Indian Journal of Engineering & Materials Sciences Vol. 13, August 2006, pp. 307-321

Optimization of inlet valve closure timing and clearance volume of a SI engine for better performance at part loads⎯A numerical and experimental approach

J M Mallikarjuna & V Ganesan Internal Combustion Engines Laboratory, Department of Mechanical Engineering

Indian Institute of Technology Madras, Chennai 600 036, India

Received 12 October 2004; accepted 5 June 2006

In this paper, a computer simulation and experimental investigations on a single cylinder, four-stroke, spark ignited engine are carried out to optimize inlet valve closure timing and clearance volume for better part-load performance. The simulation procedure involves thermodynamic and global modeling techniques. Many sub-models have been used for predicting heat transfer, friction and gas exchange processes. Unburned hydrocarbons, carbon monoxide and nitric oxide emissions are also predicted.

Experiments have been conducted on a single cylinder, air-cooled, four-stroke, spark-ignited engine. In this work, by varying inlet valve closure timing (IVCT) and clearance volume, geometric expansion ratio (GER) of engine alone is varied, while effective compression ratio (ECR) is kept constant, thereby GER/ECR ratio is altered. For modified engine, GER/ECR ratio is varied from 1.25 to 2. Experiments have been conducted for two effective compression ratios, viz., 7 and 8 at a speed of 1200 rpm. Performance and exhaust emission characteristics have been measured at different loads and GER/ECR ratios. The predicted performance and emission characteristics are compared with measured values and the agreement between the two is found to be good. It is observed that, for modified engine, considerable improvements have been found in the reduction of pumping losses (about 25.8 to 56%), increase in the brake thermal efficiency (about 13.6 to 25%), and reduction in unburned hydrocarbon emissions (about 18.5 to 58%). Finally, for modified engine, it is found that GER/ECR ratio of 1.5 gives the best performance compared to standard engine for both compression ratios.

IPC Code: F01L1/34

Previous investigations1-7 have shown that, one of the best ways to increase the thermal efficiency of a spark-ignition (SI) engine is to increase the cycle efficiency rather than increasing the combustion or mechanical efficiency. In this work, theoretical and experimental investigations have been carried out to study the improvement in the performance of a SI engine by increasing cycle efficiency using the concept of variable valve timing and clearance volume, which reduces the throttling losses to a large extent at part loads4-7. In a conventional spark-ignition engine, geometric expansion ratio (GER) is equal to geometric compression ratio (GCR) (both are based on volumes) and therefore, GER/ECR ratio is equal to one1, 6. Also, in these engines, the power output is controlled by applying throttle, which results in large pumping losses causing poor part-load efficiencies. It is possible to increase cycle efficiency of conventional engines by increasing either the GCR or GER or both. But, this is restricted by knocking characteristics of the fuel. However, if a SI engine is made to operate

on Otto-Atkinson cycle1,3,6 (in which GER/ECR ratio is more than one), the expansion can be extended until the pressure at the end of expansion is atmospheric allowing the gases to do more work on piston. But, practically there are limitations to increase the expansion to a very high value because of losses3. Therefore, in practice expansion should be stopped even when the pressure at the end of expansion is slightly higher than the atmospheric. The actual cycle of modified engine used for the purpose of theoretical analysis is as shown in Fig. 11. Practically, when inlet valve closure timing (IVCT) is delayed more than the conventional engine timing, the effective compression will be reduced. Also, volumetric efficiency (hence power output) of the engine reduces because, during early stages of compression stroke, upward motion of piston drives out the cylinder-charge into the inlet manifold through the inlet valve. In order to overcome the reduction of effective compression ratio (thereby brake thermal efficiency), clearance volume can be suitably reduced. This in turn will increase GER and results in making

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GER/ECR ratio of the engine more than one. Therefore, by selecting the appropriate values of IVCT and clearance volume, engine can be made run with desired GER/ECR ratio. This can be used to facilitate the load control of the engine without throttling and without reducing the thermal efficiency. However, the given combination of IVCT and clearance volume (at given GER/ECR ratio) of the modified engine at wide-open throttle produces less power output than the conventional engine. Further, reduction of power output can be achieved by applying a little throttle operation which is much less when compared to the conventional engine.

Mathematical Models and Theoretical Study Clearance volume and GER/ECR ratios In the present work, effective compression ratio (ECR) of the modified engine is not calculated on the basis of the ratio of the volumes at IVCT and TDC1. However, it is determined on the basis of the motored volumetric efficiency of modified engine at a given IVCT and clearance volume, and the motored volumetric efficiency of the standard engine. For the modified engine, first, GER/ECR ratio is chosen. Then, geometric expansion ratio is calculated as,

GERGER = ECRECR⎡ ⎤⎢ ⎥⎣ ⎦

… (1)

And GER is also given by,

disp_std

c_mod

GER = 1 + VV

⎛ ⎞⎜ ⎟⎜ ⎟⎝ ⎠

… (2)

In this work, effective compression ratio (ECR) of the modified engine is based on effective swept volume (ESV) and the clearance volume of the modified engine. Therefore,

c_mod

ESVECR = 1 + V

… (3)

and, effective swept volume of the modified engine is expressed as the product of the swept volume of standard engine and the ratio of motored volumetric efficiency of the modified engine to the motored volumetric efficiency of the standard engine. Then,

mod_motdisp _std

std_mot

ηESV =

η⎛ ⎞⎜ ⎟⎜ ⎟⎝ ⎠

V … (4)

Now, Eq.(3) can be written as,

disp_std mod_mot

c_mod std_mot

ηECR = 1 +

η⎛ ⎞⎜ ⎟⎜ ⎟⎝ ⎠

VV

… (5)

From Eq.(2), clearance volume can be calculated and then, motored volumetric efficiency of the modified engine required for the given combination of IVCT and clearance volume can be calculated from Eq.(5) as,

( ) c_modmod_mot std_mot

disp_std

ECR-1η = η

⎡ ⎤⎢ ⎥⎢ ⎥⎣ ⎦

VV

… (6)

Then, IVCT of the modified engine is fixed such that, the measured motored volumetric efficiency of the modified engine is as close as possible to the value obtained from Eq. (6). Practically, for fixing the IVCT, first a guess value was used, then by motoring the engine at a speed of 1200 rpm, motored-volumetric efficiency was determined and cross checked with the value calculated from Eq.(6). If they were not in good close agreement, IVCT was changed and the above said procedure was repeated until a good match was found. However, for the computer simulation, the value obtained from the experiments has been used as input data. Areas of opening for intake and exhaust valves are calculated based on the geometry of the valve mechanism1. The values of the coefficient of discharge for burned and unburned gases as a function of valve-lift to valve-throat diameter are taken from Ramos8. Experimentally measured valve-lift data with respect to crank angles is used as an input for calculations1.

Fig. 1—Pressure-volume diagram of actual cycle for simulation of modified engine

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Ineffective compression process During early stages of compression stroke when inlet valve is still open (BDC to IVC), some of the cylinder charge is driven back to the intake manifold due to the piston motion. The mass of charge driven back during this period is calculated by considering cylinder and inlet manifold as plenums, and inlet valve as orifice9-11. The following equations are used to calculate the mass flow rate. If the flow is subsonic, then,

k 1 k 1k k

d in mm m

d 2 1d ( 1)

− −⎡ ⎤⎛ ⎞ ⎛ ⎞⎢ ⎥= −⎜ ⎟ ⎜ ⎟⎢ ⎥− ⎝ ⎠ ⎝ ⎠⎢ ⎥⎣ ⎦

m K p pC A pt RT K p p

… (7)

If the flow is supersonic, then,

k 1k 1

d ind 2d 1m KC A pt RT K

+−⎛ ⎞= ⎜ ⎟+⎝ ⎠

… (8)

Effective compression process The actual compression process starts only when the intake valve is closed. During this process, trapped cylinder charge and residual gases from previous cycle are assumed to mix perfectly and perfect gas equations are assumed to hold good9-12. Pre-flame reactions are neglected during compression process due to comparatively low prevailing pressure and temperature. The composition of the cylinder charge during this process is calculated according to Heywood13. The analysis of the compression process is done according to the procedure of Horlock and Winterborne12. The equations for the calculation of change of pressure and temperature are as follows;

v v

d d d 11d d dp R V R Qp

C C V⎡ ⎤⎛ ⎞

= − + +⎢ ⎥⎜ ⎟θ θ θ⎢ ⎥⎝ ⎠⎣ ⎦ … (9)

And for the unburned mixture denoted by suffix ‘u’ is given by,

uu

d 1 d 1 dd d dT V pT

V p⎡ ⎤

= +⎢ ⎥θ θ θ⎣ ⎦ … (10)

where, Q is the heat transfer rate, Cv is the specific heat, T is the temperature of gas, m is the mass of the gas and θ is the crank angle at any instant. Combustion process Combustion process is analyzed by considering two-zone combustion model. The total combustion

process is split into two processes such as ignition delay and combustion duration. The calculation of ignition delay and combustion duration is briefly explained below. Several models were considered for the analysis of ignition delay and combustion duration, however in this work we found that the following models were suited better. Ignition delay It is assumed that, during delay period, combustion takes place at constant volume in between the spark plug electrodes. Delay period in terms of crank angle degrees is assumed to be equal to number of crank angle degrees required to burn one thousandth of the cylinder volume at the spark discharge point12, which is dictated by flame speed. As far as main cycle calculations are concerned the cylinder contents are considered to follow normal compression process due to piston motion. Ignition delay in terms of crank angle degrees is given by,

T

fid S

rN360=Δθ … (11)

13

cyl0.00123

f

Vr

⎡ ⎤⎢ ⎥

= ⎢ ⎥⎢ ⎥π⎣ ⎦

and ST = FFSL * … (12)

Different correlations for laminar flame speed suggested by Kuelh12, Tiggelens14,15 and Mallard-le-chatellier16 were tried and the correlation of Kuelh is found to be best suited for the present work. The flame front factor FF was selected in such a way that the predicted peak pressure value matches with experimental value. The value of 3.5 was found to be good enough in this study1,4. Combustion duration A two-zone combustion model is considered for the analysis with ten product species. The ten product species considered are: CO2, CO, O2, O, NO, N2, H, H2, OH, H2O. The calculation of chemical equilibrium composition of product species has been carried out according to Campbell10. Once the ignition starts, with the establishment of a stable flame kernel, the flame is assumed to propagate spherically from the spark discharge point with the release of thermal energy12. The growth of the burned volume is determined from the absolute velocity of flame. The burned volume is calculated by taking into

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account the flame radius and flame-front area. At the same time, area of cylinder and piston walls that are in contact with the burned and unburned mixture are also calculated1,14, which are required for the heat transfer calculations. Calculations of flame propagation and the accompanying changes in various thermodynamic properties are carried out in a step-by-step manner using a numerical method for integration as suggested by Horlock and Winterborne12. Combustion said to be end when the unburned mixture volume becomes zero12. The following models were used for the calculation of the change in unburned and burned temperature, pressure.

u u u

u pu u pu

d dd 1d c d c dT v Qp

m m= +

θ θ θ … (13)

u u

p p pu u

p

u u u up p

p p

ddd d d

dd ddd d d

R T mR TVT p p p

R V R Qp V pm Rpc pc p

⎡ ⎤⎛ ⎞− − −⎢ ⎥⎜ ⎟θ θ⎝ ⎠⎢ ⎥=

⎢ ⎥θ − +⎢ ⎥θ θ θ⎣ ⎦

…(14)

( )p

p

pu

u u

p pu

u u

v pup u v p u

p p

vv u u

p p v

v vvuu u

p p p p

dd1d d

d dd dd

d

c mRVp e e c T TR R

cc R Q Qc R cp

c ccR V V Vc R c R

⎧ ⎫⎡ ⎤⎛ ⎞ ⎛ ⎞⎪ ⎪− + − − −⎢ ⎥⎜ ⎟ ⎜ ⎟⎜ ⎟ ⎜ ⎟θ θ⎪ ⎪⎢ ⎥⎝ ⎠ ⎝ ⎠⎣ ⎦⎨ ⎬⎡ ⎤⎪ ⎪+ − −⎢ ⎥⎪ ⎪θ θ⎢ ⎥⎣ ⎦⎩ ⎭=

θ ⎡ ⎤− −⎢ ⎥

⎢ ⎥⎣ ⎦

… (15)

The heat transfer for both burned and unburned gases is calculated using Woschni’s correlation17. The total rate of heat transfer from the system is,

pu dddd d d

QQQ= +

θ θ θ … (16)

and rate of work done is given by, d dd dW Vp=θ θ

… (17)

The pressure change, dp/dθ, can be obtained directly from Eq. (15) since all the terms are known. The temperature changes, dTu/dθ and dTp/dθ, can be directly evaluated using Eqs (13) and (14) respectively.

Expansion process Once the combustion is completed, it can be assumed that the expansion of the products of combustion starts. During the expansion process, the variables are organized for the single-zone calculations12. The governing equations for the analysis of the expansion process are similar to that of compression process except that the working fluid is the products of combustion. Therefore, properties of the burned products should be considered. Gas exchange process The gas exchange process is analyzed by considering the cylinder and manifolds (inlet and exhaust) as plenums and valves as orifices. The mass flow rate is determined during intake and exhaust processes. The detailed analysis of the gas exchange process has been carried out according to Campbell10 and Ganesan9. And if p/po is less than critical pressure ratio8,9, then the flow is subsonic and mass flow rate is given by,

K 1 K 1K K

d o0 0

d 2 1d ( 1)m K p pC A pt RT K p p

− −⎡ ⎤⎛ ⎞ ⎛ ⎞⎢ ⎥= −⎜ ⎟ ⎜ ⎟⎢ ⎥− ⎝ ⎠ ⎝ ⎠⎢ ⎥⎣ ⎦

… (18)

where Cd is the coefficient discharge, A is the area of opening for intake/exhaust valves. The flow is supersonic if (p/po) exceeds critical pressure ratio and mass flow rate is given by,

K 1K 1

dd 2d 1m KC A pt RT K

+−⎛ ⎞= ⎜ ⎟+⎝ ⎠

… (19)

Heat transfer models The heat transfer between working fluid and internal surfaces of the combustion chamber is predominantly by forced convection and radiation. In the case of spark ignited engines, combustion due to homogeneous charge takes place entirely at gaseous phase and radiation is correspondingly limited. Radiation fraction may reach about 20% of the total heat transfer inside the cylinder12. The only gases that radiate appreciably, among those occurring in combustion products are carbon dioxide and water vapor. There are some published observations on spark ignition engines, where radiative heat transfer probably does not play an important role due to absence of the soot particles12. Therefore, it is

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common to include the radiation effect into the convective heat transfer by adjusting the appropriate coefficient. However, in the present work, radiative heat transfer is neglected. Convective heat transfer coefficient is calculated from the correlation proposed by Woschni17. Final form of the equation for heat transfer coefficient used in this work is as follows:

( ) 0.82 c c m-0.2 0.8 -0.546

c c 1 p

- =110 +

m RC V p p

h D p T C S⎡ ⎤⎢ ⎥⎣ ⎦

×1.163 (W/m2-K) … (20) Friction model Reduction of indicated mean effective pressure due to various frictional losses of the engine is evaluated using empirical equations developed by Bishop18. The working forms of the equations are obtained from Ganesan9. Unburned hydrocarbons, carbon monoxide and nitric oxide emissions For the analysis of unburned hydrocarbon emissions, the effect of flame quenching at the entrance of the various crevice volumes, open combustion chamber walls33 and oil layer absorption and desorption of fuel vapors34 are considered. CO and NO emissions are predicted from the chemical equilibrium composition during combustion and expansion processes9-12. Experimental Study Experimental investigations have been carried out for two reasons. Firstly, to validate the computer code developed and secondly, to determine experimentally the optimum GER/ECR ratio of the modified engine. Experiments have been carried out on a modified single cylinder, air cooled diesel engine. Engine specifications and experimental setup are shown in Appendix. A large plenum was provided in the inlet manifold1,6 to accommodate the pushed back charge during the early stages of the compression stroke so that minimum rise of pressure takes place in the inlet manifold. A suitable carburetor was modified and used to control the fuel and airflow independently to obtain the required air-fuel ratios1. In order to avoid the flow reversal through the carburetor due to the pushed back charge, a reed-valve in the inlet manifold was provided (Fig. 2)1. To prevent undue condensation of the charge in the inlet manifold due to large residence and over expansion of gases, warm water from cylinder head cooling was used. For the

purpose of heating the charge, a heating chamber (Fig.2) was provided in the inlet manifold immediately after the carburetor1. Also, cylinder-head cooling was required in order to prevent knocking. In the modified engine, for every configuration of the engine, a separate piston with suitable bowl volume was used. Always, piston bowl shape was kept as a part of the sphere with top land area equal to 25% of the cross sectional area of the piston1,19. This was done with the intention of keeping uniform theoretical-maximum-squish-velocity for all the configurations considered, thereby eliminating the effect of changes in squish-velocity on the combustion process23. The intake-cam of the conventional engine was modified to accommodate the changes in IVCT for modified engine. In the modified camshaft, exhaust cam profile and timings were kept the same as that of the conventional camshaft, whereas the intake cam alone was modified as shown in Fig.31,6. Intake cam was split into two halves of equal thickness perpendicular to axis of the cam called as two cam

Fig. 2—Block diagram of modified induction system

Fig. 3—Modified cam lobes of modified engine

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lobes. Lobe-1, used for inlet valve opening, has a keyway and could be moved on the camshaft to any desired position along the axis. Lobe-2, used for inlet valve closing, has thread mounting on the camshaft and could be rotated to any desired angular position relative to Lobe-1 to change the top dwell period. After adjusting the two lobes to obtain the required inlet valve opening and closing timing, both of them were held firmly by lock-nuts. When a large IVCT was required, the middle-sector was inserted in between the Lobe-1 and 2 for high top dwell periods. The rise and fall angles of the stock cams were each 61° with a top dwell angle equal to zero degree. To maintain the same acceleration characteristics of the stock cam with increased top dwell period, the modified intake cam was made with rise and fall angles of 62° and top dwell angle of 8°. Both stock and modified cams were of polydyne type6. The experimental investigations were carried out for two compression ratios of 7 and 8 on both standard and modified engines. Four valve timings corresponding to four GER/ECR ratios ranging from 1.25 to 2 in equal steps were employed in the modified engine (Table 1). Experiments have been conducted at a speed of 1200 rpm. Measurements were made for different loads. For every load, spark timing was set for minimum advance for maximum brake torque (MBT timing). In order to have knock-free operation, an air fuel ratio of 18:1 was used at all the loads. But, at very low loads, it was not possible to operate with such a lean mixture; therefore at low loads, air-fuel ratio was adjusted for steady operation of the engine. Results and Discussion

It may be noted that for the figures referred to in the following discussion, predicted results are plotted with solid and dashed lines, whereas experimental

values are represented by symbols. Here, conventional spark ignition engine is referred to as standard engine, whereas new version of the engine is referred to as modified engine in almost all places.

As a first step, motoring tests were conducted on both standard and modified engines for CRs of 7 and 8, at 1200 rpm and with different GER/ECR ratios (1.25 to 2.0) of the modified engine. Since, the actual volumetric efficiency varies with the inlet charge composition; motored volumetric efficiency was used in the calculation of clearance volume, effective compression ratio and to fix the inlet valve closure timing (IVCT) of the modified engine.

The GER/ECR ratio was selected arbitrarily and was spaced at an equal interval of 0.25, starting from standard engine’s value of 1.0. GER/ECR ratio beyond 2.0 was not tried because it reduces the volumetric efficiency of the modified engine to a very low value even at full throttle operation and also, sometimes it leads to over expansion losses2,4,6. In addition, at higher GER/ECR ratios, the spark timing would overlap with the inlet valve closure timing21.

As mentioned earlier, first GER/ECR ratio of the modified engine was chosen and then the clearance volume of the modified engine was calculated using Eq. (2). Afterwards, inlet valve closure timing was fixed according to Eq. (6). For all the configurations, reduction in the clearance volume and motored volumetric efficiencies for different GER/ECR ratios are shown in Table 1. Inlet valve closure timing and clearance volume were adjusted such that the motored peak pressure was very close to that of standard engine at the corresponding compression ratio. The IVCT setting of the modified engine with increase in GER/ECR ratio is shown in Fig. 4.

Figure 5 shows the motored peak pressures at different GER/ECR ratios. When the motored peak pressures of standard and modified engine are almost

Table 1⎯Different parameters used in the present work

Speed = 1200 rpm Inlet valve closure timing

(deg. ABDC) Clearance volume

(cc) Motored volumetric

efficiency (%)

Compression ratio

GER/ ECR ratio

8:1 7:1 8:1 7:1 8:1 7:1

1.00 35 35 94.5 110.3 81.0 81.0 1.25 95 92 73.5 85.4 60.7 60.4 1.50 108 106 60.1 69.6 49.6 49.3 1.75 128 126 50.9 58.8 42.0 41.6 2.00 135 134 44.1 50.8 36.4 36.0

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equal, the ECR of the modified engine will be the same as that of CR of the standard engine. It can be observed from Fig. 5 that, in the case of compression ratio 8, standard engine’s motored peak pressure (at GER/ECR ratio 1.0) is about 14.6 bar, whereas for modified engine, at all the GER/ECR ratios, the peak pressure values are almost equal to this value. This has been achieved by setting appropriate IVCT and clearance volume thereby ECR of the modified engine is equal to 8 at all GER/ECR ratios.

Fig. 6— Variation of volumetric efficiency with GER/ECR ratio

The variation of volumetric efficiency with increase in GER/ECR ratio of the modified engine is shown in Fig. 6. From Fig. 6, it was observed that ECR has almost no effect on the volumetric efficiency. It was found to be a function mainly of IVCT, clearance volume and throttle position. The volumetric efficiency of the standard SI engine at full throttle was around 80% for both ECRs. The original diesel version of the engine had a volumetric efficiency of about 90% at full throttle. The reduction in the volumetric efficiency for the standard SI version may be mainly due to the modifications done to the inlet manifold.

With increase in GER/ECR ratio, the volumetric efficiency of the modified engine decreases and therefore, it is obvious that the brake power output of the engine will also decrease with increase in GER/ECR ratio. Figure 7 shows the variation of the brake power output with increase in GER/ECR ratio. The brake power output of the modified engine at ECR 8, and for GER/ECR ratio of 1.25, 1.5, 1.75 and 2 were about 82.4%, 67.6%, 52.9% and 35.3% respectively compared to that of the standard engine. By reducing compression ratio from 8 to 7, brake power output has further decreased by about 9.3% at all the loads. This may be mainly attributed to the reduction in average pressures at compression ratio 7. It is proven from the facts that the reduction of compression ratio of a SI engine reduces the brake thermal efficiency and also, due to the reduction of

Fig. 4—Variation of inlet valve closure timing with GER/ECR ratio

Fig.5—Measured motored peak pressure at various GER/ECR ratio

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peak pressure and bmep, the power output reduces11−13.

In this work, the main aim is to operate the modified engine at full throttle at all the GER/ECR ratios (i.e. at a given IVCT and clearance volume). However, to obtain very low power outputs (say < 35%), a little throttling was applied. Therefore, it is to be expected that, at very low power outputs, the brake thermal efficiency will deteriorate. However, at a particular power output of both the engines, if comparison is made, it was found that the values for modified engine are better than the standard engine’s values. This is evident from Figs 8 and 9, which show the variation of brake thermal efficiency with brake power for ECR 8 and 7 at GER/ECR ratio of 1.5. Similar trends were observed for other GER/ECR ratios considered in the study1.

Considering the maximum output points of modified engine at various GER/ECR ratios, the brake thermal efficiency variations have been plotted in Fig.10. It is observed that, the efficiency increases gradually with increase in GER/ECR ratio up to 1.5 for both the compression ratios and then it decreases thereafter with further increase in GER/ECR ratio. This may be attributed mainly to the decrease of mechanical efficiency at low loads. The variation of mechanical efficiency with brake power output for GER/ECR ratio of 1.5 and 2.0, for the compression ratios 8, is shown in Fig. 11. From Fig.11, it can be seen that the mechanical efficiency reduces

considerably with increase in GER/ECR ratio. Similar trends were observed for the compression ratio of 7 also1,4. Further, the indicated power decreases with decrease in load and the fraction of the indicated power required to overcome frictional and pumping losses increases.

It is observed from Fig.10 that, at GER/ECR ratio of 1.5, for ECR 8, the brake thermal efficiency is about 31.75% as against the standard engine efficiency of about 31.25% (at GER/ECR ratio 1.0).

Fig. 7— Variation of brake power with GER/ECR ratio

Fig. 8— Variation of brake thermal efficiency with brake power

Fig. 9— Variation of brake thermal efficiency with brake power

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From the error analysis, it is estimated that the error in the measurement of brake power is about 1.33%, for fuel consumption it is about 0.9%, and for air consumption, the value is about 1.1%. Therefore, there is about 1.6% absolute improvement in brake thermal efficiency over standard engine. Similarly, for the compression ratio 7, the absolute improvement is found to be about 5%, which can be considered as

Fig. 12— Envelope of maximum brake thermal efficiency for standard and modified engine

Fig. 13— Envelope of maximum brake thermal efficiency for standard and modified engine

quite good. Even though the improvement in brake thermal efficiency appears to be low when absolute values are seen, but when the comparisons are made between the efficiencies of modified and standard engine at the same power outputs, the improvement in efficiency of the modified engine is substantial as it is evident from Figs 12 and 13. From these figures, it can be seen that the improvement in brake thermal

Fig. 10— Variation of brake thermal efficiency with GER/ECR ratio

Fig. 11— Variation of mechanical efficiency with brake power

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efficiency is more for compression ratio 7 than 8. From Figs 12 and 13, it can be noticed that, for the modified engine, at compression ratio 8, a maximum improvement of 14.6% has been achieved, and for compression ratio 7, the improvement is about 23.8% and also, the gain in efficiency increases with increase in GER/ECR ratio. Further, it can be seen from Figs 8 and 9 that with the increase in GER/ECR ratio, brake thermal efficiency increases and the predicted values match quite well with experimental values for both the engines.

Pumping power represents the power required for gas exchange processes during the intake and exhaust strokes. In standard SI engines, throttling operation decreases the absolute pressure in the intake manifold. The pressure differential between exhaust and intake manifold increases with decrease in load and consequently, so does the power required for the gas exchange process. In standard engines, pumping losses would alone account for nearly one fifth of the brake power and it increases with decrease in power output. Figure 14 shows the variation of pumping power with the load for GER/ECR ratios 1.5 and 2.0 for the compression ratio of 8. From Fig.14, it is clear that there is a substantial decrease in pumping power for modified engine compared to that of standard engines.

It can be observed from Fig. 14 that at full load operation, the pumping losses are almost remaining the same for both the standard and modified engines.

But, at a particular brake power, if the comparison is made between standard and modified engines, there is an appreciable reduction in the pumping losses. For example, with GER/ECR ratio 2.0, ECR 8, the pumping power of the modified engine is about 0.095 kW at a brake power of 1.5 kW. Meanwhile, for standard engine, at the same brake power, the pumping power is about 0.216 kW (Fig. 14). So, there is a reduction of about 56% in the pumping power. The reduction in the pumping power will be responsible for improving the brake thermal efficiency. However, improvement in brake thermal efficiency of the modified engine is not only due to reduction in pumping losses alone, but it can also be due to longer expansion1,6.

In the case of modified engine, whenever IVCT is changed, clearance volume is also changed so that motored peak pressures are more or less remaining constant (Fig. 5). Figure 15 shows the variation in firing peak pressures with increasing GER/ECR ratios. From Fig. 15, it can be observed that, the cylinder peak pressures at full throttle are remaining almost constant, at all the GER/ECR ratios. Obviously, cylinder peak pressures are comparatively less for ECR 7 than 8. In fact, cylinder peak pressures are strongly dependent on the spark timing also.

In the present work, spark timing was adjusted to MBT timing at all loads. It may be mentioned here that the engine used in this investigation was originally a diesel engine having a large bore, and

Fig. 14— Variation of pumping power with brake power

Fig. 15— Variation of peak pressure with GER/ECR ratio

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also the spark plug was not located centrally (offset by about 15 mm). Therefore, at full loads the engine in spark ignition mode was knocking and very rarely at part loads. So, at full load operation, the spark timing was adjusted to knock-free operation rather than MBT timing. As knock-free operation at full loads, calls for retarding of spark timing results in comparatively low peak pressures. However, it is reported that the variation in the peak pressure does not seem to alter the IMEP much but only very marginally24.

In case of modified engines, improvement in brake thermal efficiency is primarily achieved by lower pumping power and prolonged expansion. In addition, a constant peak pressure maintained even at lower power outputs, achieved through constant ECR may also contribute. Earlier work on VCR engine has reported that the improvement in the fuel economy is about 10%26 due to variable compression. Previous work by Nagesh6 has demonstrated that improvement in brake thermal efficiency is only about 8% for GER/ECR ratio 2, compression ratio 8 and 1200 rpm when no changes are made to clearance volume (i.e. not keeping ECR constant). But, when constant ECR was maintained, the improvement was about 17%. This shows the benefit of ECR on fuel economy.

In this work, for predicting the unburned hydrocarbon emissions, quenching of flame at the open combustion chamber wall surfaces, at the entrance of crevice volumes in the combustion

chamber and absorption and desorption of fuel vapours by lubricating oil layers have been considered. Figures 16 and 17 show the variation of UBHC emissions with brake power for standard as well as modified engine at GER/ECR ratios of 1.5 and 2 for ECR 8 and 7 respectively. Unburned hydrocarbon emissions are lower at lower compression ratios27, 28.

Generally, UBHC emissions in the exhaust gases are mainly due to quenching of flame reactions by combustion chamber surfaces29,30. Therefore, when compression ratio is reduced, UBHC emissions must be reduced. However, when GER/ECR ratio was increased, clearance volume was reduced which resulted in higher surface to volume ratio. Therefore, it can be estimated that, UBHC emissions would increase with increase in GER/ECR ratios. But, on the contrary, when GER/ECR ratio is increased, expansion is prolonged allowing more time for UBHC to oxidize. Hence, an UBHC emission at higher GER/ECR ratios depends upon the factor which predominates among the above two. In this work, the latter one seems to be predominating, since total UBHC emissions are low at high GER/ECR ratios. When compression ratio is reduced, clearance volume will be increased which results in larger volume of residual exhaust gases27. Then, the last part of the charge to be exhausted will be rich in UBHC29. Therefore, larger volume of residual gases results in increased UBHC emissions in the exhaust. But, a

Fig. 16— Variation of unburned hydrocarbons with brake power

Fig. 17— Variation of unburned hydrocarbons with brake power

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lower compression ratio engine would have comparatively higher exhaust temperatures and this would promote oxidation of UBHC in the burned mixture reducing UBHC emissions27,29. These factors are responsible for reducing UBHC emissions at lower compression ratios. Previous work by Siewert31 has demonstrated that, there were no significant changes in UBHC and NOX emissions when IVCT was changed. Therefore, it may be concluded that, changes in IVCT may not have much influence on the UBHC emissions, however changes in GER/ECR ratio, compression ratio and clearance volume can have a significant influence. The variation of UBHC emissions with different GER/ECR ratio for both ECRs of 8 and 7 is shown in Fig. 18.

Figure 19 shows the reduction in UBHC emissions from modified engine in respect to standard engine. It can be observed from Fig. 19 that there is a considerable reduction of UBHC emissions when GER/ECR ratio is increased in modified engine. Also, from Fig.19, it can be seen that reduction up to 58% is possible with ECR of 7 and up to 50% is possible with ECR of 8. Further, it is seen that, lower compression ratios give better reduction. It may be attributed to the fact that at lower CRs due to higher exhaust temperature results in better oxidation29, 31.

Carbon monoxide (CO) emissions from the internal combustion engines are primarily controlled by the air-fuel ratios. Since at full loads, lean mixtures were used, carbon monoxide emissions were almost at zero

level. Figures 20 and 21 show that carbon monoxide emission with variation in brake power for both ECR 8 and 7 at GER/ECR ratios of 1.5 and 2 respectively. Since the engine was operated with lean mixture most of the time, there are no significant variations with configuration as far as the CO emissions are considered except at very low loads (mixture is rich).

Figures 22 and 23 show predictions (measurements were not made) of the nitric oxide emissions at

Fig. 18— Variation of unburned hydrocarbons with GER/ECR ratio

Fig. 19— Percentage reduction of unburned hydrocarbons with GER/ECR ratio

Fig. 20— Variation of carbon monoxide emissions with brake power

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different configurations. NO reactions are assumed to be frozen at 1750 K during the expansion process13. Literature available on the NOX emissions indicates that, among all the factors that affect NOX emissions, mainly it is controlled by the peak temperature of the cycle and availability of oxygen. Since, it is difficult to measure the peak temperature exactly; it is calculated from the cylinder pressure data and corresponding cylinder volume. It is also found that

NOx emission reduces with increase in GER/ECR ratios. As discussed earlier, it is due to longer expansion and more time for heat transfer13. The quantity of NOX formed depends upon the air-fuel ratio, peak temperature, the time available for NOX reactions to take place at the peak temperature, and the occurrence of peak pressure during the cycle33. Obviously, the formation of NOX is expected to be low with increase in GER/ECR ratio. Summary of results (i) For modified engine, at GER/ECR ratio 1.5,

and compression ratio 8, the brake thermal efficiency is found to be a maximum with 31.75% and at compression ratio 7, it is 31.5%. The values are less for other GER/ECR ratios. Therefore, for the modified engine, GER/ECR ratio of 1.5 is considered to be optimum to give the possible maximum brake thermal efficiency.

(ii) At GER/ECR ratio 1.5 (maximum efficiency point), the improvement in brake thermal efficiency over the standard engine is about 14.7% at compression ratio 8 and at other GER/ECR ratios it varies from about 7.9 to 19%. Similarly, at compression ratio 7, the maximum improvement is 25% and at other GER/ECR ratios, it varies from 11 to 23.5%.

(iii) Reduction of compression ratio from 8 to 7 seems have a significant effect in the

Fig. 21— Variation of carbon monoxide emissions with brake power

Fig. 22— Variation of nitric oxide emissions with brake power

Fig. 23— Variation of nitric oxide emissions with brake power

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improvement in the brake thermal efficiency of the modified engine. For example, at compression ratio 8 with GER/ECR ratio 2.0, the improvement is about 19%, whereas for the same conditions with compression ratio 7, it is about 23.3%.

(iv) For the modified engine, at both compression ratios 8 and 7, with GER/ECR ratio 1.5, there seems to be about 34% reduction in the brake power output compared to the standard engine.

(v) However, at GER/ECR ratio 2, the brake power output of the modified engine is about 33% of the maximum power of the standard engine. This is an important operating condition since most of the time the average road-load of a SI engine is about one third of the maximum power output of the engine.

(vi) For the modified engine, at compression ratio 8, with GER/ECR ratio 1.5, the reduction in the pumping power is about 26% and reduction in the hydrocarbon emissions is about 31%. At GER/ECR ratio 2, the reduction in hydrocarbon emissions is about 51%.

(vii) CO emissions from both the standard and modified engine are almost nil at full load operation.

(viii) Theoretical prediction shows that the reduction in NO is possible to a great extent when the SI engine is operated with variable valve timing and variable clearance volume concept.

(ix) It is to be noted that in the present work, a diesel engine has been modified and the speed considered is only 1200 rpm. This is due to the fact that the main aim of this paper is to prove the concept.

Conclusions In this paper, it has been shown that by the use of

present day automatic and situation based valve actuating mechanisms35-38 for variable valve timing and variable clearance volume, performance of the SI can be improved considerably especially at part-loads.

The theoretical predictions match quite well with the experimental results. Therefore, the computer code developed in this work can be used with confidence for the prediction of performance and exhaust emission characteristics of the modified engine with variable valve timing and clearance volume, and to carry out the further parametric studies and also to optimize the GER/ECR ratio of the engine for the given configuration.

Nomenclature θ = crank angle A = area C1, C2 = constants Cd = coefficient of discharge Cp = specific heat at constant pressure CR = compression ratio of the standard engine D = cylinder bore ECR = effective compression ratio ESV = effective swept volume FF = flame factor GCR = geometric compression ratio GER = geometric expansion ratio GER/ECR = ratio of expansion to compression ratio of the

modified engine hc = heat transfer coefficient IMEP = indicated mean effective pressure IVC = inlet valve closure IVCT = inlet valve closure timing K = ratio of specific heats m = mass N = speed in rev/min p = pressure R = gas constant r = radius S = flame speed SI = spark ignition SP = piston speed T = temperature TDC = top dead center UBHC = unburned hydrocarbons V = volume W = work transfer η = efficiency Suffix c = clearance c, cyl = cylinder disp = displacement eee = modified engine f = flame id = ignition delay in = inlet L = laminar m = manifold mod = modified engine mot = motored o = motoring, manifold std = standard engine T = turbulent u = unburned References 1 Mallikarjuna J M, Numerical and Experimental Studies on

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Appendix Engine Specifications Model : Air-cooled, single-cylinder, diesel Bore : 87.5 mm Stroke : 110 mm Connecting rod length : 232 mm Displacement : 661.45 cm3

Compression ratio : 16.5:1 Rated output : 4.4 kW at 1500 rpm Valve timings: IVO : 4.5° bTDC IVC : 35.0° aBDC EVO : 35.0° bBDC EVC : 4.5° aTDC Above engine was converted into spark ignition mode

Fig. A.1—Experimental set-up