency with engine speed for differentspark plug locations.

Increase in combustion duration, in fact, cause to two significantnegative effects. The first one is that getting away from the idealOtto-cycle. It is well known that the theoretical upper limit of anSI engine can be reached is the thermal efficiency of the Otto-cycle,in which heat input occurs in a constant volume process, i.e., an in-stant process. Therefore, any deviation from constant volume headaddition results direct in decrease in thermal efficiency. The secondone is that increases in the heat lose as the combustion durationincreases. As combustion takes longer time, contact duration ofthe hot combustion gases with cylinder walls increases which re-sults in increase in heat losses, and hence decreases in thermalefficiency.

Fig. 5 shows variation of indicated mean effective pressure (pmi)with engine speed for single- and twin-spark configurations anddifferent plug locations. The characteristics of variations of pmi

1000 1500 2000 2500 3000Engine speed, rpm

6

6.5

7

7.5

8

8.5

9

9.5

10

10.5

Pm

i, ba

r

SS@0.0SS@0.25SS@0.50SS@0.75SS@1.0

TS@0.25TS@0.50TS@0.75TS@1.0

θS=MBT

Fig. 5. Variations of mean indicated pressure with engine speed for different sparkplug locations.

0 0.2 0.4 0.6 0.8 1rsd

6

6.5

7

7.5

8

8.5

9

9.5

Pm

i, ba

r

SSTS

n=2000 rpmθs=MBT

Fig. 6. Variations of mean indicated pressure versus spark plug locations for2000 rpm.

0 0.25 0.5 0.75 1

rsd

0.16

0.2

0.24

0.28

0.32

ISF

C, k

g/kW

.h

SSTS

n=2000 rpmθs=MBT

Fig. 8. Variations of indicated specific fuel consumption versus spark plug locationsfor 2000 rpm.

1906 _I. Altın, A. Bilgin / Energy Conversion and Management 50 (2009) 1902–1907

are quite similar to the variations of gi, i.e., the best results for pmi

have been obtained with twin-spark plug cases except for centrallylocated single-spark plug. Variations of pmi values with respect todimensionless distance of the plugs from cylinder center to side-wall were also given in Fig. 6 for an engine speed of 2000 rpm.According to this figure, the decrement ratios of indicated meanpressure values, because of moving spark plugs from center tosidewall, becomes 10.27% and 31.13%, for twin- and single-sparkconfigurations at an engine speed of 2000 rpm, respectively. Asmentioned above, these decrements are sourced mainly from devi-ating from Otto-cycle, and increasing heat loss. As a result ofincreasing heat loss and combustion duration, the work that ob-tained per cylinder volume per cycle, i.e., mean indicated pressure,decrease.

1000 1500 2000 2500 3000Engine speed, rpm

0.180

0.200

0.220

0.240

0.260

0.280

0.300

0.320

0.340

ISF

C, k

g/kW

.h

SS@0.0SS@0.25SS@0.50SS@0.75SS@1.0

TS@0.25TS@0.50TS@0.75TS@1.0

θS=MBT

Fig. 7. Variations of indicated specific fuel consumption with engine speed fordifferent spark plug locations.

Variation of indicated specific fuel consumption (bi) with enginespeed for single- and twin-spark configurations and different pluglocations has been given in Fig. 7. As given in Eq. (16), bi and gi arein contrary proportional with each other. In other words, increasein thermal efficiency results in decrease in specific fuel consump-tion, and vice versa. This situation is shown in the figure. The min-imum fuel consumptions have been obtained with centrallylocated single-spark plug arrangement. Twin-spark plugs locatedcloser the center give second best fuel economy. Steeper increasein indicated specific fuel consumption with increasing radial dis-tance of plug location from center for single-spark engine thantwin-spark engine is also given in Fig. 8. Increment in bi are11.11% and 45.45% for twin- and single-spark engine, respectively.

1000 1500 2000 2500 3000Engine speed, rpm

10

20

30

40

50

60

70

Indi

cate

d po

wer

, kW

SS@0.0SS@0.25SS@0.50SS@0.75SS@1.0

TS@0.25TS@0.50TS@0.75TS@1.0

θS=MBT

Fig. 9. Variations of indicated power with engine speed for different spark pluglocations.

0 0.2 0.4 0.6 0.8 1

rsd

24

28

32

36

40In

dica

ted

pow

er, k

W

SSTS

n=2000θs=MBT

Fig. 10. Variations of indicated power versus spark plug locations for 2000 rpm.

_I. Altın, A. Bilgin / Energy Conversion and Management 50 (2009) 1902–1907 1907

Variation of indicated power (Ni) with engine speed for single-and twin-spark configurations and different plug locations hasbeen given in Fig. 9. As is the case in other performance parame-ters, centrally located single-spark engine gives the maximum Ni,while twin-spark engine, in which twin-plugs were located atrsd = 0.25, has the second best performance. Variations of Ni versusdimensionless spark-plug locations for single- and twin-sparkarrangements were given in Fig. 10. Similar to gi and pmi, the dec-rement ratios of indicated power values, because of moving sparkplugs from center to sidewall, becomes 10.29% and 31.14%, fortwin- and single-spark configurations at an engine speed of2000 rpm, respectively.

6. Conclusions

The performance of an SI engine having twin-spark ignition sys-tem with disc-shaped combustion chamber has been investigated.For this purpose, a quasi-dimensional SI engine cycle model hasbeen used. The effects of variations of spark plug locations on en-gine performance have been investigated at different enginespeeds. Comparisons were also made to the single-spark engineconfigurations for the same conditions. The following conclusionscan be drawn in the light of results obtained:

� The presented model has ability for computing the SI enginecycles for both the cases of operating with single- and twin-spark plugs.

� The centrally located single-spark configuration gives the bestperformance and fuel economy in comparison to all otherconfigurations.

� If central location of spark plug is not possible because of thesome design constraints, twin-spark plug configurations can bepreferred. It was obtained that the twin-spark configurationsgive better performances and fuel economy than single-sparkconfigurations for all spark plug locations, except centrallylocated single-spark configuration. This is a result of faster burn-ing and lower heat losses achieved by twin-spark engines incomparison to single-spark engines.

� Using of twin-spark plugs decreases the strong dependence ofengine performance on the plug location. While moving the pluglocation from center to sidewall, for example, results in adecrease of approximately 10% in engine performance fortwin-spark engine, this becomes over 30% for single-sparkengine.

� The fuel economy has a stronger dependence on spark-pluglocation for single-spark configuration. While changing the pluglocation from center to sidewall results in an increase of 11.11%in specific fuel consumption for twin-spark engine, this becomes45.45% for single-spark engine.

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