Effect of various parameters on performance of IC engine

SEMINAR REPORT ON

EFFECT OF VARIOUS PARAMETERS ON IC ENGINE PERFORMANCE Submitted by Ajinkya D. Jagtap

(MIS NO: 121595010) Submitted in partial fulfillment of the requirement of the degree of MASTERS OF TECHNOLOGY IN Automotive Technology UNDER THE GUIDANCE OF Professor Aatmesh Jain COLLEGE OF ENGINEERING, PUNE

DEPARTMENT OF MECHANICAL ENGINEERING

ARAI ACADEMY AUTOMOTIVE RESEARCH

COLLEGE OF ENGINEERING, PUNE- 411005

ASSOSIATION OF INDIA, PUNE

YEAR 2015-2016

CERTIFICATE

This is to certify that the report entitled “EFFECT OF VARIOUS PARAMETERS ON IC ENGINE PERFORMANCE” submitted by AJINKYA D. JAGTAP (MIS No. 121595010), in the partial fulfillment of the requirement for the award of degree of Master of Technology Automotive Technology of College of Engineering Pune, Pune, and ARAI Academy, ARAI, Pune, is approved. Place: Pune Date:

Dr.S.N.Sapali

Dr.K.C.Vora

HOD mechanical department,

Dy. Director & Head, ARAI Academy

COEP, Pune

ARAI, PUNE

Prof. Aatmesh Jain Seminar guide

Prof.K.P.Wani Internal Examiner

ACKNOWLEDGEMENT I feel great pleasure in submitting my seminar report on “EFFECT OF VARIOUS PARAMETERS ON IC ENGINE PERFORMANCE”. I wish to express true sense of gratitude towards my guide Prof. Aatmesh Jain for his constant encouragement and valuable guidance and for providing all necessary facilities, which were indispensable in the completion of this seminar. I would like to thank ARAI & COEP who extended their kind support during the accomplishment of the seminar report. Finally, I express my sincere thanks to all those who helped me directly and indirectly in many ways in completion of this seminar.

ABSTRACT The invention of internal combustion engine had taken place more than century ago. Since that time these engines have continued to develop as our knowledge of engine process has increased. As nature of human has been continuous evaluation and the IC engine is no exception for that. Today’s engines have taken new dimensions towards all aspect to that of conventional engines. With a growing demand for transportation IC engines have gained lot of importance in automobile industry. It is therefore necessary to produce efficient and economical engines. While developing an IC engine it is required to take in consideration all the parameters affecting the engines design and performance. There are enormous parameters so it becomes difficult to account them while designing an engine. So it becomes necessary to conduct tests on the engine and determine the measures to be taken to improve the engines performance. In this seminar review the effect of different parameters on performance of IC engine. Engine performance can be measured in terms of Power output, various efficiencies, emission etc. There are large number of parameters which affects the engine performance. These parameters are basically of two types one is Design parameters and second is operating parameters. As scope of this subject is very large we will restrict our review to only some basic parameters such as speed, load, A/F ratio to name a few. Also we will discuss some advanced technologies. IC engines are basically classified as Spark ignition engines and Compression ignition engines. Some parameters are more significant for SI engines while that may be insignificant for CI engine. We will first discuss what are the parameters and their significance with respect to IC engine. After that we will focus on how to measure them and analyse result obtained from them.

Table of Contents 2

Introduction................................................................................................10

3

Literature review.........................................................................................12

4

Stroke to Bore Ratio...................................................................................13

5

Compression Ratio.....................................................................................15 5.1

SI Engine Performance...............................................................................................16

5.2

CI Engine Performance...............................................................................................19

6

Valve Timing and Valve Overlap................................................................23 6.1

Effect of changes to Intake Valve Opening Timing....................................................23

6.2

Effect of changes to Intake Valve Closing Timing......................................................25

6.3

Effects of Changes to Exhaust Valve Opening Timing...............................................26

6.4

Effects of Changes to Exhaust Valve Closing Timing................................................28

6.5

Effect of Overlap angle (IVO – EVC)........................................................................29

7

Air-Fuel Ratio.............................................................................................32

8

Ignition Timing...........................................................................................37

9

Load and Speed..........................................................................................42

10

Injection Timing and Injection Pressure.....................................................45

11

EGR............................................................................................................49

11.1

SI Engine Performance............................................................................................49

11.2

CI Engine Performance...........................................................................................50

12

Swirl............................................................................................................53

13

Cetane Number...........................................................................................57

14

Future Scope...............................................................................................60

15

Conclusion..................................................................................................61

16

References..................................................................................................62

TABLE OF FIGURE Figure 3.1 Effect of Stroke to Bore on IMEP............................................................................13 Figure 3.2 Effect of Stroke to Bore Ratio on IFSC...................................................................13 Figure 4.1 Effect of Variation of CR on Brake Thermal Efficiency..........................................16 Figure 4.2 Effect of Variation of CR on IMEP and Thermal Efficiency...................................16 Figure 4.3 Effect of Variation of CR on IMEP and Thermal Efficiency...................................17 Figure 4.4 Effect of Variation of CR on Brake Thermal Efficiency..........................................18 Figure 4.5 Effect of Variation of CR on Fuel Consumption......................................................18 Figure 4.6 Effect of Variation of CR on Brake Specific Fuel Consumption.............................19 Figure 4.7 Effect of Variation of CR on NOx............................................................................19 Figure 4.8 Effect of Variation of CR on PM, CO, HC...............................................................20 Figure 5.1 Standard Valve Timing for SI and CI engine...........................................................22 Figure 5.2 Effect of IVO on Brake Power.................................................................................23

1 Introduction

The more and more severe regulations on exhaust emissions from vehicles and worldwide demand for fuel consumption reduction leads to search of new ways towards achieving them. There are several options available for improving the fuel economy and reduce emissions, we will discuss them briefly. First, you can improve current technologies implemented to control the economy and emissions from present IC engines. Second option is to use alternate fuels so as to compensate the demand for conventional fuels. Third option is to implement concept of electric or hybrid electric vehicle. As last two options are still in early phase of development; although they are promising, they have to go far to compete with option one. First option is traditionally used and evolved in many aspects. As it is being said that today’s IC engines are at the verge saturation towards further evolution, but it has been seen that there is still large scope for development. Extensive research has been going on to understand the actual working of all process and effect of various variables on performance of IC engine. Optimizing major parameters have shown the improved performance. So, with the study of effect of various parameters deeply lead us to promising results. Today technology is capable of controlling most of engine variable but unless and until all variables studied thoroughly it can’t be reflected in performance of engine. Major prime movers used in automobile are four stroke Spark ignition engines and compression ignition engines. We restrict our review toward SI and CI engines only. There are large number of parameters that affects the performance of IC engine. As all parameters are contributing towards varied performance all are very important. To keep the scope of the review feasibly brief, we will discuss only the relatively important parameters that influence most. The parameters that can be controlled effectively can be said to be important. Performance of IC can be evaluated based on Power output, Fuel consumption and Emission can be evaluated in terms of part per million or g/kWhr. Power Output represented by Brake Power (BP), Indicated Power (IP), Torque, Indicated mean effective pressure (IMEP), Brake mean effective pressure (BMEP); all parameters signifies power but with different relation. Fuel consumption can be represented by Brake specific fuel consumption (BSFC), Indicated Specific fuel consumption (ISFC). Emission can be represented as Brake specific CO (bsCO), Brake specific HC (bsHC), Brake specific NO (bsNO). All above Performance parameters are affected by engine variables. Some engine variables are common for both SI and CI engines

such as Speed, Load, Compression ratio etc. Some Engine variables are different for SI and CI engine as per their different working principle. Spark timing is important parameter for SI engine and its CI engine counterpart is fuel injection timing. Engine variables can be separately discussed for SI and CI engine. Each variable is explored with its significance. Its effect on above discussed performance parameter explained with the graphical representation and reason for that is highlighted.

2 Literature review Extensive research is being done towards optimum engine variables to improve performance of Internal Combustion Engine. There is plenty of literature available on this topic. Prof. J. B. Heywood (MIT, USA) is notable person in field of engine performance. ‘Internal Combustion Engine Fundamentals’ by J. B. Heywood is one of best book to get fundamental of IC engines. Prof. Heywood (MIT) went with thorough explanation of performance characteristics and effect of engine variables with in depth analysis. We can categorize literature review in two parts as SI engine and CI engine. Diesel engine reviewed mostly with its injection parameters. Ayala, Gerty and JB Heywood (SAE 2006-01-0229) highlighted effect of Air-Fuel ratio, Compression ratio and load on SI engine efficiency. They came with experimental investigation of 5 liter naturally aspirated SI engine. In this paper effect spark timing and A/F ratio and CR on NIMEP and Net Indicated efficiency is discussed. Suwanchotchoung and Williamson (SAE 2003-32-0023) highlighted effect of Equivalence ratio (relative fuel air ratio) on Brake power, bsfc and bsCO, HC, NO emissions on 2 liter SI engine with manifold injection. Shehata and Abdel Razek (Engineering Research Journal 120, (December 2008) discussed variation of BP, BSFC, Efficiency with varied engine speed, load and EGR rate. Aina T., Folayan C. O. and Pam G. Y. (Advances in Applied Science Research, 2012, 3 (4):1915-1922) evaluated influence of compression ratio on the BP, bsfc, Brake thermal efficiency of a spark ignition engine. Hountalas, Kouremenos, Schwarz and Mavropoulos (SAE 2003-01-0340) detailed fuel injection timing effect on NOx, Soot, bsfc, Heat release rate (HER) with varied injection pressure upto 1683 bar and different load conditions. F. Mallamo, M. Badami and F. Millo (SAE 2005-01-0379) evaluated effect of variation of compression ratio (CR) and Injection pressure on emission of CRDI engine. They also came with NOx-PM trade-off with optimum injection timing and injection pressure with given CR. Kermani, Garsi, Ruhland and Kaudewitz (2013-24-0065) came with three different engine with varying Stroke/Bore ratio and all parameters remained same. They compared

these engines for IMEP, ISFC. Agarwal and Srivastava (Fuel 2013-science-direct) represented Effect of fuel injection timing and pressure on BSFC, Efficiency, bsCO, bsNOx and PM.

3 Stroke to Bore Ratio While there are many factors that contribute to an engine’s efficiency, the primary factor that needs to be considered is the engine geometry itself. Not only does the overall size of the engine matter, but the aspect ratio of the engine cylinders defined by the stroke-to-bore ratio also matters. Simple geometric relationships show that an engine cylinder with longer stroketo-bore ratio will have a smaller surface area exposed to the combustion chamber gasses compared to a cylinder with shorter stroke-to-bore ratio. The smaller area leads directly to reduced in-cylinder heat transfer, increased energy transfer to the crankshaft and, therefore, higher efficiency. Engine friction is affected by the stroke-to-bore ratio because of two competing effects: crankshaft bearing friction and power-cylinder friction. As the stroke-tobore ratio decreases, the bearing friction increases because the larger piston area transfers larger forces to the crankshaft bearings. However, the corresponding shorter stroke results in decreased power-cylinder friction originating at the ring/cylinder interface. 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) engines because it determines the overall dimensions of the engine for a given displacement. 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. Usually, in these studies, the S/B ratio changes between 0.7 and 1.4 as in modern engines and single-spark-ignition (SSI) engines having centrally located plug were commonly used. In general, a longer stroke leads to higher thermal efficiency through faster burning (reduction in combustion duration) and lowering the overall chamber heat loss. It also increases the maximum operating speed, maximum power, indicated mean effective pressure (imep), and also blow-by of the engine. In addition, the larger bore will provide more room for poppet valves in four-stroke engines. Hence, increasing the number of valves per cylinder for a given cylinder bore improves engine breathing. 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 NO emissions tended to decrease because of increasing crevice volume and decreasing temperature. The limiting factor in this relationship is the inertial forces origination from the piston motion. To achieve high power density, the engine must operate at a high engine speed (up to 18,000 rpm for the Formula 1 engine), which leads to high inertial forces that must be limited by using a small stroke-to-bore ratio. For applications that demand high efficiency, a long

stroke-to-bore ratio is necessary and, again because of the inertial forces of the piston, requires a slower engine speed and lower power density. For the marine application that has a 2.5 m stroke, the engine speed is limited to 102 rpm.

S/B 1.3 S/B 1.0 S/B 0.7

Figure 3.1 Effect of Stroke to Bore on IMEP Effect of Bore to Stroke ratio on IMEP is represented by Figure 3.1. As Engine speed increases the trend of IMEP is same for all S/B ratio but there is variation in slope of curve. It can be seen that for same engine speed and large S/B ratio IMEP is on higher side, this is due to reduction in heat loss from combustion chamber. At large B/S ratio (high speed engines) as speed increases the decrease in IMEP is small because of a shorter stroke decreases engine heat loss and friction, most noticeably at higher engine speed.

S/B 0.7 S/B 1.0 S/B 1.3

Figure 3.2 Effect of Stroke to Bore Ratio on IFSC Effect of Bore to Stroke ratio on IFSC is represented by Figure 3.2. As Engine speed increases the trend of IMEP is same for all S/B ratio but there is variation in slope of curve. It can be seen that for same engine speed and large S/B ratio IFSC is on lower side, this is due to reduction in heat loss from combustion chamber and less fuel burned for same power output. 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 NO emissions tended to decrease because of increasing crevice volume and decreasing temperature. The trend is more or less same for both SI and CI engine with small difference.

4 Compression Ratio The compression ratio is the ratio of the volume of the cylinder and the combustion chamber when the piston is at the bottom, and the volume of the combustion chamber when the piston is at the top. The compression ratio in a gasoline or petrol-powered engine will usually not be much higher than 10:1 due to potential engine knocking (detonation) and not lower than 6:1. Though there is limitation on highest CR, today’s technology pushed this limit further for gasoline engine. Mazda released new petrol engines under the brand name SkyActiv with a 14:1 compression ratio to be used in all Mazda vehicles by 2015. In a turbocharged or supercharged gasoline engine, the CR is customarily built at 10.5:1 or lower. This is due to the turbocharger/supercharger already having compressed the air considerably before it enters the cylinders. Port fuel injected engines typically run lower boost than direct fuel injected engines because port fuel injection allows the air/fuel mixture to be heated together which leads to detonation. Conversely, directly injected engines can run higher boost because heated air will not detonate without a fuel being present. In CI engines the heat of compression raises the temperature of the mixture to its autoignition point. The CR will customarily exceed 14:1 and ratios over 22:1 are common. Different methods to obtain different compression ratios are changing the cylinder head cavity volume, variation of combustion chamber height and variation of piston height. 1. Cylinder head cavity volume: The cylinder head cavity volume is plays major role in variation of compression ratio. This cylinder head cavity volume is measured separately for calculating the clearance volume. If cylinder head cavity volume is at higher side then compression ratio is at lower side and when cylinder head cavity volume is at lower side then compression ratio is at higher side. So every researcher aims to that keep compression ratio at higher side for better engine performance by using lower cavity volume cylinder head.

2.

Top dead center volume the top dead center volume is also important parameter which affecting on variation of compression ratio. This volume is measured when piston is rest at top dead center and this volume measured for calculating the clearance volume with the addition of cylinder head cavity volume. If TDC volume is at higher side then compression ratio is at lower side and when TDC volume is at lower side then compression ratio is at higher side. This top dead center volume always keeps at lower side for better engine performance.

3. Head gasket thickness: Head gasket thickness is little affecting on the variation in compression ratio. This gasket thickness measured for the calculating the clearance volume with the addition of cylinder head cavity volume and top dead center volume. For better engine performance the gasket thickness keep at lower side. 4. Piston Height from piston pin to crown: The piston height is little affecting on the variation in compression ratio. This piston height from piston pin to piston crown is helpful for the lowering clearance volume. If piston height is at higher side the TDC volume is at lower side and when piston height is lower side the TDC volume is higher side. For better engine performance keep piston height at higher side. The process of compressing a constant mass of gas into a much smaller space enables many more molecules to impinge per unit area on to the piston. When burning of the gas occurs, the chemical energy of combustion is rapidly transformed into heat energy which considerably increases the kinetic energy of the closely packed gas molecules. Therefore the extremely large number of molecules squeezed together will thus bombard the piston crown at much higher speeds. This then means that a very large number of repeated blows of considerable magnitude will strike the piston.

4.1 SI Engine Performance Theoretically, increasing the compression ratio of an engine can improve the thermal efficiency of the engine by producing more power output. The ideal theoretical cycle, the Otto cycle, upon which spark ignition (SI) engines are based, has a thermal efficiency, , which increases with compression ratio, and is given by η=1−

1 r

γ −1

Where, γ is ratio of specific heats (for air γ = 1.4) The ideal cycle analysis shows that indicated fuel conversion efficiency increases continuously with compression ratio. However, in an actual engine other processes which influence engine performance and efficiency vary with changes in compression ratio: namely, combustion rate and stability, heat transfer, and friction. Over the load and speed range, the relative impact that these processes have on power and efficiency varies also. Of course ability to increase the compression ratio is limited by the octane quality of available fuels and knock. Only a few studies have examined the effect of compression ratio on spark ignition engine performance and efficiency over a wide range of compression ratios.

Figure 4.3 Effect of Variation of CR on Brake Thermal Efficiency Figure 4.1 Effect of Variation of CR on Brake Thermal Efficiency

Figure 4.4 Effect of Variation of CR on IMEP and Thermal Efficiency Brake Mean effective pressures are shown for higher compression ratios efficiency and mep decrease slightly. This trend can be explained as being due to increasing surface/volume ratio and slower combustion, and is also due to the increasing importance of crevice volumes: at the higher compression ratios studied the combustion chamber height became very small. it is seen that the engine brake power increases as the compression ratio increases. This is due to the increase in brake torque at high compression ratios. Increase in compression ratio induces greater turning effect on the cylinder crank. That means that the engine is giving more push on the piston, and more torque is generated.

Figure 4.5 Effect of Variation of CR on IMEP and Thermal Efficiency

IMEP was achieved at higher compression ratios. As we increase the compression ratio more negative compression work must be done, but the IMEP still tends to increase. The increase in work-out at higher compression ratios overrides the additional required compression work. This is a result of the properties of the fluid and is best seen in examining thermal efficiency and the ideal Otto cycle. Thermal efficiency is work out divided by energyin. The energy-in (the product of mass of fuel and lower heating value) was held constant, and thus increasing thermal efficiency means the net work out must have increase as is seen by an increasing IMEP. It is expected that IMEP would tend to decrease after reaching a maximum due to increasing heat losses through the cylinder walls. As the surface area-to-volume ratio increases, greater amount of heat are conducted out of the cylinder. The loss of this thermal energy decreases the amount of work that can be extracted from the system.

There is no significant effect of increase in compression ratio on emission characteristic of SI engine. Compression ratio variation affects NOx emission little bit. The exhaust temperature decreases as compression ratio and efficiency increase until the compression ratio corresponding to maximum efficiency is reached. It has also been shown that heat losses to the combustion chamber walls, as a fraction of the fuel's chemical energy, also decrease as the compression ratio and efficiency both increase. The effect of compression ratio changes on NO emissions is small. Some studies show a modest increase in specific NO emissions as the compression ratio increases at constant load and speed; other studies show a slight decrease. Increasing the compression ratio increases exhaust hydrocarbon emissions. Several trends could contribute: increased importance of crevice volumes at high CR and lower gas temperatures during the latter part of the expansion stroke, thus producing less HC oxidation in the cylinder; decreasing residual gas fraction, thus increasing the fraction of in-

cylinder HC exhausted; lower exhaust temperatures, hence less oxidation in the exhaust system.

4.2 CI Engine Performance

Figure 4.6 Effect of Variation of CR on Brake Thermal Efficiency The maximum brake thermal efficiency is obtained at a compression ratio of 14.8. The least brake thermal efficiency is obtained at a compression ratio 20.2. Hence with respect to brake thermal efficiency, 14.8 can be treated as optimum. This can be attributed to the better combustion and better intermixing of the fuel and air at this compression ratio.

Figure 4.7 Effect of Variation of CR on Fuel Consumption The better fuel consumption was obtained at a compression ratio of 14.8. The high and low compression ratios than 14.8 result in high fuel consumptions. The fuel consumption at a compression ratio of 18.1 and 20.2 was almost the same. The high fuel consumption at higher compression ratios can be attributed to the effect of charge dilution. At the lower sides of the compression ratios, the fuel consumption is high due to incomplete combustion of the fuel.

The brake thermal efficiency of the engine working cycle is improved when CR rises, and firmly depends on the mechanical efficiency, which de creases when CR rises. However, in view of that fact, it is clear that the brake the mal efficiency depends on both the rate of i crease of indicated thermal efficiency and the rate of the de crease of mechanical efficiency. It is verified from the above that the brake the mal efficiency first rises at the beginning, then reaches the max i mal value for optimal CR value and then sub sequent declines. The CR value when the brake thermal efficiency reaches the max i mal value is the optimal value of CR for this load regime in engine operation.

BSFC (g/kWh) CR 17.5 CR 19

Figure 4.8 Effect of Variation of CR on Brake Specific Fuel Consumption Figure 4.6 shows that the variation of BSFC with injection timing for two different CRs. For given CR, BSFC reduces as injection timing is advanced before TDC. This is due to increase in heat release rate at TDC which generate maximum power without substantial pressure loss. For given injection timing, higher CR gives even more reduced BSFC than for lower CR. This is due to reason explained before.

. Figure 4.9 Effect of Variation of CR on NOx

In the case of the largest value of the CR (17.5) under all loads, largest temperature occurs inside the engine cylinder. Large amount of free oxygen under low-loads in spite of relatively low maximal temperature with respect to full load, leads to formation of the largest amount of NOx. Amount of NOx for that CR de creases with load increase. In the case of the lowest value of the CR (12.1) under low-loads, we have the lowest maximal temperature within the working cycle. This leads to formation of the lowest amount of NOx. With an increase in load, temperature increases as well and the amount of free oxygen decreases. Thus, at the beginning of the process, the amount of produced NOx increases, but, when the amount of free oxygen decreases, a decrease in the amount of produced NOx would occur with load increase.

Figure 4.10 Effect of Variation of CR on PM, CO, HC With the increase of CR and engine load, under the same injection timing (18.5 CAD BTDC), maximal cylinder pressure is increasing. This undesirable increase in maximal pressure is followed by a relatively improved atomizing of larger amount of fuel in cylinder under higher pressure and engine temperature. Because of improved conditions for

combustion process, the en tire working process is improved. Moreover, when the CR is increased, the temperature of exhaust gas is decreased. Leaner air-fuel mixture is used in engine operation under low-loads. Therefore, the amount of heat released during the combustion process is decreased. A consequence of this is certain decrease in temperature of the engine parts and decrease in cylinder temperature in the first phase of fuel injection. Under very low-loads, the degree of emission of PM is somewhat larger. The major reason for this is a relatively low injection pressure of the small amount of fuel that does not atomize so well. As the amount of fuel increases with a load increase, this effect is attenuated and a certain decrease in PM emission occurs, so that, under large loads, it would begin to increase again. Emission of PM increases under all loads with increase in the CR. The combustion chamber volume increases if the CR decreases. Thus, the amount of air in the cylinder increases, and it is the cause of decreasing of PM emission when the CR decreases. Poor fuel atomizing under low-loads leads to increase in the emission of CO, which is significantly decreased under the increased CR. On the other hand, the emission of CO is reduced under the improved quality of fuel atomizing, which improves with the increase in the amount of injected fuel. A similar case is with a change of the amount of HC.

5 Valve Timing and Valve Overlap

Figure 5.11 Standard Valve Timing for SI and CI engine

5.1 Effect of changes to Intake Valve Opening Timing IVO The opening of the intake valve allows air/fuel mixture to enter the cylinder from the intake manifold. (In the case of direct injection engines, only air enters the cylinder through the intake valve). The timing of IVO is the second parameter that defines the valve overlap and this is normally the dominant factor when considering which timing is appropriate for a given engine. Overlap will be discussed in more detail later in this paper. Opening the intake valve before TDC can result in exhaust gasses flowing into the intake manifold instead of leaving the cylinder through the exhaust valve. The resulting EGR will be detrimental to full load performance as it takes up space that could otherwise be taken by fresh charge. EGR may be beneficial at part load conditions in terms of efficiency and emissions as discussed above. Later intake valve opening can restrict the entry of air/fuel from the manifold and cause in-cylinder pressure to drop as the piston starts to descend after TDC. This can result in EGR if the exhaust valve is still open as gasses may be drawn back into the cylinder with the same implications discussed above. If the exhaust valve is closed, the delay of IVO tends not to be particularly significant, as it does not directly influence the amount of fresh charge trapped in the cylinder. Typical IVO timing is around 0-10° before TDC which results in the valve overlap being fairly symmetrical around TDC. This timing is generally set by full load optimization and, as such, is intended to avoid internal EGR.

Figure 5.12 Effect of IVO on Brake Power For the engine geometry and running conditions shown above, all parameters were kept constant except the IVO angle and the valve lift value. The IVO was varied from the original value 54o IVO angle BTDC opening down to 0o at TDC in steps for three valve lift values. As shown in Figure (1) the brake power is drawn versus the IVO angle opening BTDC for different valve lifts from (8.5,9 ,9.5,10 and 10.5 mm) at the engine design speed ( N = 2500 rpm). It showed an increase in power with the IVO angle reduction but this increase in power was small for values of IVO angle BTDC less than 25 o for all engine running speeds considered. The increase in power may be due to the reduction of residual gases and backflow of exhaust into the inlet manifold. The reduction of valve lift show a considerable reduction in power may be due to the restriction on charge gas inflow to the cylinder, may be due to the viscous effect of the attached jet formed by the low lift valve which is Reynolds number dependent, while the increase of valve lift shows an increase in power over the original 9.5 mm lift as they produce a larger effective area.

Figure 5.3 Effect of IVO on BSFC The variation of BSFC versus IVO angle BTDC; this showed that BSFC is hardly affected by IVO angle BTDC higher than 20 o for different valve lifts. But it shows an increase in BSFC for all valve lifts at IVO angles less than 20o.

5.2 Effect of changes to Intake Valve Closing Timing IVC The volumetric efficiency of any engine is heavily dependent on the timing of IVC at any given speed. The amount of fresh charge trapped in the cylinder is largely dictated by IVC and this will significantly affect engine performance and economy. For maximum torque, the intake valve should close at the point where the greatest mass of fresh air/fuel mixture can be trapped in the cylinder. Pressure waves in the intake system normally result in airflow into the cylinder after BDC and consequently, the optimum IVC timing changes considerably with engine speed. As engine speed increases, the optimum IVC timing moves further after BDC to gain maximum benefit from the intake pressure waves. Closing the intake valve either before or after the optimum timing for maximum torque results in a lower mass of air being trapped in the cylinder. Early intake closing reduces the mass of air able to flow into the cylinder whereas late intake closing allows air inside the cylinder to flow back into the intake manifold. In both cases, the part load efficiency can be improved due to a reduction in intake pumping losses. A typical timing for IVC is in the range of 50 0-60° after BDC and results from a compromise between high and low speed requirements. At low engine speeds, there will tend to be some flow back into the intake manifold just prior to IVC whereas at higher speeds, there may still be a positive airflow into the cylinder as the intake valve closes.

Figure 5.4 Effect of IVC on Brake Power Figure show the brake power drawn versus the IVC angle closing ABDC for valve lifts (8.5 ,9,9.5,10 and 10.5 mm) at the engine design speed ( N = 2500 rpm). It showed a decrease in power with the IVC angle reduction for all valve lifts considered. But it is less sever at higher valve lift. Late IVC reduce the volumetric efficiency. In contrast early IVC leads to greater reduction in volumetric efficiency, and this limits the output power. Late closing of the intake valve, long after the BDC, leads to a higher cylinder charge.

Figure 5.5 Effect of IVC on BSFC Figure shows the variation of BSFC versus IVC angle ABDC; this showed that BSFC is slightly affected by IVC angle ABDC, as it is increased slightly by reducing IVC angle ABDC.

5.3 Effects of Changes to Exhaust Valve Opening Timing EVO As the exhaust valve opens the pressure inside the cylinder resulting from combustion is allowed to escape into the exhaust system. In order to extract the maximum amount of work (hence efficiency) from the expansion of the gas in the cylinder, it would be desirable not to open the exhaust valve before the piston reaches Bottom Dead Centre (BDC). Unfortunately, it is also desirable for the pressure in the cylinder to drop to the lowest possible value, i.e. exhaust back pressure, before the piston starts to rise. This minimizes the work done by the piston in expelling the products of combustion (often referred to as blow down pumping work) prior to the intake of a fresh charge. These are two conflicting requirements, the first requiring EVO to be after BDC, the second requiring EVO to be before BDC. The choice of EVO timing is therefore a trade-off between the work lost by allowing the combusted gas to escape before it is fully expanded, and the work required raising the piston whilst the cylinder pressure is still above the exhaust back-pressure. With a conventional valve train, the valve lifts from its seat relatively slowly and provides a significant flow restriction for some time after it begins to lift and so valve lift tends to start some time before BDC. A typical EVO timing is in the region of 50 0-60° before BDC for a production engine. The ideal timing of EVO to optimize these effects changes with engine speed and load as does the pressure of the gasses inside the cylinder. At part load conditions, it is generally beneficial if EVO moves closer to BDC as the cylinder pressure is much closer to the exhaust back pressure and takes less time to escape through the valve. Conversely, full load operation tends to result in an earlier EVO requirement because of the time taken for the cylinder pressure to drop to the exhaust back-pressure.

Figure 5.6 Effect of EVO on Brake Power Figure shows the brake power versus the (EVO) angle opening BBDC for different engine speeds between (1000 - 5000 rpm). It shows a decrease in power with the (EVO) angle reduction for all engine running speeds. But it is less sever at lower engine speed (less than 1500 rpm), at low speeds, a late (EVO) reduce the volumetric efficiency η vol. In contrast at high engine speeds early (EVO) leads to greater reduction in volumetric efficiency, and this limits the output power.

Figure 3.7 Effect of EVO on BSFC The effect on BSFC is the opposite as shown in Figure. It shows the variation of BSFC versus (EVO) angle that BSFC is highly affected by (EVO) angle at high engine speeds, while it is increased slightly by reducing (EVO) angle at low engine speeds emissions.

5.4 Effects of Changes to Exhaust Valve Closing Timing EVC The timing of EVC has a very significant affect on how much of the Exhaust gas is left in the cylinder at the start of the engine’s intake stroke. EVC is also one of the parameters defining the valve overlap, which can also have a considerable affect on the contents of the cylinder at the start of the intake stroke. For full load operation, it is desirable for the minimum possible quantity of exhaust gas to be retained in the cylinder as this allows the maximum volume of fresh air & fuel to enter during the Intake stroke. This requires EVC to be at, or shortly after TDC. In engines where the exhaust system is fairly active (i.e. Pressure waves are generated by exhaust gas flow from the different cylinders), the timing of EVC influences whether pressure waves in the exhaust are acting to draw gas out of the cylinder or push gas back into the cylinder. The timing of any pressure waves changes with engine speed and so a fixed EVC timing tends to be optimized for one speed and can be a liability at others. For part load operation, it may be beneficial to retain some of the exhaust gasses, as this will tend to reduce the ability for the cylinder to intake fresh air & fuel. Retained exhaust gas thus reduces the need for the throttle plate to restrict the intake and results in lower pumping losses (see Appendix A) in the intake stroke. Moving EVC Timing further after TDC increases the level of internal EGR (Exhaust Gas Recirculation) with a corresponding reduction in exhaust emissions. There is a limit to how much EGR the cylinder can tolerate before combustion becomes unstable and this limit tends to become lower as engine load and hence charge density reduces. The rate of combustion becomes increasingly slow as the EGR level increases, up to the point where the process is no longer stable. Whilst the ratio of fuel to oxygen may remain constant, EGR reduces the proportion of the cylinder contents as a whole that is made up of these two constituents. It is this reduction in the ratio of combustible to inert cylinder contents which causes combustion instability.

Figure 5.8 Effect of EVC on Brake Power As shown in Figure 5.8 the brake power is drawn versus the EVC angle for different engine speeds between (1000 - 5000 rpm). It shows an increase in power with the (EVC) angle

reduction but this increase in power was small for values of (EVC) less than 25° for all engine running speeds considered. This effect is more recognized at higher engine speeds (2500 -5000 rpm). The increase in power may be due to the reduction of residual gases and backflow of exhaust into the inlet manifold, but a late (EVC) closing causes the high pressure exhaust gas reducing the amount of inlet mixture incoming through the inlet manifold. Retarded valve close angles induce a considerable reverse flow and results in the reduction of η vol and internal exhaust gas recirculation.

Figure 5.9 Effect of EVC on BSFC The BSFC is hardly affected by EVC angle higher than 25° for low engine speeds (less than 1500 rpm). But it was sensitive to (EVC) angle ATDC variation for higher engine speeds.

5.5 Effect of Overlap angle (IVO – EVC) Valve overlap is the time when both intake and exhaust valves are open. In simple terms, this provides an opportunity for the exhaust gas flow and intake flow to influence each other. Overlap can only be meaningfully assessed in conjunction with the pressure waves present in the intake and exhaust systems at any particular engine speed and load. In an ideal situation, the valve overlap should allow the departing exhaust gas to draw the fresh intake charge into the cylinder without any of the intake gas passing straight into the exhaust system. This allows the exhaust gas in the combustion chamber at TDC to be replaced and therefore the amount of intake charge to exceed that which could be drawn into the cylinder by the swept volume of the piston alone. A given amount of overlap unfortunately tends to be ideal for only a portion of engine speed and load conditions. Generally, the torque at higher engine speeds and loads can benefit from increased overlap due to pressure waves in the exhaust manifold aiding the intake of fresh charge. Large amounts of overlap tend to result in poor emissions at lower speeds as fuel from the intake charge can flow directly into the exhaust. High overlap can also result in EGR which, although beneficial to part load economy, reduces full load torque and can cause poor combustion stability especially under low load conditions such as idle. Poor idle quality can therefore result from too much overlap.

The valve overlap tends to be fairly symmetrical about TDC on most engines. The further away from TDC that valve overlap is present, the more effect the piston motion will have on the airflow. Early overlap may result in exhaust gasses being expelled into the intake manifold and late overlap may result in exhaust gasses being drawn back into the cylinder. Both of these situations result in internal EGR that can be beneficial to part load emissions and efficiency. As discussed earlier, internal EGR tends to be avoided due to the detrimental effect it has on full load torque

Figure 5.10 Effect of Valve Overlap on Brake Power For the engine geometry and running conditions shown above, all parameters were kept constant except the overlap angle between IVO and EVC at TDC and valve lifts. The former was varied from the original value 108o overlap angle at TDC down to 0o at TDC in steps while for the later three values were considered (8.5, 9,9.5,10 and 10.5 mm) . For the brake power drawn versus the overlap angle at TDC, for all valve lifts the power showed an increase with reducing overlap angle till about 60o angle then it starts a considerable decrease to lower values at 0o overlap. It has a larger reduction for less valve lifts.

Figure 5.11 Effect of EVO on Brake Power

The variation of BSFC versus overlap angle at TDC, this showed that BSFC decreases to a minimum at lower overlap angles at TDC then an increase toward the 0 o overlap angle. This is quite noticeable at low valve lifts while it approaches a minimum at around overlap angle around 60o.

6 Air-Fuel Ratio In spark ignition engine air and fuel is mixed before induction in cylinder. Mixing is done with the help of either carburetor or fuel injector. For complete combustion the amount of air required and corresponding Air- Fuel ratio is denoted by Stoichiometric A/F ratio. Often the A/F ratio is represented relative to stoichiometric A/F or in terms of Fuel-Air ratio represented by Equivalence ratio (ϕ). A/F ratio determines whether mixture is fuel rich or lean according to which performance of engine varies.

Figure 6.1 Effect of Equivalence ratio variation on IMEP

Power (kW)

Figure 6.2 Effect of Equivalence ratio variation on Power output

Equivalence ratio (relative F/A ratio) affects IMEP and as represented by above fig. General trend shows that indicated fuel conversion efficiency and mean effective pressure are

function of equivalence ratio. The IMEP peaks slightly rich of stoichiometric, about ϕ= 1.1. Due to dissociation at the high temperature following combustion, molecular oxygen is present in the burned gases under stoichiometric condition, so some additional fuel can be added and partially burned. This increases the temperature and the number of moles of burned gases in cylinder. These effects increase the pressure to give increased power and IMEP. Fuel conversion efficiency decreases with increase in equivalence ratio, as the mixture is richened above stoichiometric due to the decreasing combustion efficiency associated with the richening mixture. For mixtures lean of stoichiometric, the theoretical fuel conversion efficiency increases linearly as equivalence ratio decreases below 1.0. Combustion of mixtures leaner than stoichiometric produces products at lower temperature, and with less dissociation of the tri-atomic molecules CO2 and H2O. Thus the fraction of the chemical energy of the fuel which is released as sensible energy near TC is greater; hence a greater fraction of the fuel's energy is transferred as work to the piston during expansion, and the fraction of the fuel's available energy rejected to the exhaust system decreases.

Figure 6.3 Effect of Equivalence ratio variation on Thermal Efficiency and ISFC

BSFC (g/kWh)

Figure 6.4 Effect of Equivalence ratio variation on BSFC Figure 6.4 illustrates the Brake specific fuel consumption (bsfc) for this engine. The bsfc decreases as the mixture becomes leaner until the low supply of fuel energy creates poor combustion and misfire which Causes the bsfc to Increase. Interestingly, the bsfc reaches a minimum at 331g/kWh at ϕ=0.93 and then rapidly increases as is further decreased. This particular point is of interest when the emissions results are considered. On the other hand, since the efficiency is inversely proportional to the bsfc, brake thermal efficiency initially increases as the mixture becomes leaner. The fuel/air equivalence ratio is an important parameter controlling spark-ignition engine emissions. The critical factors affecting emissions that are governed by the equivalence ratio are the oxygen concentration and the temperature of the burned gases. Excess oxygen is available in the burned gases lean of stoichiometric. The maximum burned gas temperatures occur slightly rich of stoichiometric at the start of the expansion stroke, and at the stoichiometric composition at the end of expansion and during the exhaust process.

HC (ppm)

Figure 6.5 Effect of Equivalence ratio variation on HC emission

HC Emissions from Spark-Ignition Engines Unburned hydrocarbon levels in the exhaust of a spark-ignition engine under normal operating conditions are typically in the range 1000 to 3000 ppm. When combustion quality deteriorates, e.g. with very lean mixtures, HC emissions can rise rapidly due to incomplete combustion or misfire in a fraction of the engine's operating cycles. There are several mechanisms that contribute to total HC emissions. Also, any HC escaping the primary combustion process may oxidize in the expansion and exhaust processes. Four possible HC emissions formation mechanisms for spark-ignition engines (where the fuel-air mixture is essentially premixed) have been proposed: (1) flame quenching

at the combustion chamber walls, leaving a layer of unburned fuel-air mixture adjacent to the wall; (2) the filling of crevice volumes with unburned mixture which, since the flame quenches at the crevice entrance, escapes the primary combustion process; (3) absorption of fuel vapor into oil layers on the cylinder wall during intake and compression, followed by desorption of fuel vapor into the cylinder during expansion and exhaust; (4) incomplete combustion in a fraction of the engine's operating cycles (either partial burning or complete misfire). Figure 6.5 shows the effect of variations in equivalence ratio on hydrocarbon emissions in ppm. For rich mixtures, emissions of HC are high. This is primarily due to the lack of oxygen for afterburning of any unburned NC that escapes the primary combustion process within the cylinder and the exhaust system. The leaner: mixtures, which increase the oxygen concentration and yet produce adequate internal gas temperatures result in lower HC and provide lower until the lean operation limit is approached. The minimum of HC emission occurs at ϕ=0.8. Then HC emission rise again as combustion quality becomes poor.

bsCO (g/kWh)

Figure 6.6 Effect of Equivalence ratio variation on bsCO Carbon monoxide (CO) emissions from internal combustion engines are con-trolled primarily by the fuel/air equivalence ratio. For fuel-rich mixtures CO concentrations in the exhaust increase steadily with increasing equivalence ratio as the amount of excess fuel increases. For fuel-lean mixtures, CO concentrations in the exhaust vary little with equivalence ratio. Since spark-ignition engines often operate close to stoichiometric at part load and fuel rich at full load, CO emissions arc significant and must be controlled.

NO (ppm)

Figure 6.7 Effect of Equivalence ratio variation on NO emission Effect of variation in the Fuel-Air equivalence ratio on NOx formation is as follows. Maximum burned gas temperature occurs at 1.1; however, at this equivalence ratio oxygen concentrations are low. As the mixture is enriched, burned gas temperatures fall, As the mixture is leaned out, increasing oxygen concentration initially offsets the falling gas temperatures and NO emissions peak at ϕ=0.9. Detailed predictions of NO concentrations in the burned gases suggest that the concentration versus time histories under fuel-lean conditions are different in character from those for fuel-rich conditions. In lean mixtures NO concentrations freeze early in the expansion process and little NO decomposition occurs. In rich mixtures, substantial NO decomposition occurs from the peak concentrations present when the cylinder pressure is a maximum. Thus in lean mixtures, gas conditions at the time of peak pressure are especially significant. The formation rate of NO strongly depends on the gas temperature and oxygen concentration. The maximum exhaust gas temperature occurs ϕ= 0.9, and at this equivalence ratio the oxygen concentrations are also high. This combination maximizes NO emissions. As the mixture is leaned out with an increase in oxygen concentration, the decrease in burned gas temperature dominates the reaction. This effect results in lower NOx emissions throughout when the equivalence ratio is below 0.98. On the other hand as the mixture becomes richer even though the burned gas temperature is high the oxygen concentrations are lower which decreases NOx emissions.

7 Ignition Timing Ignition timing in a spark ignition engine is the process of setting the time that an ignition will occur in the combustion chamber (during the compression stroke) relative to piston position and crankshaft angular velocity. Setting the correct ignition timing is crucial in the performance and exhaust emissions of an engine.

Figure 7.1 Maximum Braking Torque (MBT)

If combustion starts too early in the cycle, the work transfer from the piston to the gases in the cylinder at the end of the compression stroke is too large: if combustion starts too late, the peak cylinder pressure is reduced and the expansion stroke work transfer from the gas to the piston decreases. There exists a particular spark timing which gives maximum engine torque at fixed speed, and mixture composition and flow rate. It is referred to as MBT— maximum brake torque—timing. This timing also gives maximum brake power and minimum brake specific fuel consumption. Fig .1 shows that at MBT, maximum torque is generated if ignition timing deviate from MBT timing significant reduction in torque occurs.

Figure 7.2 Effect of ignition timing advance on BMEP and IMEP Fig. 7.2 show that BMEP and IMEP tends to increase with ignition timing advance till 31° Before Top Dead Centre (BTDC) and then drop off. Best performance will be achieved with greatest ignition advance of 31° BTDC. It is expected that IMEP should increase with timing angle advance to a point, and thendrop off. Best performance will be achieved when the greatest portion of the combustion takes place near top dead center. If the ignition timing is not advanced enough, the piston will already be moving down when much of the combustion takes place. In this case we lose the ability to expand this portion of the gas through the full range, decreasing performance. If the ignition timing is too advanced, too much of the gas will burn while the piston is still rising. The work that must be done to compress this gas will decrease the net work produced. These competing effects cause there to be a maximum in the IMEP as a function of ignition timing advance The maximum BMEP is at an ignition timing 31°BTDC minimum advance for Maximum Brake Torque (MBT) is defined as the smallest advance that achieves 99 % of the maximum power.

Figure 7.3 Effect of ignition timing advance on BSFC

Figure 7.4 Effect of ignition timing advance on Heat Release Rate

Figure 7.5 Effect of ignition timing advance on Efficiency BSFC and Heat release curves can be analysed as follow. It can be seen that as ignition timing advanced from TDC the BSFC curve giving lower fuel consumption for same power output. One of the best advantage of optimum ignition timing is that as it gives maximum heat release rate at top dead centre which gives rise to maximum tempreture and pressure increasing power output. For figure 7.4 The blue curve is for maximum possible spark advance for effective performance and the spark timing retards towards TDC upto yellow curve. As a result the efficiency of the engine decreases with spark retard.

Figure 7.6 Effect of ignition timing advance on HC emission In the process of obtaining MBT spark timing the prssure inside the combution chamber rises rapidly at TDC. Higher pressure gives rise to more charge going in crevices and also more leakage from piston rings forming absorption and adsorption of it. Combustion tempressure later in power stroke decreases rapidly leads to incomplete combution and incresed HC emission at MBT.

Figure 7.7 Effect of ignition timing advance on Pressure variation

Figure 7.8 Effect of ignition timing advance on NOx emission Spark timing significantly affects NO emission levels. Fig. 7.7 shows peak cylinder pressure variation with spark timing. Advancing the timing so that combustion occurs earlier in the cycle increases the peak cylinder pressure as more fuel is burned before TC and the peak pressure moves closer to TC where the cylinder volume is smaller; retarding the timing decreases the peak cylinder pressure. Higher peak cylinder pressure and temperature gives higher NOx formation. Retarding ignition timing helps to reduce HC and NOx formation as the peak cylinder pressure and temperature decreases lowering NOx and post combustion temperature is increases gives rise to late oxidation of unburned HC. Though this reduces performance it is effective in emission control.

8 Load and Speed `Load and Speed One common way to present the operating characteristics of an internal combustion engine over its full load and speed range is to plot brake specific fuel consumption contours on a graph of brake mean effective pressure versus engine speed. Operation of the engine coupled to a dynamometer on a test stand, over its load and speed range, generates the torque and fuel flow-rate data from which such a performance map is derived. The upper envelope of the map is the wide-open-throttle performance curve. Points below this curve define the part-load operating characteristics, While details differ from one engine to another, the overall shapes of these maps for spark-ignition engines are remarkably similar. When mean piston speed Sp is used instead of crankshaft speed for the abscissa, the quantitative similarity of such maps over a wide range of engine sizes is more apparent. Maximum bmep occurs in the mid-speed range; the minimum bsfc island is located at a slightly lower speed and at part load. These map characteristics can be understood in terms of variations in volumetric efficiency, gross indicated fuel conversion efficiency and mechanical efficiency.

Figure 8.1 Effect of Engine speed on Performance

Figure 8.2 Torque and Power trade off

Figure 8.3 Engine Performance Characteristics

Figure 8.4 Effect of Speed and Load on BP, BSFC and Efficiency Increasing load at constant speed from the minimum bsfc point increases bsfc due to the mixture enrichment required to increase torque as the engine becomes increasingly airflow limited, Decreasing load at constant speed increases bsfc due to the increased magnitude of friction (due to increased pumping work), the increased relative importance of friction, and increasing importance of heat transfer. The effect of speed and load variation on NO and HC emission are can be elaborated as follows. NO concentration increase moderately with increasing speed at constant load. At lower loads, the proportional increase in NO is greater than at higher loads. The residual gas fraction decreases as speed increases, this effect being greater at lower inlet manifold pressures (lighter loads). Also, the relative importance of heat transfer per cycle is less as speed increases , which would also be expected to increase NO concentration. With increasing load (at constant speed), NO concentrations also increase. Again, as inlet manifold pressure and load increase, the residual gas fraction decreases also, the relative importance of heat transfer per cycle decreases with increasing load.

The hydrocarbon concentration trends with speed and load changes are the opposite of the NO concentration trends. As indicated, speed and load are likely to affect several of the HC formation mechanisms, the in-cylinder mixing of unburned hydrocarbons which escape combustion with the bulk gases, and the fraction of the in-cylinder HC which escape into the exhaust. However, not enough is yet known about the details of these processes to make these dependencies explicit. If oxygen is available, oxidation of unburned hydrocarbons both within the cylinder and in the exhaust system will be significantly enhanced by increases in speed since the expansion stroke and exhaust process gas temperatures increase substantially, due to the reduced significance of heat transfer per cycle with increasing speed. This more than offsets the reduced residence time in the cylinder and in the exhaust. As load increases at constant speed, expansion and exhaust stroke temperatures increase, and the in-cylinder oxidation rate, if oxygen is available, will increase. However, as the exhaust gas flow rate increases, the residence time in critical sections of the exhaust system decreases and a reduction in exhaust port HC oxidation occurs. The net trend is for HC concentration to decrease modestly as load is increased.

9 Injection Timing and Injection Pressure Fuel-injection timing essentially controls the crank angle at which combustion starts. While the state of the air into which the fuel is injected changes as injection timing is varied, and thus ignition delay will vary. The fuel-injection rate, fuel nozzle design (including number of holes), and fuel-injection pressure all affect the characteristics of the diesel fuel spray and its mixing with air in the combustion chamber. At normal engine conditions (low to medium speed, fully warmed engine) the minimum delay occurs with the start of injection at about 10 to 15° BTDC. The increase in the delay with earlier or later injection timing occurs because the air temperature and pressure change significantly close to TDC. If injection starts earlier, the initial air temperature and pressure are lower so the delay will increase. If injection starts later (closer to TDC) the temperature and pressure are initially slightly higher but then decrease as the delay proceeds.

IP-1040 bar IP-1215 bar IP- 1683 bar

Figure 9.1 Effect of injection Timing and Pressure on BSFC Figure 9.1 and 9.2 shows effect of injection timing and pressure on BSFC. For given injection pressure BSFC decreases with advanced injection timing because as injection timing is advanced, time span for complete mixing and atomization of fuel increases and charge burns at high pressure and temperature closer to TDC position hence work transfer increases. For given injection timing increase in injection pressure gives lower BSFC due to formation of small liquid droplets increasing surface area giving instant atomization of fuel giving complete combustion.

BSFC (g/kWh) IP-400bar IP-550bar IP-700bar

Figure 9.2 Effect of injection Timing and Pressure on BSFC

Nox (g/kWh)

Figure 9.3 Effect of injection Timing on NOx Figure 9.3 shows effect of injection timing on NOx emission. As injection timing is advanced NOx formation increases. As explained earlier advanced injection gives peak pressure and temperature at TDC which is favorable condition for NOx formation. Retarded injection reduces peak temperature and hence NOx.

PM (g/kWh)

Figure 9.4 Effect of injection Timing on PM Figure 9.4 shows effect of injection timing on particulate matter emission. Advancing injection decreases PM due to complete atomization and mixing giving complete combustion reducing PM.

NOx (g/kWh) IP-400bar IP-550bar IP-700bar

Figure 9.5 Effect of injection Timing and Pressure on NOx Figure 9.5 shows effect of injection pressure on NOx. Higher injection pressure reduces time required for atomization and hence giving favorable condition for NOx formation.

PM (g/kWh) IP-400bar IP-550bar IP-700bar

Figure 9.6 Effect of injection Timing and Pressure on PM Injection timing variations have a strong effect on NOx emissions for DI engines. Retarded injection is commonly used to help control NOx emissions. It gives substantial reductions, initially with only modest bsfc penalty. Retarding timing generally increases smoke, though trends vary significantly between different types and designs of diesel engine. Mass particulate emissions increase as injection is retarded. The injection rate depends on injection pressure. Higher injection rates result in higher fuel-air mixing rates, and hence higher heat-release rates for a given amount of fuel injected per cylinder per cycle, as the injection pressure is increased the optimum injection timing moves closer to TDC. The effects of injection pressure and timing on bsfc in a natural aspirated DI diesel engine are shown. The higher heat-release rates and shorter overall combustion process that result from the increased injection rate decrease the minimum bsfc at optimum injection timing: however, a limit to these benefits is eventually reached. Increasing

the injection rate increases NOx emissions and decreases smoke or particulate emissions. The engine designer's goal is obviously to achieve the best bsfc possible.

10 EGR In internal combustion engines, exhaust gas recirculation (EGR) is a nitrogen oxide (NOx) emissions reduction technique used in petrol/gasoline and diesel engines. EGR works by recirculating a portion of an engine's exhaust gas back to the engine cylinders. This dilutes the O2 in the incoming air stream and provides gases inert to combustion to act as absorbents of combustion heat to reduce peak in-cylinder temperatures. NOx is produced in a narrow band of high cylinder temperatures and pressures. EGR acts as diluents to the combustion mixture. Introduction of EGR is to reduce oxygen concentration. Increase specific heat of incoming charge which ultimately reduces peak combustion temperature and hence NOx reduction.

10.1SI Engine Performance

Figure 10.1 Effect of EGR on Performance The effect of exhaust gas recycle on engine performance and efficiency, for mixtures with ϕ<1, is similar to the addition of excess air. Both EGR and excess air dilute the unburned mixture. In practice since EGR is only used at part-throttle conditions, ϕ<1, is the region of interest. Because three-way catalysts are now used where NOx emission constraints are severe, greatest attention has focused on dilution with EGR at ϕ=1. At constant burn duration, bsfc and exhaust temperature decrease with increasing EGR. Only for very long combustion processes is the burn rate especially significant_ This improvement in fuel consumption with increasing EGR is due to three factors :(1) reduced pumping work as EGR is increased at constant brake load (fuel and air flows remain almost constant; hence intake pressure increases); (2) reduced heat loss to the walls because the burned gas temperature is decreased significantly; and (3) a reduction in the degree of dissociation in the high-temperature burned

gases which allows more of the fuel's chemical energy to be converted to sensible energy near TDC.

Figure 10.2 Effect of EGR on Emission These burned gases are comprised of both residual gas from the previous cycle and exhaust gas. Since the burned gases dilute the unburned mixture, the absolute temperature reached after combustion varies inversely with the burned gas mass fraction. Hence increasing the burned gas fraction reduces the rate of formation of NOx emissions. Effect on NOx emissions of increasing the burned gas fraction by recycling exhaust gases to the intake system substantial reductions in NOx concentrations are achieved with 10 to 25 percent EGR. The amount of EGR a particular combustion chamber design will tolerate depends on its combustion characteristics, the speed and load, and the equivalence ratio. Faster-burning engines will tolerate more EGR than slower-burning engines Because of the decrease in burn rate and increase in cycle-by-cycle combustion variations, Hydrocarbon emissions increase with increasing EGR. At first the increase in HG is modest and is due primarily to decreased HC burn up due to lower expansion and exhaust stroke temperatures. The HC increase becomes more rapid as slow combustion, partial burning, and even misfire, in turn, occur with increasing frequency EGR has no significant effect on engine CO emissions.

10.2CI Engine Performance CI engine subjected to high pressure and temperature combustion and also CI engine mostly run in lean condition its favorable condition hence NOx formation is considerable in this process. One possible method to reduce is reduce combustion temperature by recirculating exhaust gas diluting charge. But in this process formation of PM in CI engine increases. Figure 10.3 shows how NOx and PM trade occurs with EGR rate. At about 20% EGR the NOx and PM are at lowest possible in tradeoff. Increased PM can be filtered out with PM trap hence EGR is most effective method for NOx control.

Figure 10.3 NOx-Soot Tradeoff for different EGR rate

Figure 10.4 Effect of EGR on Indicated Power Figure 10.4 shows effect of EGR on Indicated Power. It can be seen that for part load 15% EGR is superior to 10 or 20% EGR rate.

Figure 10.5 Effect of EGR on SFC Figure 10.5 shows effect of EGR on specific fuel consumption. It indicates the variations of brake specific fuel consumption with increasing EGR rate. There is remarkable improvement in fuel consumption with increasing EGR. One of the main reason for that effects is due to the reduction of pumping work as the amount of EGR rate is increased(with

fuel and air flow rate remains constant),the pump work get reduced and hence the entire inlet charge needing to come passed the throttle. Again due to the reduction in heat loss to the wall of cylinder the significant reduction in burnt gas reduction, improve the fuel consumption trends. The reductions in degree of dissociation in high temperature burn gases also improve specific fuel consumption.

Figure 10.6 Variation of NOx with Hot and Cold EGR NOx emission from hot EGR is comparatively higher than without EGR. Cold EGR of higher rates shows much effective in reducing NOx emission. at 10% cold EGR percentages is very high than that of higher EGR rates.CO emissions with EGR was increased in part loads and decreases with higher loads as compared without EGR.

11 Swirl Changes in swirl rate change the fuel evaporation and fuel air mixing processes. They also affect wall heat transfer during compression and, hence, the charge temperature at injection. At normal operating engine speeds, the effect of swirl rate change on the delay is small. Under engine starting conditions (low engine speeds and compression temperatures) the effect is much more important presumably due to the higher rates of evaporation and mixing obtained with swirl. There are different methods for swirl generation. Modification of manifold is one such method to produce swirl in combustion chamber. Making manifold spiral, helical and combination of both produces induced swirl, which affect the combustion characteristic as follow.

Figure 11.1 Effect of Swirl on BSFC The bsfc is a measure of engine efficiency. In fact, bsfc and engine efficiency are inversely related, so that the lower the bsfc the better the engine. Engineers use the bsfc rather than thermal efficiency because a more are less universally accepted definition of thermal efficiency does not exist. The variation of brake specific fuel consumption at different load for normal, spiral, helical and helical-spiral inlet manifolds is shown in figure. 4 Brake specific fuel consumption of different inlet manifolds are very similar to normal manifold. BSFC increases with load up to 0.5kW, however as load further increases from 0.5 to 3 kW. It can be observed from the Figure.4 that brake specific fuel consumption for all new technique manifolds is less compared to normal manifold. It is significant to note that 32.03% of reduced in bsfc is observed at 2.5kW load for helical-spiral inlet manifold compared to normal inlet manifold.

Figure 11.2 Effect of Swirl on BMEP Figure 11.2 depicts the variation of Brake mean effective pressure with respect to at different loads .The brake mean effective pressure is the indication of external shaft work per unit displacement volume done by the engine. Brake mean effective pressures were higher for new intake manifold technique than normal intake manifold The increase in brake mean effective pressure may increase the power output and decrease the exhaust emissions. It is significant to note that 81.17kN/m2 of incresed in Brake Mean Effective Pressure observed at 2.5kW load for helical-spiral inlet manifold compared to normal inlet manifold.

Figure 11.3 Effect of Swirl on CO emission

The variation of carbon monoxide with respect to load can be observed that as the load increases the CO emission is increased.CO emissions are low at low load and high at full load for normal manifold compared to other manifolds. It can be observed that CO emissions are decreased in case of helical-spiral manifold up to a load of 2kW. The reason behind increased

CO emission may be incomplete combustion. The maximum CO emission was observed at the full load 3kW.

Figure 11.4 Effect of Swirl on HC emission Figure 11.4 depicts the variation of hydrocarbons with respect to load for tested different inlet manifold. Unburned hydrocarbon emissions are caused by incomplete combustion of fuel air mixture. HC emissions varies from no load to full load Unburned hydrocarbons are higher in case of spiral manifold compared to normal manifold but in case of helical and helical-spiral manifold will be less. The values of unburned hydrocarbons of spiral, helical and helical-spiral manifolds for constant speed at 2.5kw load are 46, 24 and 22 ppm as compared to 27 ppm of normal manifold. The probable reason for emission may be some portion of the fuel-air mixture in the combustion chamber comes into direct contact with combustion chamber wall and get quenched and some of this quenched fuel-air mixture is forced out during the exhaust which contributes to the high HC emission.

Figure 11.5 Effect of Swirl on NOx emission

Figure 11.5 depicts the Oxide of nitrogen from the engine exhaust at different loads. NOx results from reaction of nitrogen and oxides at relatively high temperature. No is major component in the NOx emission .As the load increases the oxides of nitrogen emission increases .The oxides of nitrogen were higher for spiral and helical manifold at lower loads, as the load increases the emissions were less for all new technique inlet manifold compared to normal manifold.

12 Cetane Number Cetane number (CN) is a measure of the ignition quality of the diesel fuel and is determined by a standard engine test as specified by ASTM. The ignition quality is quantified by measuring the ignition delay, which is the period between the time of injection and the start of combustion (ignition) of the fuel. A fuel with a high CN has a short ignition delay period and starts to combust shortly after it is injected into an engine. Ignition delay is the time interval between the start of fuel injection and the beginning of the oxidation reaction. Ignition delay period starts with the injection of fuel and consists of physical and chemical delay periods until the auto ignition occurs. Fuels with a high CN have a very short ignition delay time; that is, ignition occurs in a very brief interval of time after injection begins. Conversely, the longer the ignition delay time the lower the CN. The ignition delay time of diesel cycle engines is a fundamental parameter to effectively control the combustion process, allowing for high thermal efficiency through maximum pressures close to 15° after reaching the top dead center (TDC), with which the maximum torque characteristic of Diesel cycle engines is obtained.

Figure 12.1 Effect of CN on BSFC Figure represents the effect of CN on brake specific fuel consumption (bsfc) for the four tested fuels. Increasing fuel CN reduces bsfc, although it is still high at low loads. Increasing fuel’s CN improves combustion and raises combustion chamber temperatures. Increasing combustion chamber temperatures gives low fuel delay period, and gives better ignition. Reducing the load reduces temperatures inside combustion chamber, and increases fuel delay period, resulting in bad combustion that needs more fuel to compensate the lost power.

Figure 12.2 Effect of CN on Brake Power Brake power (BP) increased with increasing engine speed, as Fig. 6 illustrates. Increasing CN increases BP also. Brake power increased by 1.1, 3.88 and 5.6% for CN 50, 52 and 55 respectively compared with baseline diesel fuel (CN=48.5).

Figure 12.3 Effect of CN on NOx Exhaust gas temperatures are increased by increasing load, while increasing CN reduces these temperatures, as shown in Fig. 5. Increasing load needs more fuel to be burned which rises exhaust temperatures. On the other hand, increasing CN improves delay period, making the burning process to be completed at top dead given the operating conditions, it is easy to see why cetane level is important. In addition to improving fuel combustion, increasing cetane level also tends to reduce emissions of nitrogen oxides (NOx). These emissions tend to be more pronounced when working with lower cetane number fuels as Figure shows. The decrease in CN caused an increase in NO, because of the long ignition delay

Figure 12.4 Effect of CN on CO Fig.12.4 shows the variation of the CO concentration in exhaust gas with variable engine loads, when the engine was operated on commercial diesel fuel of 48.5 CN, and modified fuel of 50, 52 and 55 CN diesel fuels. Carbon monoxide is the primary intermediate product in the hydrocarbon oxidation. The presence of CO in lean fuel- air mixtures exhaust is an indication that some of the CO produced through the oxidation reactions could not be oxidized further to carbon dioxide. With very lean engine operation and small load within the partial motoring region, the CO concentrations recorded imply that they also partially originate from the incomplete combustion. These emissions are reduced with increasing CN in the fuel by 11.79, 31.2 and 56.34 for CN 50, 52 and 55 respectively compared with baseline diesel fuel (CN=48.5).

Figure 12.5 Effect of CN on HC At higher loads, when the diesel concentration in the cylinder charge is high enough, the UBHC tend to reduce. Diesel fuel with CN= 55 improved the utilization of the fuel up to 20% compared with baseline diesel. Also, it reduced the UBHC concentration in the exhaust,

as compared to the 48.5 CN fuel. The CN 50 fuel had a slightly adverse effect. Little differences were found at very light loads, as well as at full load.

13 Future Scope In this seminar report we have studied several parameters that affect performance of IC engines. Future demand of emission reduction and improvement in fuel consumption leads to more in depth study with the help of some software assistance is going to be of much importance. Following are the some topics on which further research can be done. 1. 2. 3. 4.

Modeling of different processes with the help of computer. Modeling the performance to match real-time conditions Study of multiple parameters simultaneously theoretically and experimentally. Optimization of parameters for better performance using different methods e.g. Taguchi Method, Neural network method.

14 Conclusion 1. Several parameters have been studied and effect of them on IC engine performance and cause of performance variation is discussed. 2. Bore to Stroke ratio as surface to volume ratio should be taken considering optimum between heat loss effect and volumetric efficiency due to valve size. 3. Though higher compression ratio enhances performance by improving efficiency with little expense of increase of Nox emission, knocking phenomenon limit the large CR use. 4. Valve opening and Closing and also valve overlap is critical for volumetric efficiency and hence performance and emission. 5. Air-Fuel ratio is very crucial in performance and emission as it affects every parameter, little lean mixture gives better performance. 6. In SI engine 200 to 300 advance of ignition timing BTDC gives lower emission and Max Brake Torque. 7. In CI engine optimum injection timing (100 to 150 BTDC) and high injection pressure gives high BSFC and low emission. 8. Use of EGR is better option for NOx control. Increased swirl rate improves combustion characteristic and hence improved performance. Higher Cetane Number reduces delay period and gives better combustion property.

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