Engine structure modifications effect on the flow behavior, combustion, and performance characteristics of DI diesel engine

Energy Conversion and Management 85 (2014) 20–32

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Engine structure modifications effect on the flow behavior, combustion, and performance characteristics of DI diesel engine Hadi Taghavifar ⇑, Shahram Khalilarya, Samad Jafarmadar Mechanical Engineering Department, Technical Education Faculty, Urmia University, Urmia, West Azerbaijan 57561-15311, Iran

a r t i c l e

i n f o

Article history: Received 4 April 2014 Accepted 20 May 2014 Available online 12 June 2014 Keywords: Bowl radius CFD simulation DI diesel engine Engine modification Homogeneity Factor

a b s t r a c t The simulation was carried out based on 1.8 L Ford diesel engine and the geometrical modification in structure of piston were considered in terms of bowl movement and the bowl size in four equal increments. Two major conflicting parameters in combustion and engine efficiency were taken into account and visualized in contour plots as the bowl geometry was varied: (1) the air/fuel mixing process demonstrated by Homogeneity Factor and equivalence ratio, (2) combustion initiation and work delivery by heat release rate, pressure curves, and indicated thermal efficiency. A new version of Coherent Flame Model’s sub-model (ECFM-3Z) was adopted during the calculations to shed light into the combustion chemistry and reaction rate in detail. It was found that the bowl displacement toward the cylinder wall, increases the mixture uniformity (higher HF) thus higher pressure and heat release rate peak were obtained with the penalty of combustion delay which substantially reduces the effective in-cylinder pressure. Furthermore, it was demonstrated that smaller bowl size induces better squish and vortex formation in the chamber, although lesser spray penetration and flame quenching owing to the spraywall impingement reduces ignition delay. Ó 2014 Elsevier Ltd. All rights reserved.

1. Introduction Energy issues along with ever-increasing demand of environmental communities to reduce the emissions, made researchers to analyze the various parameters affecting the engine efficiency and burnt gas production. It is well known that air/fuel mixing has drastic impact on subsequent combustion, engine efficiency, and emission production [1]. Among other methods to achieve a homogenous mixture, combustion chamber modification in structure has proved to be excellent scheme in increasing efficiency and emission reduction [2–7]. Recently Ji et al. [8] conducted a numerical study to investigate the role of bowl geometry on combustion and emissions of biodiesel–fueled diesel engine. They concluded that the application of the bowl with less surface area is preferred at low engine speed. Jaichandar et al. [4] studied the influence of injection timing and bowl geometry with two different configurations. They showed that brake thermal efficiency was enhanced by 5.64% while brake specific fuel consumption was decreased by 4.6% when TCC (Toroidal Combustion Chamber) geometry was used instead of HCC (Hemispherical Combustion Chamber). More

⇑ Corresponding author. Tel.: +98 441 2770508; fax: +98 441 277 1926. E-mail addresses: [email protected], [email protected] (H. Taghavifar). http://dx.doi.org/10.1016/j.enconman.2014.05.076 0196-8904/Ó 2014 Elsevier Ltd. All rights reserved.

recently, the combustion chamber geometry was optimized for compression ignition engine [5]. They were able to optimize the chamber configuration in terms of cup depth and Bezier curves with micro-genetic algorithm. Wei et al. [9] has investigated swirl ratio impact on combustion phase of DI diesel engine and concluded that better air–fuel mixture and lowest NOx fraction pertains to swirl ratio of 0.2 and 0.8, respectively. In addition, the influence of combustion chamber geometry was investigated on mixture preparation in a CNG direct injection SI engine [10]. They found that pentroof cylinder head can induce a non-symmetrical flow field inside cylinder. Raj et al. [11] conducted study concerning energy efficient piston configuration for effective air motion, wherein four configurations namely flat, inclined, centre bowl, and inclined offset bowl pistons were investigated. The flat bowl piston showed higher TKE, and turbulence intensity compared to other cases. More recently, fair share of emphasis was placed on survey of simultaneous effect of alternative fuels and EGR rate on combustion and emission characteristics of diesel and SI engines [12,13]. Pang et al. and Kiplimo et al., [14,15] implemented temporal and spatial simulation of soot evolution in diesel engine. They demonstrated that soot cloud mostly distributed towards the cylinder wall when a large separation of 20° was used. They found optimum spray targeting spot for lower HC, CO, and soot but higher NOx without EGR effect. Maghbouli et al. [16] used 3DCFD simulation to study combustion and emission characteristics

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Nomenclature ATDC AFR CAD D1–D4 U, ER HCP HF HRR IMEP ISFC

after top dead center air fuel ratio crank-angle (degree) four typical Distance of the bowl outer wall(m) equivalence ratio high central pedestal Homogeneity Factor heat release rate (J/deg) indicated mean effective pressure (MPa) indicated specific fuel consumption (g/kW h)

IVC ITE LHV M Ncell P R1–R4 T TKE Dmi

injection valve closing indicated thermal efficiency (%) lower heating value (MJ/kg) total mixture mass (kg) number of cells pressure (MPa) four typical bowl radius (m) temperature(K) turbulent kinetic energy (m2/s2) mixture mass at i-th computational cell

are contoured in a cross section of the cylinder at different CA positions. Furthermore, NOx emission, engine performance, and vortex formation were discussed during the paper. 2. Methodology

Fig. 1. Meshing model of domain.

of dual-fuel engines, thereby IMEP and ITE were increased, and NOx and CO levels were decreased by higher pilot fuel employment. As literature survey implies, no study was devoted to consider the bowl movement and radius impact on mixture formation in terms of Homogeneity Factor, combustion initiation (quality), and emissions. The current study deals with the engine structure modification in piston head shape at constant 1500 rpm engine speed. The results are graphed and commented upon at detail-oriented manner. The important flow and combustion characteristics

A 1.8 L Ford diesel engine has been adopted for simulation purpose. Adequately refined mesh domain with 108,456 cell numbers is presented in Fig. 1 at TDC position. The modifications in engine structure were proposed in two frameworks according to Fig. 2: in the first case, the bowl distance from HCP center was varied from D1 to D4 in 0.005 m paces (D1 = 0.045 m and D4 = 0.06 m), in the second case, the bowl radius was varied from R1 to R4 in 0.001 m paces (R1 = 0.0047 m and R4 = 0.0077 m). During all the structural modifications, the compression ratio was locked and we have constant CR by handling the final volume of combustion chamber for each case. The validity of the results can be confirmed by comparison of simulated and experimental results for pressure courses, HRR, and accumulated heat as a function of CA under 1500, 2000, and 2500 rpm [17] and also engine power and AFR based on different engine speeds of 1500–4000 rpm [25,26] (Fig. 3). The highest discrepancy of pressure at peak pressure course under 1500 and 2500 rpm is less than 1%. Regarding HRR, it was observed that the trend was captured successfully. The mean error for AFR at different engine speed is 1.43%, whereas numerical power values match experimental values within 0.2–1% depending on engine speed. The operational conditions along with boundary condition and used sub-models of the engine simulation are provided in Tables 1 and 2. Additionally, the main

Fig. 2. Different modifications in bowl radius and outer bowl diameter from R1 and D1 to R4 and D4.

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physical and chemical properties of diesel fuel in current computational investigation are listed in Table 3. The combustion model is based on Coherent Flame Model that is coupled to the spray model and has the capacity of stratified combustion modeling and NO formation [18]. The extended combustion flame model is based on. ECFM-3Z [19] constitutes and recognizes three zones of unmixed air plus EGR (if any exists), the mixed air and fuel zone, and unmixed fuel. In this approach, flame propagation occurs from burned gas to unburned gas section and takes into account the three main combustion modes, i.e. Auto Ignition, Premixed Flame (oxidation), and Diffusion Flame. For a diesel spray, the fuel droplets have close proximity to each other

and the fuel droplets can be classified at unmixed fuel part of the computational cell. After evaporation of droplet fuels, a definite period is needed to transfer from the pure fuel area to mixed fuel and air region. In this situation, the mixing of fuel with air is modeled by initially placing the fuel into the ‘unmixed fuel’ zone of the ECFM-3Z model [19]. The transport equation from the unmixed to the mixed zone is solved and presented in detail in Ref. [19]. The turbulent flow of the spray and combustion is modeled by k–e. Additionally, the spray atomization and ligament disintegration is implemented by modified KH–RT simulation methodology [20]. Passing particles in the flow field are being deflected by eddy turbulence, hence additional turbulence effect on spray droplets

Fig. 3. The comparison of modeling and experimental data (a) accumulated heat @1500 rpm, (b) pressure @1500 rpm, (c) pressure @ 2500 rpm, (d) HRR @2000 rpm, and (e) engine power and AFR vs. various engine speeds.

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H. Taghavifar et al. / Energy Conversion and Management 85 (2014) 20–32 Table 1 Boundary condition and sub-models. Head temperature Piston temperature Cylinder temperature Spray breakup Combustion model Turbulence model Evaporation NO Fuel injection quantity Residual gas ratio Initial pressure Initial temperature Injection valve closed Exhaust valve open

550.15 K 575.15 K 475.15 K Modified KH-RT ECFM-3Z k–e Dukowicz Extended Zeldovich 31.3 (mg/cycle) 0.5 0.1 MPa 330 K 52°CA ATDC 110°CA ATDC

2.2. Parameter calculations

Table 2 Engine operational condition. Bore  stroke Displacement Compression ratio Swirl ratio @ IVC Rail pressure Nozzle geometry Number of nozzles Conrod Clearance Injection start timing Injection spray angle

not available, start of injection (SOI) has to be regulated from the instant where the raise pressure trace shows its first sharp reduction. All results were obtained at the limiting torque curve with limits set to the following parameters over different speed ranges: Air–fuel ratio, A/F:P19 for Ne (engine speed) between: 1250 and 1500 rpm. Maximum cylinder pressure, Pmax:6140 bar for Ne between: 1500 and 3000 rpm. Pre. Turbine temperature, Texhaust 6 800 °C for Ne: 3000– 4000 rpm. More details about experimental setup and information about data acquisition apparatus and procedure are mentioned in [25,26].

82.5  82 mm 438 cm3/cylinder 19.5:1 3 54–125.5 MPa (based on engine speed) 5  0.15 mm 4 130 mm 0.86 mm 3°CA ATDC 160°

In order to support the understanding of engine efficiency, two parameters are calculated for different designs of bowl geometry. Indicated torque is an index which shows the engine’s rotational ability and is calculated as the following equation:

Indicated Torque ¼

Indicated Thermal Efficiencyð%Þ ¼ ¼

IMEP  V d  N  60  100 2mf LHVf

ð1Þ W Q in ð2Þ

where Vd is the Volumetric displacement and N represents the engine speed. LHVf is the lower heating value of the diesel fuel while mf and IMEP are fuel mass and indicated mean effective pressure, respectively.

Table 3 Main properties of diesel fuel used in computations. Energy content (MJ/kg) Density (kg/m3) Stoichiometric A/F ratio (–) Molecular mass (kg/kmol) Dynamic viscosity (Pa s) Cetane no. LHV (MJ/kg) Auto ignition temperature (°C)

IMEP  V d 2p

43.09 828 14.7 179 0.00214 50–55 42.496 200–220

were modeled through stochastic dispersion method developed by Gosman and Ioanniodis [21]. Interaction between wall and spray droplets were modeled by walljet1 approach [22]. Evaporation model is Dukowicz with constant coefficients equal to 0.246 [23].

2.3. Homogeneity Factor The equivalence ratio is proved to fail at giving a suitable measurement of air/fuel mixture quality. Since the amplitude of equivalence ratio higher or lower than 1 denotes the worsening mixture quality (rich or lean fuel distribution), one needs to modify this quality as a parameter rearranged and calibrated in the range of 0–1. In this manner, Nandha and Abraham [27] developed a formula to quantify the heterogeneity of air/fuel mixture as follows:

rffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi PNcell 2 i¼0

½ðui u0 Þ Dmi  M

DOHðhÞ ¼

u0

ð3Þ

where: 2.1. Experimental description In order to set up flexible fuel injection and air charge, variablegeometry turbocharging (VGT) and high-pressure common rail (HPCR) systems are applied. The HPCR system assures fuel injection pressure to be independent of crankshaft speed, which accordingly better air/fuel mixture can be obtained. The experiments were carried out under limiting torque condition (LTC) of 1.8 L prototype direct injection diesel engine with a baseline build comprised of a fixed-geometry turbocharging (FGT) with a distributor electronic fuel injection system. HSDI (high-speed direct injection) Ford Diesel engine was equipped with a prototype Lucas CAV HPCR system, and an allied Signal VGT. The experimental results were obtained using an AVL heat release data acquisition system together with associated AVL Concerto software [24]. AVL 409 SMOKE (Bosch) and AVL AFR (Spindt) was used for measurement of smoke and Air/Fuel ratio. Since a needle lift transducer was

u0 ¼ M¼

PNcell i¼0

u i Di

M

Ncell X

Dmi

ð4Þ

ð5Þ

i¼0

Dmi and ui are mixture mass and equivalence ratio at i-th computational cell. Ncell represents the total number of cells while u0 is the overall equivalence ratio of the mixture and M is the total mass of the mixture. The proposed modeling of Degree of Heterogeneity (DOH) represents the standard deviation from the normalized equivalence ratio [27]. For diesel engine combustion, Mobasheri and Peng [28] presented Homogeneity Factor (HF) characterized such that increasing fuel amount can be interpreted as decrease of fuel amount in nearby cells as a relation of half standard deviation from equivalence ratio. In terms of average total equivalence

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ratio and stoichiometric air–fuel ratio (A/F)st, the deviation of fuel from average fuel amount is defined as the following:

ui ðA=FÞst þ ui ¼

Dmi 

u0 ðA=FÞst þ u0

Dmi

ðA=FÞst ðui  u0 Þ Dmi ½ðA=FÞst þ ui ½ðA=FÞst þ u0 

ð6Þ

Describing the total fuel amount in terms of u0:

u0 ðA=FÞst þ u0

M

Heterogeneity Factor (HeterF(h)) is the given as:

HeterFðhÞ ¼

i¼0

pffiffiffiffiffiffiffiffiffiffiffiffiffiffi2ffi ðui u0 Þ ðA=FÞst þui

Dmi

2u0 M

ð8Þ

Accordingly, Heterogeneity Factor is obtained:

HFðhÞ ¼ ð1  HeterFðhÞÞ

Engine structure was modified in two ways. Firstly, the bowl of piston was shifted outward and secondly the radius of the bowl was incremented evenly. The effects of these two parameters were taken into account as to explain the flow and combustion behavior within the cylinder.

3.1. Influence of engine geometry modification on equivalence ratio

ð7Þ

PNcell

3. Results and discussion

ð9Þ

Heterogeneity factor (HeterF) demonstrates the accumulation of fuel and non-uniformity in combustion chamber normalized in the range of 0–1. As Eq. (9) suggests the Heterogeneity factor is recognized as complement of Homogeneity Factor such that increasing in one of these parameters lead to decrease of the other one.

Fig. 4 shows the effect of outward movement of the bowl and the bowl radius on the equivalence ratio distribution at TDC, 10, and 20 ATDC. Fig. 4a implies that with 5 mm increment of the bowl section towards the lateral wall, better air–fuel mixture occurs up to D3 and further outward shifting of the bowl resulted in reverse mixing trend. With CA evolution and more fuel injection into the chamber, the fuel quotient increases and the highest equivalence ratio was observed at 20ATDC. At 10ATDC the over rich mixture zone is located in the bowl (up to D2) and increasing the bowl distance with central axis (D3 and D4) yields the transfer of over rich zone to the squish. At subsequent 20CA ATDC, the best air–fuel mixture distribution belongs to D2 and D3 with the majority of the cross-section covered with 2.7 equivalence ratio. The poor performance on the equivalence ratio is seen for D1 and D3 in the extreme situations where it reaches 6.5 in the swirl and top side of piston wall. Note that outward swirl chamber shift lets effective, uniform distribution of the fuel via provision of suitable volume for

Fig. 4. Equivalence ratio contours for (a) different outer wall of the bowl, (b) the bowl radius configurations at TDC, 10, and 20°CA ATDC.

H. Taghavifar et al. / Energy Conversion and Management 85 (2014) 20–32 Table 4 Air/fuel distribution characteristics according to various bowl designs. Bowl design

ER@780°CA

ER standard deviation

HF@780°CA

D1 D2 D3 D4 R1 R2 R3 R4

2.54123 2.356 2.226 2.139 2.191 2.257 2.368 2.4536

1.1274 1.0542 1 0.944 0.99 1.0167 1.06521 1.0817

0.5712 0.5946 0.6293 0.6786 0.619 0.5836 0.5696 0.5527

spray injection. Failure of D4 case in presenting of evenly fuel concentration chiefly stems from the impingement of injected fuel to the top pedestal of swirl chamber. Fig. 4b contrasts the equivalence ratio contours for different swirl chamber radii of 0.004, 0.005, 0.006, and 0.007 associated with D1, D2, D3, and D4 with CA evolution from TDC to 20°CA ATDC. At 10°CA ATDC, the over rich mixture zones appear in the bottom of swirl chamber. It is found that radius increment led to extension of high concentration (red region) across the bottom side of swirl wall. In the R4 case, fuel share of the mixture tends to increase. At R2 and R3 high concentration zones fade from 10 to 20°CA ATDC, however the mixture of 3.9 equivalence ratio expands throughout the cross section of combustion chamber. What is more, it can be deduced that the optimum structure of combustion chamber in terms of air–fuel mixture is concerned with D3 and R2;

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therein the uniform distribution of fuel and the accessible air was noticed. The fuel/air mixture characteristics for various bowl geometry are listed in Table 4. 3.2. Influence of engine geometry modification on temperature Fig. 5a illustrates temperature contours of modified combustion chamber (swirl chamber slide) with CA change at different values of outer bowl diameters, i.e. D1, D2, D3, and D4. At TDC position, higher temperature distribution was developed for D1 with 2000 K in high temperature zone. At 10°CA ATDC, high temperature zone across combustion chamber decreases with increase of outer bowl diameter whereas the contrary pattern was witnessed at 20°CA ATDC such that high concentration region of 2709 K covers the bowl and crevice section of the combustion chamber for D4 case. It can be appreciated from Fig. 5a that there is a huge gap of temperature intensity and extension between 10° and 20°CA ATDC for D4 (1369–2709 K), hence, steep temperature gradient is expected. Fig. 5b shows temperature distribution over combustion chamber with the same manner as previous Fig. 5a but rather considers various bowl radii. It can be seen that at TDC increasing radius brings about temperature reduction, although significant temperature drop happens with CA evolution from TDC to 10°CA ATDC. The highest temperature distribution is at 10°CA ATDC and the high temperature zone is concentrated in the bowl. The high concentration zone decreases gradually and moves towards squish area with 10°CA downward piston displacement.

Fig. 5. Temperature contours for (a) different outer wall of the bowl, (b) the bowl radius configurations at TDC, 10, and 20°CA ATDC.

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3.3. Influence of engine geometry modification on Homogeneity Factor and turbulent kinetic energy Fig. 6 shows the Homogeneity Factor and inlet mass flow of fuel in the combustion chamber for different designs of chamber geometry as a function of crank-angle. HF is defined as index of the quality of air/fuel mixture in the range of 0–1. Approaching zero denotes poor mixing process, whereas near 1 magnitudes represent better uniformity of the mixture. With that in mind, Fig. 6a shows better air/fuel homogeneity between 732 and 780°CA for D4 geometry. HF is increased for final state 0.5712, 0.5946, 0.6293, and 0.6786 with D1, D2, D3, and D4, respectively. The main reason for this lies in the more space for spray penetration and air entrainment without wall impingement. D2 chamber geometry shows well fuel distribution capability in spite of higher fuel injection rate in comparison with D1 geometry design. The results are in agreement with equivalence ratio distribution illustrated in Fig. 4a that with bowl distance increment, more air/fuel uniformity was observed. Fig. 6b shows the influence of bowl radius on air/fuel mixing and mass flow as a function of crank-angle. Higher HF is

seen with R1 configuration and with the bowl’s radius increment from R1 to R4 the homogeneity is reduced from 0.619 to 0.5527. This is because of stronger squish and vortex formation in the flow field of the cylinder which mixes the fuel with air more efficiently. Fig. 7a shows TKE concentration across cross section area of the combustion chamber with CA variation after TDC onwards at different values of outer bowl wall. The high speed of injection from the nozzle tip has the maximum amount of TKE equals to 307 m2/ s2 at TDC whilst injection of the spray. TKE reaches the maximum level at 10°CA ATDC whereby the bowl section was also affected for D2, D3, and D4. TKE fades at 20°CA ATDC as spray injection was ceased. The energy of spray injection defines the TKE values, thus decreasing TKE with increase of outer wall diameter can be explained since closer bowl wall makes more spray-wall interaction. Taken Fig. 7a together with Fig. 4a, it can be noticed that decreasing the bowl distance (from D4 to D1) causes the reduced distance of injector-combustion chamber inner wall. This makes more spray-wall impingement and thus higher TKE distribution in close bowl placement. The close bowl distance leaves positive

Fig. 6. Variation of HF and inlet mass flow rate as a function of crank-angle at different bowl configurations (a) bowl outward wall (D1–D4), (b) bowl radius (R1–R4).

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Fig. 7. TKE contours for (a) different outer wall of the bowl, (b) the bowl radius configurations at TDC, 10, and 20°CA ATDC.

influence on the flow turbulence and mixture quality, nonetheless, due to spray-wall interaction, a portion of liquid fuel sediments on the wall and losses the chance of vaporization which deteriorates the combustion process (better equivalence ratio is seen when the bowl is getting far away: Fig. 4a). Fig. 7b demonstrates TKE variation with CA at different bowl radii. The observed trend indicates that high TKE was intensified at TDC near the nozzle exit orifice and with 10°CA piston expansion the intensity decreases but high TKE was more extended across the cylinder. The high TKE zone remains unaffected with the bowl radius increase at TDC and 20°CA ATDC whereas increasing trend of TKE with radius increase was noted up to R3. High TKE section from nozzle tip was shortened and the cone angle was decreased at R4. 3.4. Influence of engine geometry modification on combustion, emission, and engine performance The pressure and HRR curves of different combustion chamber designs are illustrated in Fig. 8. According to Fig. 8a, from D1 to D4 configuration, the pressure curves are shifted to the right while the 2-stage peak pressures are forming more clearly. Due to existence of more requisite time for premixed air/fuel preparation with D4, higher peak pressure and HRR were obtained. However, the pressure generation because of combustion was further postponed from TDC. That is to say, lower combustion work in a short interval was exerted on piston, hence lower torque and engine efficiency

should be expected accordingly. As for HRR histories with crankangle, the highest peak is of D4, although due to coincidence of majority of expansion stroke and its heat lessening effect with combustion, a sharp decrease of heat release is noticeable. Increasing the bowl distance from D1 to D4 provides suitable space for efficacious air/fuel mixing which gives way to increasing stoichiometric combustion as well as higher indicated torque and mechanical efficiency (refer to Table 5). On the other hand, with increasing the bowl distance after D3, the initiation of combustion was delayed and its work production was applied during shorter period of expansion stroke. It can be concluded that there is a competition between two factors of the combustion heat magnitude because of better premixed-control combustion and the timing of work impact with increasing the bowl distance from D1 to D4. It seems that at D4 configuration, the impact of delayed combustion initiation outweighs the constructive effect of better air/fuel mixture preparation. Fig. 8b shows that no perceptible expansion pressure was obtained with the bowl size manipulation, although the overall pressure tends to decrease along with heat release delay to be increased. It is obvious that smaller bowl size induces higher squish and swirl/tumble flows which subsequently yields higher pressure. As a result of better squish and pressure a uniform flow was created within cylinder and more mixing time, a large premixed mixture can be achieved, which in turn increases Homogeneity Factor.

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Fig. 8. Pressure trace and HRR as a function of CA for (a) different outer wall of the bowl, (b) the bowl radius configurations.

Table 5 The engine efficiency parameters with different structure specifications. Quantities

D1

D2

D3

D4

R1

R2

R3

R4

ISFC (g/kW h) IMEP (MPa) ITE (%) Indicated torque (N.m)

441 1.12 48 15.25

374 1.21 51 16.41

174 1.56 58 21.27

206 1.49 56 20.3

350 1.24 52 16.87

374 1.21 51 16.41

380 1.2 51 16.3

358 1.23 52 16.71

As shown in Fig. 9 there are more vortex flow motion formation in the velocity field of R1 than R4, implying that with R1 design, better swirl/ tumble can be obtained to enhance the homogeneity of mixture. Moreover, the HRR of R4 is tangibly higher than other designs due to more oxygen content and air accessibility in bigger bowl geometry. Due to more time for air/fuel mixing and spray penetration, the HRR peak is augmented from 29.07 J at 724°CA to 76.13 J at 726°CA from R1 to R4. Although R4 has faster ascending with crank-angle, the heat generation duration is lower and demonstrates faster descending trend owing to the occurrence of the late combustion initiation accompanied with more bowl sur-

face area for wall heat transfer. Comparing Fig. 8a and b reveals that the outward movement of the bowl leaves stronger impact than that of the bowl radius increment in terms of combustion phenomenon (drastic changes of pressure and HRR occurs in the former). Fig. 8a shows that by the outward bowl movement from D1 to D4, compression and combustion pressure decreases while the expansion pressure increases substantially. On the other hand, two peak pressures were observed in this state and the peak pressure shifts towards late combustion period. With regard to HRR trace, increasing outer wall of the bowl resulted in higher PMC peak and higher ignition delay. More HRR period and the higher

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Fig. 9. The in-cylinder fluid flow pattern and swirl motion illustration for (a) R1, (b) R4.

effective pressure at TDC make better combustion output for D1. A single heat release peak was observed for D1 and D4 that implies PMC combustion is dominant and slight diffusion controlled combustion exists in these states. The results also indicate that with the bowl outward movement, the combustion phasing was retarded and starts at 7°CA ATDC. It can be attributed to the elongated time of the spray contact with the distant bowl wall for D4 case as well as lower in-cylinder temperature at initial CA period. Fig. 8b shows that there is negligible difference of expansion pressure while the main difference relies on the pressure course during compression stroke and combustion phasing. As the bowl radius of combustion chamber increases, the in-cylinder pressure decreases since more volume was provided for spray combustion and the produced gases of fuel burning in the bowl. Two stages HRR process associated with two peaks are evident in the graph for all cases except D1, which shows mild, prolonged session of heat generation over CA. As shown in Fig. 8b, increase of the bowl radius caused more ignition delay and higher PMC peak for HRR. To sum up, HRR courses are explained as follows for different configurations: (1) with increasing D, more fuel injection volume was created as well as lower spray-wall interaction leading to longer time for ignition initiation. Longer ignition delay and more HF bring about higher HRR peak for D4 case, however mild decreasing trend for D1 stems from lower surface area for heat transfer. (2) Higher peak of HRR was observed for R4 case since there is more space for spray injection and lower squish flow, which retards the ignition of air/fuel mixture. Due to more oxygen content available for higher bowl radius and more air/fuel mixture duration, acute burning rate and chemical reaction energy was resulted. Meanwhile due to late combustion time and coincidence

of combustion with expansion stroke, the heat release is diminished drastically for higher R. Fig. 10 represents combustion initiation postponement as bowl wall diameter and the bowl radius was increased, furthermore it was shown that outer wall of the bowl’s outward movement has more influence in ignition delay increase. Fig. 11 shows the temperature and NOx emission mass fraction with outer wall movement of the bowl as a function of CA. A direct proportional correspondence of temperature and NOx trend is seen as expected. The lowest NOx concentration and temperature was observed for

Fig. 10. The comparison of ignition delay for different engine configurations.

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Fig. 11. Temperature trace and NOx concentration as a function of CA for (a) different outer wall of the bowl, (b) the bowl radius configurations.

D4 until 13°CA ATDC, thereafter a sudden rise of temperature and NOx happens. With the outward bowl displacement from D1 to D4, the maximum Temperature and NOx concentration increases in a way that 38.2% and 82.6% increase of temperature and NOx from D1 to D4 takes effect, respectively. The highest temperature is of D1 until 10°CA ATDC, while the contrary trend was reported at later CA periods and D4 prevails in this sense. Comparing Fig. 11a and b determines that there exists lower range of temperature and NOx emission when the bowl radius modification was applied. Fig. 11b shows the highest NOx emission was resulted from R1, although high temperature in this case is only up to 10°CA ATDC and afterwards the maximum temperature is lower than R4. Fig. 12 indicates the soot trend as a function of CA for different geometrical cases of piston shape. As Fig. 12(a) presents with increasing the outer wall diameter (D), the soot concentration has been decreased substantially. Having the HF quantity in mind, the reason for soot decrement can be explained clearly, as it defines the quality of mixture uniformity. Increasing bowl radius divulged rather different impact on soot emission development. According to Fig. 12(b), R1 and R4 demonstrated the lowest soot mass fraction within combustion chamber. R1 dominates in low

soot fraction because of better squish and vortex flow formation, which results in better air/fuel mixture and low equivalence ratio, while R4 owes its low soot content to negligible flame quenching as a result of spray jet collision to inner bowl wall. A close inverse correlation can be deduced between the amount of Soot mass fraction and engine performance indices provided in Table 5. The engine performance was also mentioned in Table 5. According to Table 5, IMEP, indicated torque, and indicated thermal efficiency increases up to D3 and decreases slightly for D4. Moreover, the specific fuel consumption was decreased from D1 to D3, and decrease of ISFC is seen at D4. All of these data are indicator of superior engine performance in terms of work delivery and fuel consumption for D3. This shows that in the competition between two conflicting parameters in engine performance boost (i.e. mixture quality and combustion timing) the late combustion initiation and lower heat to work transformation span outperforms, therefore inflicts D4 with slightly low engine performance. The engine performance indices show insignificant changes with the bowl radius alteration when compared with that of bowl movement. The engine specifications of IMEP, indicated torque, and ITE are decreased and ISFC increased up to R3 and slightly that shows poor performance with R3. The better performance of engine in terms of

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content and air accessibility with higher bowl radius. This was indeed attributed to lower squish and vortex flow motion together with more heat loss from piston wall due to higher surface area associated with R4 design. (5) Comparing two factors of bowl design, change in bowl displacement demonstrates higher flow dynamic and engine performance change. To sum up, it is highly recommended that engineers and manufacturers focus on this element in order to acquire best efficiency. (6) Since lower D and R is conducive to the risk of spray-wall collision, it is recommended that higher swirl ratio applied at lower R and D to deflect direct spray injection and thus acquire better air/fuel mixture. Moreover, at bigger bowl radius lower vortex and air/fuel motion exists that can potentially lead to fuel accumulation on the inner wall. Therefore higher swirl ratio can remain positive effect on overall mixture uniformity for this particular case. (7) As mentioned lower R and D configurations makes higher squish flow and hence higher air compression and mixture temperature. Start of injection schemes have to be such regulated as to avoid spray injection at higher combustion chamber temperature due to occurrence of knock phenomenon. The authors suggest that at lower bowl volume structures (D1 and R1) SOI should be postponed before top dead center. References

Fig. 12. Soot mass fraction traces as a function of CA for (a) different outer wall of the bowl, (b) the bowl radius configurations.

combustion and operational efficiency can be originated from better recirculation of the flow. 4. Concluding remarks 3-D CFD study is presented to analyze the effect of modification in piston head shape at constant compression ratio on the fluid flow, combustion, emission, and engine operation. The results were discussed elaborately and the gists of conclusion are as follows: (1) It was found that increasing outward bowl movement (D1– D4 configuration) provides better uniformity in air/fuel mixture either qualitatively (equivalence ratio) or quantitatively (Homogeneity Factor) which results in higher peak pressure and HRR. However, combustion initiation was postponed to late expansion period with lower work delivery that acts adversely on engine performance and combustion heat generation. (2) TKE amplitude tends to decrease from D1 to D4 due to decrease in spray-wall collision, affecting spray-air distribution unfavorably, although benefits the combustion process since lesser fuel amount was deposited on the bowl wall and evaporation was enhanced dramatically. (3) It was reported that at 10°CA ATDC, increasing bowl displacement is conducive to temperature and NOx emission reduction, whereas at 20°CA ATDC there noticed a sharp ascending gradient of temperature increase from D1 to D4 configuration. (4) Taken together, increasing the bowl radius proved affecting the engine performance adversely, in spite of more oxygen

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