Effects of the Stroke/Bore Ratio on the Performance Parameters of a Dual-Spark-Ignition (DSI) Engine † PDF

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Energy & Fuels 2009, 23, 1825–1831 1825 Effects of the Stroke/Bore Ratio on the Performance Parameters of a Dual-Spark-Ignition (DSI) Engine† I˙smail Altin,*,‡ I˙smet Sezer,§ and Atilla Bilgin| Trabzon Vocational School, Karadeniz Technical UniVersity, Trabzon 61300, Turkey, Bes¸ikdu¨zu¨ Vocatio...

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Effects of the Stroke/Bore Ratio on the Performance Parameters of a Dual-Spark-Ignition (DSI) Engine † I. Sezer Energy & Fuels

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Energy & Fuels 2009, 23, 1825–1831

1825

Effects of the Stroke/Bore Ratio on the Performance Parameters of a Dual-Spark-Ignition (DSI) Engine† I˙smail Altin,*,‡ I˙smet Sezer,§ and Atilla Bilgin| Trabzon Vocational School, Karadeniz Technical UniVersity, Trabzon 61300, Turkey, Bes¸ikdu¨zu¨ Vocational School, Karadeniz Technical UniVersity, Trabzon 61800, Turkey, and Department of Mechanical Engineering, Karadeniz Technical UniVersity, Trabzon 61080, Turkey ReceiVed May 10, 2008. ReVised Manuscript ReceiVed July 3, 2008

In this study, the effects of the stroke/bore (S/B) ratio on the performance parameters of an engine having dual-spark ignition (DSI) were investigated theoretically by using a quasi-dimensional thermodynamic cycle model. A range of S/B ratios for various spark-plug locations were examined in the scope of the study. The engine performance parameters, i.e., engine power, indicated mean effective pressure, specific fuel consumption, and thermal efficiency, were computed for various conditions. The obtained results for DSI operation have been compared to those of a single-spark-ignition (SSI) condition. Such comparisons showed that increases in the S/B ratio cause improvements in engine performance parameters. The dual-plug operation also improves the overall engine performance at different percentages for the selected S/B ratios.

1. Introduction The geometric dimensions and design parameters have considerable effects on internal combustion engine (ICE) performance. Therefore, it is vital that the efforts intended to optimize these parameters achieve the best engine performance. The stroke/bore (S/B) ratio is one of the most important geometric parameters for modern spark-ignition (SI) engines1 because it determines the overall dimensions of the engine for a given displacement.2 However, there are only a few studies performed to investigate S/B ratio effects on engine performance and exhaust emissions for two- and four-stroke engines.2-6 Usually, in these studies, the S/B ratio changes between 0.7 and 1.4 as in modern engines1,3 and single-spark-ignition (SSI) engines having centrally located plug were commonly used.2,5,6 In general, a longer stroke leads to higher thermal efficiency through faster burning (reduction in combustion duration) and lowering the overall chamber heat loss.2 A shorter stroke decreases engine friction, most noticeably at higher engine †

From the Conference on Fuels and Combustion in Engines. * To whom correspondence should be addressed. Telephone: +90-462377-71-21. Fax: +90-462-248-22-34. E-mail: [email protected]. ‡ Trabzon Vocational School. § Bes ¸ikdu¨zu¨ Vocational School. | Department of Mechanical Engineering. (1) Chon, D. M.; Heywood, J. B. Performance scaling of spark-ignition engines: Correlation and historical analysis of production engine data. SAE Paper 2000-01-0565, 2000; pp 1-12. (2) Filipi, Z. S.; Assains, D. N. The effect of the stroke-to-bore ratio on combustion, heat transfer and efficiency of a homogeneous charge spark ignition engine of given displacement. Int. J. Engine Res. 2000, 1 (2), 191– 208. (3) Siewert, R. M. Engine combustion at large bore-to-stroke ratios. SAE Paper 780968, 1978; pp 3637-3651. (4) Alsterfak, M.; Filipi, Z. S.; Assanis, D. N. The potential of the variable stroke spark-ignition engine. SAE Paper 970067, 1997; pp 1-10. (5) Yamin, J. A. A.; Dado, M. H. Performance simulation of a fourstroke engine with variable stroke-length and compression ratio. Appl. Energy 2004, 77 (4), 447–463. (6) Thornhill, D.; Douglas, R.; Kenny, R.; Fitzsimons, B. An experimental investigation into the effect of bore/stroke ratio on a simple twostroke cycle engine. SAE Paper 1999-01-3342, 1999; pp 1-14.

speed.7 It also increases the maximum operating speed, maximum power, indicated mean effective pressure (imep), and also blow-by of the engine.1,3 In addition, the larger bore will provide more room for poppet valves in four-stroke engines.6 Hence, increasing the number of valves per cylinder for a given cylinder bore improves engine breathing.1 On the other hand, the variation of the S/B ratio has impacts on exhaust emissions. The CO and HC emissions increase with a decreasing S/B ratio, while NOx emissions tended to decrease6 because of increasing crevice volume8 and decreasing temperature.6 Furthermore, turbulence characteristics, detailed heat-transfer analysis, and burned gas temperatures have been investigated in the studies on the effects of the S/B ratio.2,4,9 Effects of the S/B ratio on SI engine performance parameters and exhaust emissions were investigated at various spark-plug configurations.4 It is concluded that centrally located single spark plug among the various sparkplug configurations could produce better results,4 but it would be impossible because of design limitations.10 The performance parameters and exhaust emissions of a variable stroke SI engine were investigated with centrally located single-plug configuration.3 The use of multiple spark plugs is suggested to improve the engine performance in a variable stroke engine.3 It has been clearly understood in the literature review that investigations on the S/B ratio are limited. The centrally located single plug is suggested to have an optimal configuration, but it cannot be applied commonly because of some designing limitations. In that case, it is considered to use multiple spark-plug configurations as a suitable solution. For these reasons, this study is (7) Patton, K. J.; Nitschke, R. G.; Heywood, J. B. Development and evaluation of a friction model for spark-ignition engines. SAE Paper 890836, 1989; pp 1-21. (8) Trinker, F. H.; Cheng, J.; Davis, G. C. A feedgas HC emission model for SI engines including partial burn effects. SAE Paper 932705, 1993; pp 1-17. (9) Brognakke, C.; Arpaci, V. S. A model for the instantaneous heat transfer and turbulence in a spark ignition engine. SAE Paper 800287, 1980; pp 1-15. (10) Martins, J. J.; Pereira, E. C. O.; Rodrigues, E. Design of a fuel efficient uniflow two-stroke semi-direct injection engine. SAE Paper 970367, 1997; pp 85-93.

10.1021/ef800332m CCC: $40.75  2009 American Chemical Society Published on Web 09/10/2008

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Altin et al.

devoted to investigate the effect of the S/B ratio on the performance parameters for dual-spark-ignition (DSI) engine and SSI engine conditions. The cylinder mid-radius point was selected as the optimum location for dual (twin or identical) spark-plug operation.11 The results of the identical plug were compared to those of single-plug configuration. Such comparisons showed that a mid-radius dual spark plug generally gives slightly higher engine performance parameters than the centrally located single plug for selected S/B ratios. On the other hand, increasing the S/B ratio improves the engine performance in a decreasing manner.

p˙ ) T˙b )

-hgAb(Tb - Tw) Vb ∂ln Vb + p˙ + ωmCpbxb Cpb ∂ln Tb

1 ˙ VC V+ (1) m ω Tw Vu ∂ln Vu Tw hg Vb ∂ln Vb B) 1A + 1A (2) ωm Cpb ∂ln Tb Tb b Cpu ∂ln Tu Tu u

[

)

( )

( )] [ ]

(xb - xb2)CL ∂ln Vb hb - hu C ) -(Vb - Vu)x˙b - Vb x˙b (3) ∂ln Tb CpbTb ω

[ ( ) [ ( )

Vb2 ∂ln Vb D ) xb CpbTb ∂ln Tb

2

Vb ∂ln Vb + p ∂ln p

Vu2 ∂ln Vu E ) (1 - xb) CpuTu ∂ln Tu

2

[

T˙u )

-hgAu(Tu - Tw) Vu ∂ln Vu + p˙ ωmCpu(1 - xb) Cpu ∂ln Tu ˙ ) pV˙ W

2.1. Thermodynamic Model. In this study, a quasidimensional thermodynamic cycle simulation with a two-zone combustion model, mainly based on Ferguson’s work12 has been used. Equations of the original model have been reconstructed to investigate the effects of dual spark on SI engine performance.11 This model has been developed to predict the cylinder pressure, temperatures of the burned and unburned gases, work done, heat loss, and the enthalpy of exhaust gas. The fuel burning rate has been determined by the flame propagation approach, which will be expressed below. The two-zone combustion model assumes that the cylinder volume is divided into burned and unburned zones by an infinitesimally thin flame front (namely, the volume of the reaction zone, which is neglected here).13 Each zone is assumed to be in thermodynamic equilibrium, with both zones having the same uniform pressure at any instant of time (here crank angle).14 The properties of the fuel-air-residual gas mixture during the compression and the equilibrium properties of the burned gas mixture during the combustion and expansion were determined by means of FORTRAN subroutines FARG and ECP, respectively, developed by Ferguson.12 The constants used in the model, such as heat-transfer coefficients (hg), the cylinder wall temperature (Tw), and the blow-by constant (CL), were taken as 500 W m-2 K-1, 400 K, and 0.8 s-1, respectively.12 The following governing equations of the model are based on the first law of thermodynamics and derived as in the literature.11

(

(6)

]

Vu ∂ln Vu + p ∂ln p

(4)

]

(5)

(11) Altin, ˙I. A quasi-dimensional two-zone thermodynamic cycle model for spark ignition engines having dual ignition system. M.S. Thesis, Karadeniz Technical University, Trabzon, Turkey, 2004 (in Turkish). (12) Ferguson, C. R. Internal Combustion Engines Applied Thermosciences; John Wiley and Sons, Inc.: New York, 1986. (13) Bayraktar, H.; Durgun, O. Mathematical modeling of spark-ignition engine cycles. Energy Sources 2003, 25 (5), 651–666. (14) Chan, S. H.; Zhu, J. Modelling of engine in-cylinder thermodynamics under high values of ignition retard. Int. J. Therm. Sci. 2001, 40 (1), 94–103.

]

CL hu - hb x˙ - (xb - xb2) (7) xbCpb b ω

2. Engine Cycle Simulation

A)

A+B+C D+E

˙L) Q

hg [A (T - Tw) + Au(Tu - Tw)] ω b b

(8) (9) (10)

CLm [(1 - xb2)hu + xb2hb] (11) ω 2.2. Combustion Model. The combustion model used here consists of the geometrical and burn-rate sub models given below. 2.2.1. Geometric Model. Experimental observations have shown that the actual flame shape is approximately spherical for a wide range of operating conditions.15 For this reason, the model is based on the premise of spherical flame propagation for simplicity. The developed geometric model was used to determine flame front area, heat-transfer surface area, and burned gas volume as a function of the flame radius, crank angle, and combustion chamber design. The computational method employed is based on the technique presented by Blizzard and Keck16 and Bilgin17 (for the dual-spark case) for a simple disk chamber, as in Figure 1a for single-plug configuration and Figure 1b for dual-plug configuration. Equations were derived to define the flame front area, enflamed volume, and wetted combustion chamber surface area as a function of the axial location. These equations were integrated to obtained the volume and areas required. The details of the geometric model relating the single-plug and dual-plug application can be found in the literature.16,17 2.2.2. Burn Rate Model. The combustion process was modeled considering the details of the turbulent flame propagation process. The method used here was postulated originally by Blizzard and Keck,16 and extended later by Keck and coworkers.18 The following equations have been used for the determination of the burn rate in the combustion simulation.16,18 ˙L) H

dme ) FuAfUe dθ

(12)

dmb ) FuAfSL + (me - mb)/τb dθ

(13)

τb ) lT/SL

(14)

Here, Af is the area of the flame front, and τb is the characteristic burning time of eddy at a size of lT. The flame front area Af has (15) Lucas, G. G.; Brunt, M. F. J. The effect of combustion chamber shape on the rate of combustion in a spark ignition engine. SAE Paper 820165, 1982; pp 714-729. (16) Blizard, N. C.; Keck, J. C. Experimental and theoretical investigation of turbulent burning model for internal combustion engines. SAE Paper 740191, 1974; pp 846-864. (17) Bilgin, A. Geometric features of the flame propagation process for an SI engine having dual-ignition system. Int. J. Energy Res. 2002, 26 (11), 987–1000. (18) Keck, J. C. Turbulent flame structure and speed in spark-ignition engines. Nineteenth International Symposium on Combustion/The Combustion Institute, 1982; pp 1451-1466.

Energy & Fuels, Vol. 23, 2009 1827

Stroke/Bore Ratio on SI Engine Performance

been computed instantaneously depending upon the enflamed volume Vf using a geometrical submodel. The equations given below were also used to determine turbulent entrainment speed Ue, turbulent speed UT, and characteristic length scale of the turbulent flame lT. (15)

Ue ) UT + SL ji UT ) 0.08U

() Fi Fe

1/2

j i ) ηv(Apc/Aiv) Sn U 30 Fe 3/4 lT ) 0.8Liv Fi

()

(16) (17) (18)

j i is the speed of gases entering the cylinder during where U induction, Apt is the area of the piston crown, Aiv is the maximum opening area of the intake valve, and Liv is the maximum intake valve lift. Laminar flame speed SL is calculated by the equations developed by Gu¨lder.19 SL(φ, T, p) ) SL,0[Tu/T0]R[p/p0]β(1 - ψf)

(19)

SL,0 is the laminar flame speed at standard conditions of p0 and T0 and calculated as SL,0(φ) ) ZWφη exp[-ξ(φ - 1.075)2]

(20)

The values of R, β, Z, W, η, and ξ are given in ref 21, and ψ is selected as 2.5 for 0 e f e 0.3. 2.3. Solution Procedure. A simulation program has been developed for the presented quasi-dimensional SI engine cycle

Table 1. Specifications of the Engines specification r

D (mm) S (mm) Lcr (mm) div (mm) Liv (mm)

rs

experimental 7 0.1 engine I21 experimental 5 0.3 engine II16

76.2

111.125

220

30

4.2

63.5

76.2

127

25

4.8

model. Input parameters of the program are engine speed n, equivalence ratio φ, distance of ignition point from the cylinder axis rs, spark-plug number, spark advance angle θs, properties of fuel, ambient pressure, and temperature. After determining the intake conditions, the thermodynamic state of the cylinder charge was predicted by solving the arranged first-order ordinary differential equations for each process by taking the crank angle increment as 1°. The DVERK subroutine, which uses the Runge-Kutta methods, was used to integrate the governing equations. At the start of combustion [at θs before top dead center (TDC)], the initial value of the burned gas temperature was determined as the adiabatic flame temperature.13 The initial value for the mass fraction burned was predicted from the cosine burn rate formula. Laminar burning was assumed during the ignition delay period. After the delay period was terminated, combustion was computed as a fully developed turbulent flame process. Throughout the simulation, the thermodynamic state of cylinder content was determined. Finally, after the computation of cycle was completed, the results were corrected according to the error analysis as in the following equations and computations were repeated until the desired convergence was achieved.12 ε1 ) 1 - Vm/V ε2 ) 1 +

W ∆(mu) + QL + HL

(21) (22)

2.4. Computation of the Engine Performance Parameters. The governing equations of the mean indicated pressure (pmi), indicated thermal efficiency (ηi), indicated specific fuel consumption (bi), and indicated power (Ni) were specified as in the literature.11,20 pmi )

Wi Vd

pmiVdzn k60 pmiRTo ηi ) FsφΗupoηv Ni )

bi )

3600 Huηi

(23) (24) (25) (26)

In eq 24, coefficient k is dependent upon the stroke number of the engine cycle and is taken as 1 for two-stroke engines and 2 for four-stroke engines. 2.5. Model Validation. In this section, predicted values were compared to the experimental data given in the literature.16,21 The specifications of engines were used for comparisons given in Table 1. Here, cylinder pressure and mass fraction burned

Figure 1. Flame front geometry for (a) single-plug and (b) dual-plug configurations.

¨ . Correlations of laminar combustion data for alternative (19) Gu¨lder, O SI engine fuels. SAE Paper 841000, 1984. (20) Bayraktar, H. Theoretical investigation of the effects of gasolineethanol blends on spark-ignition engine combustion and performance. Ph.D. Thesis, Karadeniz Technical University, Trabzon, Turkey, 1997 (in Turkish). (21) Rakopoulos, C. D. Evaluation of a spark ignition engine cycle using first and second law analysis techniques. Energy ConVers. Manage. 1993, 34 (12), 1299–1314.

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Altin et al.

Figure 2. Comparison of predicted cylinder pressure values with experimental data.

Figure 3. Comparison of predicted mass fraction burned values with experimental data. Table 2. Engine Specifications2 S/B ratio stroke (S) (mm) bore (B) (mm) connecting rod length (Lcr) (mm) compression ratio (rc) maximum valve lift (Liv) (mm) displacement volume (Vd) (cm3)

0.7 62.8 90 126 9 1.007 400

1.0 79.5 80 159 9 8.95 400

1.3 95.5 73 191 9 8.17 400

were selected as comparison parameters. Figure 2 shows predicted and experimental cylinder pressure values versus the crank angle. The predicted values of the mass fraction burned are compared to experimental values in Figure 3. The predicted values are in excellent agreement with the experimental data in both figures; therefore, the presented model has an enough level of confidence for parametric investigation.

3.1.1. Wall Areas in Contact with Burned Gases. As known, heat transfer during combustion has a very significant effect on cycle efficiency. The areas in contact with burned or unburned gas play an important role in the heat-transfer process. Wall areas in touch with burned gases should be evaluated to assess the importance of real geometric effects on the heat-transfer process in SI engines because the gas temperature in the burned zone is much higher than in the unburned zone.2 Figure 5 shows a comparison of wall areas contacted by burned gases for different S/B ratios at single- and dual-plug configurations (i.e., cylinder axis and mid-radius). For centrally located single plug, the normalized total wetted area increases with a decreasing S/B ratio and it is the largest for the short-stroke engine in Figure 5a. The larger bore expands the cylinder inside area, which naturally increases the wetted area by burnt gas. On the other hand, the wetted area for dual plug is lower than that of single plug, at mid-radius plug location in all S/B ratios in Figure 5b. This can be attributed to the faster burning when using dual plug; therefore, the combustion process ends before the piston moves away from TDC and the lower wetted area exists. It is evident that dual plug in various S/B ratio configurations has an advantage in terms of heat loss because of the less normalized total wetted area. 3.1.2. Mass Fractio...