Chapter 10 Internal Combustion Engine Systems – Thermal Engineering

Chapter 10

Internal Combustion Engine Systems


An engine may be defined as a device which converts one form of energy into mechanical energy. Mechanical energy can be further easily converted into electrical energy which is easier to transport and control. Heat engine is a cyclic device which transforms heat energy into mechanical energy.

Heat engines are classified into the following:

  1. External combustion engines
  2. Internal combustion engines

In an external combustion engine, the fuel is burned outside the engine and the generated heat is supplied to the working fluid of the engine for power generation. Steam engine is an example of an external combustion engine in which fuel is burned in the boiler to generate steam, which is used in the engine for power generation. The working fluid is not mixed with fuel, and the same working fluid (water in the form of steam) is repeatedly used in the system.

In an internal combustion (IC) engine, the fuel is mixed with air and burned inside the engine to generate power. In this case, the same working fluid (air-fuel mixture) cannot be used again in the cycle. Some examples of internal combustion engine are spark ignition (SI) engine, compression ignition (CI) engine, etc.

IC engines offer the following advantages over external combustion engines:

  1. Greater mechanical simplicity
  2. Lower weight to output ratio
  3. Higher overall thermal efficiency
  4. Less water requirement
  5. Easy and quick starting

Figure 10.1 shows the classification of IC engines on the basis of cycle of operation into cylinder, type of fuel, method of supply of fuel, type of ignition, among others.

Figure 10.1 IC engine classification

  1. Basic engine design: Reciprocating engines, rotary (Wankel) engines
  2. Working cycle: Otto cycle and diesel cycle
  3. Number of strokes: Four-stroke and two-stroke engines
  4. Fuel: Gasoline (or petrol), compressed natural gas (CNG), liquefied petroleum gas (LPG), diesel oil (light diesel oil—LDO and high speed diesel oil—HSD), fuel oil, alcohols (methanol, ethanol)

    Figure 10.2 Classification of IC engines as per cylinder arrangement: (a) Inline engine, (b) Vee engine, (c) Radial engine, (d) Opposed piston engine, (e) Opposed cylinder engine, (f) Delta type engine

  5. Fuel supply and mixture preparation: Carburetted type and injection type
  6. Method of ignition: CI (compression-ignition), SI (spark-ignition) engines: battery ignition or magneto ignition
  7. Method of cooling: Water cooled or air-cooled
  8. Cylinder arrangement (Fig. 10.2): Inline, V, or Vee, radial, opposed piston and opposed cylinder, delta
  9. Valve or port design and location: Overhead valves (I-head), side valve (L-head); in two-stroke engines: cross scavenging, loop scavenging, uniform scavenging
  10. Application: Automotive engines for land transport, marine engines for propulsion of ships, aircraft engines for aircraft propulsion, industrial engines, prime movers for electrical generators
  1. Four-stroke spark-ignition engine: The cross section of a four-stroke spark-ignition engine is shown in Fig. 10.3. The major components of the engine are as follows:
    1. Cylinder block: It is the main supporting structure for the various components. The cylinder head is mounted on the cylinder block. The bottom of the cylinder block is called the crank case. The lubricating oil is kept in the crank case sump.
    2. Cylinder: It is a cylindrical vessel in which the piston reciprocates.

      Figure 10.3 Cross-section of a spark-ignition engine

    3. Piston: It is a cylindrical component fitted into the cylinder forming the moving boundary of the combustion system.
    4. Combustion chamber: It is the space between the cylinder and the piston top where combustion takes place.
    5. Inlet manifold: It is the pipe through which air or air-fuel mixture is drawn in the cylinder.
    6. Exhaust manifold: It is the pipe through which the products of combustion escape into the atmosphere.
    7. Inlet and exhaust valves: The valves are used for regulating the incoming charge into the cylinder (inlet valve) or discharging the products of combustion (exhaust valve) from the cylinder.
    8. Spark plug: It is a component to initiate the combustion process in SI engine.
    9. Carburettor: It is used for mixing fuel and air in correct proportion in SI engine.
    10. Connecting rod: It is used to interconnect the piston and the crank to transmit force from the piston to the crankshaft.
    11. Crankshaft: It is used to convert reciprocating motion of the piston into rotary motion of the output shaft.
    12. Piston rings: They provide a tight seal between the piston and the cylinder wall.
    13. Gudgeon pin: It connects the small end of the connecting rod to the piston.
    14. Camshaft: It is a shaft on which cams are mounted to operate the valves. It is driven from the cam shaft by gears.
    15. Flywheel: It absorbs surplus energy during working stroke and delivers during idle stroke.
  2. Four-stroke compression-ignition engine: Except for the spark plug and carburettor, all other components are the same as for the SI engine. It also requires a fuel pump and fuel nozzle. Figure 10.4 shows outline of a CI engine.

    Figure 10.4 Outline of diesel engine

  3. Two-stroke SI engine: In the case of two-stroke engine, the valves are replaced by the exhaust port and the transfer port. The charging of the cylinder with the air-fuel mixture takes place through the carburetor and ignition by a spark plug. Other components are same as for a four-stroke engine.
  4. Two-stroke CI engine: Here, the carburettor and the spark plug are replaced by a fuel pump and fuel nozzle. All other components are same as for two-stroke SI engine.

10.4.1 Four-stroke Spark-ignition Engine

The details of various processes of a four-stroke spark-ignition engine with overhead valves are shown in Fig. 10.5. Within the four strokes, there are five events to be completed namely, suction, compression, combustion, expansion, and exhaust.

Figure 10.5 Working principle of a four-stroke SI engine: (a) Suction stroke, (b) Compression stroke, (c) Expansion or power stroke, (d) Exhaust stroke

  1. Suction stroke: It starts when the piston is at the top dead centre (TDC) and is about to move downwards. The inlet valve is open and the exhaust valve is closed at this time. The charge consisting of the fuel-air-mixture is drawn into the cylinder. When the piston reaches the bottom dead centre (BDC), the suction stroke ends and the inlet valve closes.
  2. Compression stroke: The charge is compressed to the clearance volume by the return stroke of the piston with both inlet and exhaust valves are closed. At the end of the compression stroke, the mixture is ignited with the help of spark plug to convert chemical energy of fuel to heat energy.
  3. Expansion stroke: The high pressure of burnt gases forces the piston towards BDC with both the valves in closed position to produce power.
  4. Exhaust stroke: At the end of the expansion stroke, the exhaust valve opens with the inlet valve closed. The piston starts moving towards the TDC and sweeps the burnt gases out from the cylinder. The exhaust valve closes when piston has reached TDC.

10.4.2 Four-stroke Compression-ignition Engine

The four cycles of operation of a CI engine are shown in Fig. 10.6.

  1. Suction stroke: Only air is induced during the suction stroke with the inlet valve open and the exhaust valve closed.

    Figure 10.6 Cycle of operation of a four-stroke CI engine: (a) Suction, (b) Compression, (c) Expansion, (d) Exhaust

  2. Compression stroke: The sucked air is compressed into the clearance volume with both valves closed.
  3. Expansion stroke: Fuel injection starts nearly at the end of the compression stroke, resulting in combustion. The products of combustion expand and both the valves remain closed.
  4. Exhaust stroke: The piston moves from BDC to TDC and pushes out the products of combustion with the exhaust valve open and intake valve closed.

10.4.3 Two-stroke Spark-ignition Engine

In a two-stroke engine, the cycle is completed in one revolution of the crankshaft. The filling process is accomplished by the charge compressed in the crankcase or by a blower. The induction of the compressed charge moves out the products of combustion through the exhaust ports. Therefore, two piston strokes are required for these two operations. Two strokes—one for compressing the fresh charge and the other for expansion or power stroke—are sufficient to complete the cycle. Figure 10.7 shows the crankcase scavenged two-stroke engine.

10.4.4 Two-stroke Compression-ignition Engine

Figure 10.8 shows a two-stroke compression-ignition engine. It is a crankcase scavenged type engine. S is the plate valve for admission of air in the crank case, E is the exhaust port, and A is the port in the cylinder communicating to the crankcase through the cylinder block casting; cams and valves are not required. The four operations—air induction, air compression and fuel injection, expansion, and exhaust are completed in two strokes, that is, in one revolution of crankshaft.

In the upward motion of the piston, suction is created in the crankcase and air enters through the plate valves for full 180° of crank rotation as shown in Fig. 10.8(a). Above the piston, the compression starts after both the points have been covered by the piston. At the end of compression, fuel is injected and ignites at TDC giving products of combustion at high pressure.

On downward stroke, the high pressure products of combustion expand and the air below the piston compresses, closing the plate valves, as shown in Fig. 10.8(b). As soon as the piston uncovers the exhaust port, the products of combustion are released into the atmosphere. A little later, on the downward stroke, the other port communicating with the crankcase gets uncovered and the air compressed in the crankcase gets transferred to the space above the piston as shown in Fig. 10.8(c).

Figure 10.7 Crankcase scavenged two-stroke SI engine

Figure 10.8 Working of two-stroke C.I. Engine


Table 10.1 shows a comparison of four-stroke and two-stroke engines.


Table 10.1 Comparison of two-stroke and four-stroke engines



Table 10.2 gives the comparison of SI and CI engines.


Table 10.2 Comparison of SI and CI engines


10.7.1 Merits

  1. A two-stroke engine gives twice as many power strokes as a four-stroke cycle engine at the same engine speed; therefore, a two-stroke engine of the same size should develop twice the power of a four-stroke engine. In practice, the actual power developed by a two-stroke engine is about 1.7 to 1.8 times the power developed by a four-stroke engine of the same dimensions and speed. This happens because some of the power is used for compressing the charge in the crank case and the effective stroke is less than the actual stroke.
  2. For the same power developed, the two-stroke engine is much lighter, less bulky, and occupies less floor area. Therefore, it is more suitable for uses in marine engines and transport purposes.
  3. It provides mechanical simplicity as valves, rocker arms, push-rods, cam, and cam shafts are not required. The friction loss is also less, and therefore, it gives higher mechanical efficiency.
  4. The two-stroke engines are much easier to start.
  5. A crankcase compression and valve-less type two-stroke engine can run in either direction, which is useful in marine applications.
  6. The initial cost of the engine is considerable less.
  7. The weight/kW ratio is considerably less.

10.7.2 Demerits

  1. The thermodynamic efficiency of an engine is only dependent on the compression ratio. The effective compression ratio for a two-stroke engine is less than that for four-stroke engine for the same stroke (actual) and clearance volume. Therefore, the thermodynamic efficiency of two-stroke cycle is always less than a four-stroke cycle engine.
  2. The actual efficiency of a two-stroke cycle is less than a four-stroke cycle engine because greater overlapping of ports is necessary in a two-stroke engine for effective scavenging. A portion of fresh charge in the case of an SI engine always escapes unused through the exhaust ports; therefore, the specific fuel consumption is usually higher.
  3. As the power-strokes per minute are twice the power stroke of four-stroke cycle engines, the capacity of the cooling system used must be higher. The cooling of the engine also presents difficulty as the quantity of heat removed per minute is large. Due to firing in each revolution, the piston is likely to get overheated and oil cooling of the piston is necessary.
  4. The consumption of lubricating oil is sufficiently large because of high operating temperatures.
  5. Sudden release of the gases makes the exhaust noisier.
  6. A two-stroke petrol engine with crankcase compression 50–60% of the swept volume is filled with fresh charge, whereas a four-stroke petrol engine contains 80–95%. This happens because the space occupied by the rotating parts in the crankcase prevents a full charge being sucked in.
  7. The scavenging is not complete, particularly in high speed engines, as very short time is available for exhaust; hence, the fresh charge is highly polluted. This can be reduced using an opposed piston two-stroke diesel engine which provides unidirectional scavenging.
  8. The turning moment of a two-stroke engine is more non-uniform as against with four-stroke engine, so it requires heavier flywheel and strong foundation.

10.8.1 Four-stroke SI Engine

The valve timing diagram shows the regulation of the positions in the cycle at which the valves are set to open and close. The valve timing diagrams for low and high speed four-stroke S.I. engine are shown in Fig. 10.9. Typical valve timings are given in Table 10.3.

10.8.2 Four-stroke CI Engine

A typical valve timing diagram for a four-stroke CI engine is shown in Fig. 10.10. The typical timing valves are also displayed.

IVO up to 30° before TDC; IVC up to 40° after BDC

EVO about 45° before BDC; EVC about 30° after TDC

Fuel valve opens (FVO) about 15° before TDC; FVC about 25° after TDC.

10.8.3 Two-stroke SI Engine

The transfer port opens 35° before the BDC and closes 35° after the BDC. The exhaust port opens 45° before the BDC and closes 45° after the BDC. The spark ignition occurs 20° before the TDC. The typical valve timing diagram for two-stroke engine is shown in Fig. 10.11.

Figure 10.9 Valve timing for low and high speed four-stroke SI engine: (a) Low speed engines, (b) High speed engines


Table 10.3 Typical valve timings for four-stroke SI engines

Figure 10.10 Valve timing diagram four-stroke CI engine

Figure 10.11 Typical valve timing diagram of a two-stroke SI engine

10.8.4 Two-stroke CI Engine

The transfer port opens 45° before the BDC and closes 45° after the BDC. The exhaust port opens 60° before the BDC and closes 60° after the BDC. The fuel valve opens 15° before the TDC and closes 20° after the TDC. The typical valve timing diagram is shown in Fig. 10.12.

Figure 10.12 Valve timing diagram for a two-stroke CI engine


At the end of the expansion stroke, the combustion chamber of a two-stroke engine is left full of products of combustion as there is no exhaust stroke available to clear the cylinder of burnt gases. The process of clearing the cylinder, after the expansion stroke, is called scavenging process. The scavenging process is the replacement of the products of combustion in the cylinder from the previous power stroke with fresh-air charge to be burned in the next cycle.

There are three types of scavenging systems as follows:

  1. Uniflow scavenging system: In this system, as shown in Figs 10.13(a) and (b), air enters the cylinder from one end and leaves from the other end. Air acts like an ideal piston and pushes out the residual gas in the cylinder and replaces it with fresh charge. Due to absence of any eddies or turbulence, in a uni-flow scavenging system, it is easier to push the products of combustion out of the cylinder without mixing with it and short circuiting. Thus, this system has the highest scavenging efficiency.
  2. Cross scavenging: This process shown in Fig. 10.14. It employs inlet and exhaust ports placed in the opposite sides of the cylinder wall. The air moves up to combustion chamber on one side of the cylinder and then down on the other side to flow out of the exhaust ports. This process requires that air should be guided by the use of either a suitably shaped detector formed on piston top or by the use of inclined ports. The main disadvantage of this system is that the scavenging air is not able to get rid of the layer of exhaust gas near the wall, resulting in poor scavenging. Some of the fresh charge also goes directly in the exhaust port.

    Figure 10.13 Uni-flow scavenging system: (a) Exhaust valve, (b) Opposed piston

    Figure 10.14 Cross scavenging

    Figure 10.15 Loop or reverse scavenging

  3. Loop or Reverse Scavenging: Figures 10.15(a) and (b) show the loop or reverse scavenging system. This avoids the short-circuiting of the cross-scavenged engine and improves upon the scavenging efficiency. The inlet and exhaust ports are placed on the same side of the cylinder wall. The major mechanical problem with this system is that of obtaining an adequate oil supply to the cylinder wall consistent with reasonable lubricating oil consumption and cylinder wear.

The detail applications with capacity and type of engine used are listed in Table 10.4.


Table 10.4 Applications of IC engines



10.11.1 Four-stroke Petrol Engine

The theoretical and actual p-v diagrams for a four-stroke petrol engine are shown in Figs. 10.16(a) and (b), respectively.

The theoretical p-v diagram is drawn with the following assumptions:

  1. Suction and exhaust take place at atmospheric pressure through 180° rotation of crank.
  2. Compression and expansion take place through 180° rotation of crank.
  3. Compression and expansion processes are isentropic.
  4. The combustion takes place instantaneously at constant volume at the end of compression stroke.
  5. Pressure suddenly falls to the atmospheric pressure at the end of expansion stroke.

The various processes of theoretical p-v diagram are as follows:

5-1: Suction stroke (pa = const)

1-2: Compression stroke (pvγ = const)

Figure 10.16 Theoretical and actual p-v diagrams for a four-stroke petrol engine: (a) Theoretical, (b) Actual

2-3: Instantaneous combustion (v = const)

3-4: Expansion stroke (pvγ = const)

4-1: Sudden fall in pressure (v = const)

1-5: Exhaust stroke (pa = const)

In practice, the actual conditions differ from the ideal as follows:

  1. The suction of mixture in the cylinder is possible only if the pressure inside the cylinder is below atmospheric pressure.
  2. The burnt gases can be pushed out into the atmosphere only if the pressure of the exhaust gases is above atmospheric pressure.
  3. The compression and expansion do not follow the isentropic law, as there will be heat exchange during these processes.
  4. Sudden pressure rise is not possible after the ignition as combustion takes some time for completion and actual pressure rise is less than theoretical considered. The pressure increase takes place through some crank rotation, or increase in volume.
  5. Sudden pressure release after the opening of expansion valve is not possible and it also takes place through some crank rotation.

If all these modifications are taken into account, then the cycle can be represented on p-v diagrams as shown in Fig. 10.16(b).

The area 4′-5-1-4′ representing negative work is called negative loop or pumping loop. This work is required for admitting the fresh charge and for exhausting the burnt gases. This loss of work is known as pumping loss and power consumed for this is known as pumping power.

The net work per cycle of the engine is given by the area (A1 A2). This area (A1 A2) is always less than the area A as shown in Fig. 10.16(a) due to the actual deviations of operations from the theoretical ones.

If both the areas are represented in the form of rectangles taking vs as base, then the ordinates give the mean effective pressures as shown in Fig. 10.17.

pma (Actual mean effective pressure) = pmt (theoretical mean effective pressure) × DF (diagram factor)

10.11.2 Four-stroke Diesel Engine

The theoretical and actual p-v diagrams for a four-stroke diesel engine are shown in Figs. 10.18(a) and (b), respectively.

Figure 10.17 p-v diagrams

Figure 10.18 Theoretical and actual p-v diagrams for a four-stroke diesel engine: (a) Theoretical, (b) Actual

The theoretical p-v diagram for a four-stroke diesel engine is drawn with the following assumptions:

  1. Suction and exhaust take place at atmospheric pressure through 180° of crank rotation.
  2. Compression and expansion take place during 180° of crank rotation.
  3. Compression and expansion are isentropic.
  4. The combustion takes place at constant pressure during a small part of expansion stroke.
  5. Pressure suddenly falls to atmospheric pressure at the end of expansion stroke.

With the above assumptions, the working cycle can be represented on a p-v diagram as shown in Fig. 10.18(a), and it is similar to the theoretical diesel cycle. The various processes of theoretical p-v diagram are as follows:

5-1: Suction stroke (pa = const)

1-2: Compression stroke (pvγ = const)

2-3: Constant pressure combustion (p = const)

3-4: Expansion stroke (pvγ = const)

4-1: Sudden fall in pressure (v = const)

1-5: Exhaust stroke (pa = const)

However, in practice, the actual conditions differ from the ideal described as follows:

  1. The suction of the air inside the cylinder is possible only if the pressure inside the cylinder is below atmospheric.
  2. Exhausting of gasses is possible only if the pressure of the exhaust gases is above atmospheric pressure.
  3. The compression and expansion do not follow the isentropic process, as there are heat and pressure losses.
  4. The combustion at constant pressure is not possible as the fuel will not burn as it is introduced into the cylinder.
  5. The sudden pressure release after the opening of expansion valve is not possible and it takes place through some crank rotation.

The operations of the cycle, taking the modifications into account, are represented on the p-v diagram in Fig. 10.18(b). Actual area (A1 A2) on p-v diagram per cycle is less than theoretical.

10.11.3 Two-stroke Petrol Engine

The theoretical and actual p-v diagrams for a two-stroke petrol engine are shown in Fig. 10.19. The following assumptions are made for drawing the theoretical p-v diagram:

  1. The expansion during power stroke and compression during compression stroke are isentropic.
  2. The combustion takes place instantaneously at constant volume at the end of compression.
  3. The pressure falls instantaneously to the atmospheric pressure as the piston uncovers the exhaust ports during power stroke.
  4. The scavenging takes place at atmospheric pressure.

It may be observed from Fig. 10.19(a) that the compression of the charge starts from ‘1’ instead of point 6.

Effective compression ratio,

where vse = effective stroke volume =

vsa = actual stroke volume =

Figure 10.19 Theoretical and actual p-v diagrams for a two-stroke petrol engine: (a) Theoretical, (b) Actual

Le, La = effective and actual stroke lengths, respectively

d = cylinder diameter

The various processes are as follows:

1-2: isentropic compression (pvγ = c)

2-3: instantaneous combustion (v = c)

3-4: isentropic expansion (pvγ = c)

4-1: release of burned charge to atmosphere (v = c)

1-6 and 6-1: sweeping out of exhaust gases to atmosphere

Point 5: inlet port opens

5-6 and 6-5: charging the cylinder with fresh charge and scavenging action. Actual theoretical mean effective pressure,

pma, pme = actual mep on the basis of actual and effective stroke respectively. In practice, the actual conditions differ from the ideal as described below:

  1. The compression and expansion processes do not follow the isentropic law strictly.
  2. Instantaneous combustion at the end of compression is not possible. The actual pressure rise takes place through some crank angle, resulting in rounding of the diagram.
  3. The scavenging always takes place above atmospheric pressure.
  4. Instantaneous fall of pressure at the time of release is not possible.

10.11.4 Two-stroke Diesel Engine

The theoretical and actual p-v diagrams for a two-stroke diesel engine are shown in Fig. 10.20. The various processes are as follows:

1-2: isentropic compression (pvy = c)

2-3: instantaneous combustion (p = c)

3-4: isentropic expansion (pvy = c)

4-1: release of burned charge to atmosphere (v = c)

1-6 and 6-1: sweeping out of exhaust gases to atmosphere

5-6 and 6-5: charging the cylinder with fresh charge and scavenging action.

Point 5: inlet port opens.

Figure 10.20 Theoretical and actual p-v diagrams for a two-stroke diesel engine: (a) Theoretical, (b) Actual


In SI engines, the air and fuel are mixed outside the engine cylinder and partly evaporated mixture is supplied to the engine. The process of preparing this mixture is called carburetion. The device used for this purpose is known as carburettor. The carburettor atomises the fuel and mixes it with air. This complicated process is achieved in the induction system, which is shown in Fig. 10.21. The pipe that carries the prepared mixture to the engine cylinder is called the intake manifold.

During the suction stroke, vacuum is created in the cylinder which causes the air to flow through the carburettor and the fuel to be sprayed from the fuel jets. Due to high volatility of the SI engine fuels, most of the fuel vaporises and forms a combustible fuel-air mixture. However, some of the larger droplets may reach the cylinder in the liquid form and must be vapourised and mixed with air during the compression stroke before ignition takes place by the electric spark.

Figure 10.21 Induction system for SI engine


Table 10.5 F:A ratios for various running conditions of SI engines

The fuel-air ratios for various running conditions of SI engines are given in Table 10.5.

10.12.1 Simple Carburettor

The details of a simple carburettor are shown in Fig. 10.22. It consists of a float chamber, fuel discharge nozzle and a metering orifice, a venturi, a throttle valve, and a choke. A float and a needle valve system maintains a constant level of gasoline in the float chamber. If the amount of fuel in the float chamber falls below the designed level, the float goes down, thereby opening the fuel supply valve and admitting fuel. When the designed level is reached, the float closes the fuel supply valve, thus stopping additional fuel flow from the supply valve and stopping additional fuel flow from the supply system. The float chamber is vented either to the atmosphere or to the upstream side of the venturi.

During suction stroke, air is drawn through the venturi (or choke tube). As the air passes through the venturi, the velocity increases, reaching a maximum at the venturi throat and pressure decreases to a minimum. From the float chamber, the fuel is fed to a discharge jet, the tip of which is located in the throat of the venturi, fuel is discharged into the air stream. To avoid overflow of fuel through the jet, the level of the liquid in the float chamber is maintained at a level slightly below the tip of the discharge jet.

The throttle valve controls the amount of charge delivered to the cylinder, which is situated after the venturi tube. As the throttle is closed, less air flows through the venturi tube and less quantity of the air-fuel mixture is delivered to the cylinder; hence, the power output is reduced. The reverse takes place when the throttle is opened.

10.12.2 Compensating Jet

The function of the compensating jet is to make the mixture leaner as the throttle opens progressively. As shown in Fig. 10.23, the compensating jet is connected to the compensating well which is vented to the atmosphere. The compensating well is supplied with fuel from the main float chamber through a restricting orifice. With the increase in air flow rate, there is a decrease in the fuel level in the compensating well, with the result that fuel supply through the compensating jet decreases. The compensating jet, thus, progressively makes the mixture leaner as the main jet progressively makes the mixture richer.

Figure 10.22 Simple carburettor

Figure 10.23 A compensating jet device

10.12.3 Theory of Simple Carburettor

The air from the atmosphere is sucked through the carburettor by the pressure difference across it created when the piston moves on its suction stroke. The velocity of the air increases as it passes through the venturi and reaches maximum at venturi throat (Fig. 10.24). The pressure also changes and is maximum at section 2-2, because this is the minimum area in the induction track. The fuel is sucked through the nozzle because of suction created in the venturi.

Figure 10.24 Principles of a simple carburettor

  1. Approximate analysis (neglecting compressibility of air): Let z be the height in metre of fuel nozzle tip higher than the float chamber level. Assuming initial velocity of air to be negligible (c1 = 0), density of air to be constant and considering sections at entrance and venturi throat, by applying, Bernoulli’s theorem, we get

    where p1, p2 = pressure at sections 1 and 2, respectively, N/m2

    c2 = velocity of air at section 2, m/s

    ρa = density of air, kg/m3

    Mass of air per second,

    where A2 = area of venturi throat, m2

    Similarly, for the flow of fuel, we have

    where ρf = constant density of fuel, kg/m3

    cf = velocity of flow of fuel, m/s

    Mass flow rate of fuel,

    where Af = cross-sectional area of fuel nozzle, m2

    Taking the coefficient of discharge of fuel nozzle and venturi into account, we have

    where Cd = coefficient of discharge.

    If z = 0.

  2. Exact analysis: Considering compressibility of air and applying steady flow energy equation to sections 1 and 2, we get

    where h1, h2 = enthalpies at sections 1 and 2 respectively

    Since Q = 0, W = 0 and c1 = 0

    For isentropic flow between the atmosphere and venturi throat, we have

    where v = specific volume


    Eq. (10.7) becomes

10.12.4 Limitations of Single Jet Carburettor

The drawbacks of single jet carburettor are as follows:

  1. A single jet carburettor cannot provide a very rich mixture as required at the time of starting the engine. This is because at low speed (starting or idling), the pressure difference causing the fuel flow is very small as the throttle is nearly closed. It is not possible to discharge fuel to make the mixture considerably rich.
  2. It cannot provide a very rich mixture required for sudden acceleration of the engine.
  3. For gradually increasing pressure difference over the jet (at higher speed of the engine), the weight of the petrol discharged from a single jet increases at a greater rate than does the air supply. Hence, a single jet carburettor gives a progressively richer mixture as the air speed increases when set to given a correct mixture at low air speeds.
  4. It cannot reduce the quantity of air flow during starting as required in cold weather conditions.
  5. The automatic control of air and fuel according to the required conditions is not possible.

The carburettor used with a variable speed engine must fulfil all the following requirements:

  1. It must atomise the fuel and mix it homogeneously with the air.
  2. It must be able to run the engine smoothly without hunting or fuel wastage.
  3. It must provide rich mixture during starting and idling.
  4. It must provide a constant air-fuel ratio during normal running of the engine which is the maximum period.
  5. It must provide a rich mixture required for quick acceleration of the engine.
  6. It must be able to start the engine even in very cold weather conditions (during snowfall).
  7. All operations should be automatic.

To fulfil the above requirements, the following devices are introduced.

  1. A starting or a pilot jet (to start engine)
  2. Compensating devices (to provide constant A:F ratio during normal operation conditions)
  3. An automatic control of choke valve (to start the engine in cold weather)

10.12.5 Different Devices Used to Meet the Requirements of an Ideal Carburettor

The main function of the carburettor is to vapourise the petrol in the current of air created by means of engine suction and supply the required quantity of air and petrol mixture in proper proportion in accordance with the load on the engine and its speed.

In the design of a carburettor, the following points should be considered:

  1. It is seen that for a gradually increasing pressure difference over the fuel jet, the mass of petrol discharged from a single jet increases at a greater rate than does the air supply. Hence, a carburettor of this type will give a progressively rich mixture as air speed is increased. There must be some device to maintain A: F ratio constant over a wide range of engine speed.
  2. To ensure rapid and complete combustion, it is necessary that the fuel should be finely divided and intimately mixed with air supply. This can be done by proper design of venturi and inlet manifold.
  3. For starting and accelerating, a rich mixture must be supplied momentarily but the supply should come to the correct mixture strength automatically when the engine attains the desired speed.
  4. There must be a provision to supply extra-rich mixture to start the engine in a very cold weather (where temperature falls to zero or sub-zero temperature).
  5. The float in the float chamber of a carburettor must maintain the fuel level constant irrespective of load or speed of the engine.

10.12.6 Complete Carburettor

In order to satisfy the demands of an engine under all conditions of operations, the following additional systems are added to the simple carburettor:

  1. Main metering system
  2. Idling system
  3. Power enrichment by economiszer system
  4. Acceleration pump systems
  5. Choke.

Main Metering System

The main metering system of a carburettor is designed to supply a nearly constant basis fuel-air ratio over a wide range of speeds and loads. This mixture corresponds approximately to best economy at full throttle (A/F ratio ≈ 15.6 or F/A ratio 0.064). Since a simple or elementary carburettor tends to enrich the mixture at higher speeds automatic compensating devices are incorporated in the main metering system to correct this tendency. These devices are as follows:

  1. Use of a compensating jet that allows an increasing flow of air through a fuel passage as the mixture flow increases.
  2. Use of emulsion tube for air bleeding. In this device, the emphasis is on air bleeding alone.
  3. Use of a tapered metering pin that is moved in and out of the main or auxiliary fuel orifice either manually or by means of some automatic mechanism changing the quantity of fuel drawn into the air charge.
  4. Back-suction control or pressure reduction in the float chamber.
  5. Changing the position or jet in the venturi. The suction action is highest at the venturi throat, therefore by raising the venturi the nozzle relatively moves to points with smaller suction and the flow of fuel is decreased.
  6. Use of an auxiliary air valve or port that automatically admits additional air as mixture flow increases.

The main devices are explained in detail.

  1. Compensating jet device: This device is shown in Fig. 10.23. In this device, in addition to the main jet, a compensating jet is provided which is in communication with a compensating well.

    The compensating well is open to atmosphere and gets its fuel supply from the float chamber through a restricting orifice. As the air flow increases, the level of fuel in the well decreases, thus reducing the fuel supply through the compensating jet. The compensating jet thus tends towards leanness as the main jet tends towards richness, the sum of the two remaining constant as shown in Fig. 10.25. At even higher rates of air flow, when the compensating jet has been emptied, air is bled through the compensating jet to continue the leanness effect, and incidentally, to assist in fuel atomisation.

  2. Emulsion tube or air bleeding device: In modern carburettors, the mixture correction is done only by air bleeding. In this arrangement, the main metering jet is fitted about 25 mm below the petrol level and it is called a submerged jet (see Fig. 10.26). The jet is situated at the bottom of a well, the sides of which have holes which are in communication with the atmosphere. Air is drawn through the holes in the well, the petrol is emulsified, and the pressure difference across the petrol column is not as great as that in the simple or elementary carburettor. Initially, the petrol in the well is at a level equal to that in the float chamber. On opening the throttle this petrol, being subject to the low throat pressure, is drawn into the air. This continues with decreasing mixture richness as the boles in the central tube are progressively uncovered. Normal flow takes place from the main jet.

    Figure 10.25 Variation of air-fuel ratio vs air flow with main and compensating jet

    Figure 10.26 Correction in modern carburettors by air bleeding

  3. Back-suction control or pressure reduction method: A common method of changing the air-fuel ratio in large carburettors is the back-suction control as shown in Fig. 10.27. In this arrangement, a relatively large vent line connects the carburettor entrance (say point 1) with the top of the float chamber. Another line, containing a very small orifice line, connects the top of the float chamber with the venturi throat (say point 2). A control valve is placed in the large vent line. When the valve is wide open, the vent line is unrestricted the pressure in the float chamber equal to p1, and the pressure difference acting on the fuel orifice is (p1 p2). If the valve is closed, the float chamber communicates only with the venturi throat and the pressure on the fuel surface will be p2. Then ∆ pf will be zero, and no fuel will flow. By adjusting the control valve, any pressure between p1 and p2 may be obtained in the float chamber, thus changing the quantity of fuel discharged by the nozzle.

    Figure 10.27 Back-suction control or pressure reduction method

  4. Auxiliary valve carburettor: An auxiliary valve carburettor is illustrated in Fig. 10.28. With an increase of engine load, the vacuum at the venturi throat also increases. This causes the valve spring to lift the valve admitting additional air and the mixture is prevented from becoming over-rich.
  5. Auxiliary part carburettor: An auxiliary port carburettor is illustrated in Fig. 10.29. By opening the butterfly valve, additional air is admitted and at the same time the depression at the venturi throat is reduced, decreasing the quantity of fuel drawn in. This method is used in aircraft carburettors for altitude compensation.

    Figure 10.28 An auxiliary valve carburettor

    Figure 10.29 An auxiliary port carburettor

Idling System

It has already been shown that at idling and low lead, the engine requires a rich mixture (about A/F 12:1). However, the main metering system not only fails to enrich the mixture at low air flows but also supplies no fuel at all at idling. For this reason, a separate idling jet must be added to the basic carburettor. An example of idling jet is shown in Fig. 10.30. It consists of a small fuel line from the float chamber to a little on the engine side of the throttle. This line contains a fixed fuel orifice. When the throttle is practically closed, the full manifold suction operates on the outlet to this jet. In addition, the very high velocity past be throttle plate increases the suction locally. Fuel can, therefore, be lifted by the additional height up to the discharge point, but this occurs only at very low rates of air flow. As the throttle is opened, the main jet gradually takes over while the idle jet becomes ineffective. The desired air-fuel ratio for the idling jet is regulated manually to idle adjust, which is a needle valve controlling the air bleed.

Power Enrichment of Economiser System

As the maximum power range of operation (75% to 100% load) is approached, some device must allow richer mixture (A/F about 13: 1, F/A 0.08) to be supplied despite the compensating jet. Such a device is the meter rod economiser shown in Fig. 10.31. The name economiser is rather misleading. It stems from the fact that such a device provides a rich uneconomical mixture at high load demand without interfering with economical operation in the normal power range. The meter rod economiser shown in Fig. 10.31, simply provides a large orifice opening to the main jet as the throttle is opened beyond a certain point.

The rod may be tapered or stepped. Other examples provide for the opening of auxiliary jets through some linkage to the throttle movement or through a spring action when manifold vacuum is lost as the throttle is opened.

Acceleration Pump System

It has already been shown that when it is desired to accelerate the engine rapidly, a simple carburettor will not provide the required rich mixture. Rapid opening of the throttle will be immediately followed by an increased airflow, but the inertia of the liquid fuel will cause at least a momentarily lean mixture just when richness is desired for power. To overcome this deficiency, an acceleration pump is provided; an example is shown in Fig. 10.32. The pump consists of a spring-loaded plunger. A linkage mechanism is provided so that when the throttle is rapidly opened the plunger moves into the cylinder and forces an additional jet of fuel into the venturi. The plunger is raised again against the spring force when the throttle is partly closed. Arrangement is provided so that when the throttle is opened slowly, the fuel in the pump cylinder is not forced into the venturi but leaks past plunger or some holes into the float chamber.

Figure 10.30 Idling jet

Figure 10.31 Meter rod economiser

Instead of the mechanical linkage shown, some carburettors have a plunger held up by a manifold vacuum. Wherever that vacuum is reduced by the rapid opening of the throttle, a spring forces the plunger down in a pumping action identical to that of the pump illustrated.


During cold starting period, at low cranking speed and before the engine has warmed up, a mixture much richer than usual mixtures (almost five to 10 times more fuel) must be supplied because a large fraction of the fuel will remain liquid even in the cylinder, and only the vapour fraction can provide a combustible mixture with the air. The most common means of obtaining this rich mixture is by the use of a choke, which is a butterfly-type valve placed between the entrance to the carburettor and the venturi throat as shown in Fig. 10.33. By partially closing the choke, a large pressure drop can be produced at the venturi throat that would normally result from the amount of air flowing through the venturi. This strong suction at the throat will draw large quantities of fuel from the main nozzle and supply a sufficiently rich mixture so that the ratio of evaporated fuel to air in the cylinders is within combustible limits. Choke valves are sometimes made with a spring loaded so that high-pressure drops and excessive choking will not result after the engine has started and has attained a higher speed.

Figure 10.32 Acceleration pump

Figure 10.33 Choke valve with spring-loaded bypass

Some manufacturers make the choke operate automatically by means of a thermostat such that when the engine is cold, the choke is closed by a bimetallic element. After starting and as the engine warms up, the bimetallic element gradually opens the choke to its fully open position.

An alternative to the choke is the provision of auxiliary fuel jets that are opened manually or automatically only as required.

Example 10.1

A single jet carburettor has to supply 5 kg of air per minute. The air is at a pressure of 1.013 bar and at a temperature of 27°C. Calculate the throat diameter of the choke for air flow velocity of 90 m/s. Take the velocity coefficient as 0.8. Assume the flow to be incompressible and isentropic.

[IES, 1992]


Given that a = 5 kg/min, p1 = 1.013 bar, T1 = 273 + 27 = 300 K, c2 = 90 m/s, Cda = 0.8

Applying Bernoulli’s theorem between inlet and throat, we have


Density of air at inlet, ρa1 =

Density of air at throat, ρa2 =

Mass flow rate at throat, a = ρa2 A2c2

throat dia, d2 = 0.0325 m or 32.5 mm

Example 10.2

A six-cylinder, four-stroke engine, 80 mm in diameter and 120 mm stroke length runs at 3600 rpm. The volumetric efficiency of the engine is 0.8. If the maximum head causing the flow is limited to 11.765 cm of mercury, find the throat diameter of the venturi required. Find the diameter of the nozzle orifice if the desired A: F ratio is 15:1. Take Cda = 0.9, Cdf = 0.7, ρa = 1.3 kg/m3 and ρf = 720 kg/m3.


Given that i = 6, d = 0.08 m, L = 0.120 m, N = 3600 rpm, ηvol. = 0.8, hHg = 11.765 cm

Volume flow rate of air,

Mass flow rate of air, a = ρaa = 1.3 × 0.0868 = 0.113 kg/s


hw = hHg × γHg = 11.765 13.6 = 160 of water

Example 10.3

The throat diameter of a carburettor is 80 mm and the nozzle diameter is 5.5 mm.The nozzle lip is 6 mm. The pressure difference causing the flow is 0.1 bar. Find (a) the air-fuel (A:F) ratio supplied by the carburettor neglecting nozzle lip. (b) A:F ratio considering nozzle lip and (c) the minimum velocity of air required to start the fluid flow. Neglect air compressibility. Take Cda = 0.85, Cdf = 0.7, ρa = 1.2 kg/m3 and ρf =750 kg/m3.


Given that da = 80 mm, df = 5.5 mm, z = 6 mm, ∆p = 0.1 bar, Cda = 0.85, Cdf = 0.7, ρa = 1.2 kg/m3, ρf = 750 kg/m3

Example 10.4

A 4-stroke petrol engine of 2 litres capacity is to develop maximum power at 4000 rpm. The volumetric efficiency at this speed is 0.75 and the air-fuel ratio is 14:1. The venturi throat diameter is 28 mm. The coefficient of discharge of venturi is 0.85 and that for fuel jet is 0.65. Calculate (a) the diameter of the fuel jet and (b) the air velocity at the throat.

The specific gravity of petrol is 0.76. Atmospheric pressure and temperature are 1 bar and 17°C respectively.

[IAS, 2002]


Given: Vs = 2 litres or 2 × 10−3 m3, N = 4000 rpm, n = 2, ηvol = 0.75, A/F = 14/1, d2 = 28 mm, Cda = 0.85, Cdf = 0.65, ρf = 760 kg/m3, p1 = 1 bar, T1 = 17 + 273 = 290 K

  1. or df = 28 × 0.32733 = 9.165 mm

    ∴ Diameter of fuel jet = 9.165 mm

  2. Volume of air supplied per cycle = stroke volume × ηvol
    V1 = 2 × 10−3 × 0.75 = 1.5 × 10−3 m3

    Mass of this air per cycle at suction condition,

    = 1.8 × 10−3 kg/s

    Mass of air supplied,

    a = A2 C2 ρa2     (ρa1 = ρa2)

    or c2 = 81.2 m/s

Example 10.5

The venturi of a simple carburettor has a throat diameter of 20 mm and the fuel orifice has a diameter of 1.12 mm. The level of petrol surface in the float chamber is 6.0 mm below the throat venturi. Coefficient of discharge for venturi and fuel orifice are 0.85 and 0.78 respectively. Specific gravity of petrol is 0.75. Calculate (a) the air-fuel ratio for a pressure drop of 0.08 bar, (b) petrol consumption in kg/hr and (c) the critical air velocity. The intake conditions are 1.0 bar and 17°C. For air cp = 1.005 and cv = 0.718 kJ/kg-K.

[IAS, 2003]


Given: d2 = 20 mm, df = 1.12 mm, z = 6 m, Cda = 0.85, Cdf = 0.78, sf = 0.75, ∆ p = 0.08 bar, p1 = 1.0 bar, T1 = 290 K, cp = 1.005 kJ/kg.K, cv = 0.718 kJ/kg.K

  1. Fuel-air ratio,
  2. Petrol consumption,
  3. Critical air velocity

Example 10.6

The air-fuel ratio of a mixture supplied to an engine by a carburettor is 15. The fuel consumption of the engine is 7.5 kg/h. The diameter of the venturi is 2.2 cm. Find the diameter of fuel nozzle if the lip of the nozzle is 4 mm. Take ρf = 750 kg/m3, Cda = 0.82, Cdf = 0.7, atmospheric pressure = 1.013 bar and temperature = 25°C. Neglect compressibility effect of air.


Given that A: F = 15:1, f = 7.5 kg/h, da = 2.2 cm, z = 4 mm, ρf = 750 kg/m3, Cda = 0.82, Cdf = 0.7, p = 1.013 bar, T = 273 + 25 = 298 K

Density of air,

Mass flow rate of air,

Example 10.7

A four-stroke petrol engine of 1710 cm3 capacity is to be designed to develop maximum power at 5400 rpm, when the mixture A:F ratio supplied to the engine is 13:1. Two carburettors are to be fixed. The maximum velocity of air is limited to 107 m/s. Find the diameters of venturi and fuel nozzle jet, if ηvol (engine) = 0.7, Cda = 0.85, Cdf = 0.66, nozzle lip = 6 mm and ρf = 750 kg/m3. Atmospheric pressure and temperature are 1.013 bar and 300 K. Take cp = 1 kJ/kgK.


Given that Ve = 1710 cm3, N = 5400 rpm, A:F = 13, ca = 107 m/s, ηvol = 0.7, Cda = 0.85, Cdf = 0.66, z = 6 mm, ρf = 750 kg/m3, p = 1.013 bar, T = 300 K, cp = 1 kJ/kgK

Volume of air supplied to engine at STP,

Air flow through each carburettor,

Mass flow rate of air,

Air velocity through venturi,

Assuming the flow to be isentropic,

Volume of air passing through each carburettor,

2 = AacaCda

or da = 2 cm

Mass flow rate of fuel per carburettor,

Example 10.8

A 4-stroke petrol engine has a swept volume of 2.0 litres and is running at 4000 rpm. The volumetric efficiency at this speed is 75% and the air-fuel ratio is 14:1. The venturi throat diameter of the carburettor fitted to the engine is 30 mm. Estimate the air velocity at the throat if the discharge coefficient for air is 0.9. The ambient conditions are, pressure = 1.0 bar, temperature = 20°C. Calculate the diameter of the fuel jet if the fuel density is 760 kg/m3. For air cp = 1.005 kJ/kgK and R = 287 J/kgK. Assume Cdf = 1.0.

[IAS, 2005]


Given: n = 2, Vs = 2 litres or 2 × 10−3 m3, N = 4000 rpm, ηvol = 0.75, A/F = 14, d2 = 30 mm, Cda = 0.9, Cdf = 1.0, p1 = 1.0 bar, T1 = 293 K, ρf = 760 kg/m3, cp = 1.005 kJ/kgK, R = 287 J/kgK

Actual volume sucked by the engine is, V1 =

Air velocity at venturi throat,

or p2 = 0.964 bar

p = p1p2 = 1 − 0.964 = 0.036 bar

Mass of air flowing through the venturi/s,

f = 0.0408/14 = 2.916 × 10−3 kg/s

or df = 1.26 × 10−3 m or 1.26 mm

Example 10.9

An engine fitted with a single jet carburettor having a jet diameter of 1.25 mm has a fuel consumption of 6 kg/hr. The specific gravity of fuel is 0.7. The level of fuel in the float chamber is 5 mm below the top of the jet when the engine is not running. Ambient conditions are 1 bar and 17°C. The fuel jet diameter is 0.6 mm. The discharge coefficient of air is 0.85. Air-fuel ratio is 15. Determine the critical velocity of flow at throat and the throat diameter. Express the pressure at throat in mm of water column. Neglect compressibility effect. Assume discharge coefficient of fuel flow is 0.60.


Given: dj = 1.25 mm, j = 6 kg/hr, Sf = 0.7, z = 5 mm, p1 = 1 bar, T1 = 273 + 17 = 290 K, df = 0.6 mm, Cda = 0.85, A:F = 15, Cdf = 0.60.

pV = mRT
ρf = Sf ρw = 0.7 × 103 = 700 kg/m3

Applying Bernoulli’s equation for air flow at entrance and venturi throat, we have

Mass of air flow per second,

For the flow of fuel, we have


Mass of fuel flow per second,

Critical velocity of flow at throat,

f = Cdf ρf Af cf


or cf = 14.035 m/s


or p2 = 0.31 bar

Again a = Cda ρa Ca2 A2


or d2 = 9.6 mm

Pressure at throat = p2 /ρw ghw



Although the modern carburettor is cheap and reliable, it has several inherent disadvantages that make the supply of correct A:F mixture always difficult. The problem is further accentuated when a single carburettor has to supply the mixture to a multi-cylinder engine. In addition, the throat restricts the flow of air to the engine and the maximum power. The maximum power can be increased by using a large throat as this affects the economy at low speed because of low air speed and worst control on the fuel spray and atomisation. Multiple carburettors assist the distribution problem but they also increase fuel consumption due to low air velocity at low engine speed and capital cost.

For these reasons, many attempts have been made to design a satisfactory petrol injection system, in which each cylinder is supplied with its correct quantity of petrol for each working cycle under all operating conditions. It is also claimed that the system devised to do this, allow the engine to produce more power and gives less vapourisation troubles.

The advantages of fuel injection system over the conventional carburettor have been appreciated for many years. The quest for improved engine performance linked with fuel economy with legislation regarding to control of exhaust emission has enormously increased the use of fuel injection system. The incorporation of electronic control systems has also considerably helped the development of efficient and commercially viable systems.

There are two different injection systems for SI engines which are mechanically operated as follows:

  1. Combustion chamber injection
  2. Continuous port injection

The combustion chamber injection system is just similar to CI engines. However, this system is not adopted as today’s emission and fuel economy requirements make it impractical.

10.13.1 Continuous Port Injection System (Lucas Mechanical Petrol Injection System)

Figure 10.34 shows a simplified line diagram of the aforementioned system. The petrol is sucked from the tank by a pump and pressurised petrol at 3 bar is supplied through a distributor to the fuel injector to a particular cylinder. The relief valve shown in figure maintains the pressure and allows excess petrol to return to the tank. The pump may be driven by engine or by engine or by a separate electric motor. The latter is preferred because it starts pumping at its normal running speed and pressurises as soon as ignition is switched on.

In this system, petrol is continuously injected into the inner port as shown in Fig. 10.34 at a varied rate. The distribution may be made by having a separate metering pump for each cylinder, timed by the arrangement of a series of cams on one cam shaft or by having one pump operated by a single cam with a lob for each cylinder (similar to ignition system) that feeds to distribution unit which passes the petrol to each inlet port in turn. The unit may be driven by shaft, chain, or V-belt.

Nowadays, electronically controlled fuel injection systems are commonly used as they function rapidly and respond automatically to the change in manifold air pressure, engine speed, crankshaft angle, and many other secondary factors. The electronic control unit assesses data (manifold pressure, engine speed, crank angle) received from various sensing devices and then adjusts the A:F supply for the best performance of the engine.

Figure 10.34 Lucas mechanical petrol injection system

This system has to contain a means of supplying additional fuel for cold starting, during warming, and enriching the mixture during acceleration. Ice formation is virtually impossible with this system and the danger of vapourisation is minimised because petrol is under pressure right up to the injection point.

10.13.2 Electronic Fuel Injection System

The amount of fuel supplied to the engine is controlled by three factors namely, injector cross sectional area, fuel pressure, and duration of injection. Injector orifice size is fixed for a particular design. Fuel pressure supplied is controlled by electric pump and regulator. Therefore, the only variable factor in electronic fuel injection is to control the period of injection. This is done by translating the data supplied to the electronic control unit from the various sensors into electric pulses which are in turn relayed to the solenoids which operate the injectors thus determining the moment and period of injection.

The main parts of the system are the injector and electronic control unit.


The solenoid operated fuel injector is shown in Fig. 10.35(a). It consists of a valve body and a needle valve to which the solenoid plunger is rigidly attached. The fuel is supplied to the injector under pressure from the electric fuel pump passing through the filter. The needle valve is pressed against a seat in the valve body by a helical spring to keep it closed until the solenoid winding is energised. When a current pulse is received from the electronic control unit, a magnetic field builds up in the solenoid coil which attracts a plunger and lifts the needle valve from its seat. This opens the path to pressurised fuel to emerge as a finely atomised spray.

Electronic Control Unit

This unit is the heart of electronic injection system which is responsible for metering the quantity of fuel supplied to each cylinder. The unit contains a number of printed circuit boards on which, a series of transistors, diodes, and other electronic components are mounted. This makes the vital data analysing circuits responding to various input signals. After processing the input data, the power output circuits in the control unit generates current pulses which are transmitted to the solenoid injectors to operate the injector for the required period.

Figure 10.35 Fuel injection valve: (a) Solenoid operated valve, (b) Electronically controlled valve

All the electrical units of this system are connected to a cable terminating in a 25-pole plug to the mating socket. Hence, as soon as the ignition switch is actuated, the fuel pump is switched on and the fuel injection system becomes operational. The arrangement of this system is shown in Fig. 10.35(b).

Advantages of Petrol Injection System

  1. The fuel injection system can precisely match fuel delivery to engine requirements under all load and speed conditions. This reduces fuel consumption with no loss of engine performance.
  2. The manifold in an injection system carries only air, so there is no problem of the air and fuel separation and design of manifold becomes simple.
  3. Due to absence of venturi which abstracts the air passage, the petrol injection results in better volumetric efficiency and increased power.
  4. Fuel injection system is relatively free from icing and surge when tilted, cornering, and braking.
  5. Starting and acceleration is much simpler than the carburettor system.
  6. It provides an identical A:F ratio mixture to all the cylinders and maintains better balancing. In addition, the engine can work more economically closer to its lean limit, whereas in a carburettor engine, one or two ratios richer than best economy mixture are to be used to compensate for driveability. This gives lower specific fuel consumption.
  7. By maintaining a precise A:F ratio according to engine requirements, exhaust emissions are lowered. The improved air fuel flow in injection system also helps to reduce emission levels.

Disadvantages of Petrol Injection System

  1. The major disadvantage is its high capital cost due to precise and complicated components of electronic circuit.
  2. The maintenance of this system is difficult and costly as this system contains 4 times the components of mechanical system. A Junker 12 cylinder engine has 1576 parts against 433 parts used with 12 cylinder Mercedes carburettor system.
  3. The weight of the mechanical injection system is considerably higher than that of a carburettor.
  4. The mechanical injection system has many wearing parts such as camshaft, rotor, and so on.
  5. The injection system generates more noise.

10.13.3 Rotary Gate Meter Fuel Injection System

We have seen that a conventional fuel injection system requires individual injector to inject fuel into each cylinder. The additional components increase the overall cost compared to the carburettor system.

This new system, shown in Fig. 10.36, reduces cost because fuel is injected at a central location and then distributed to each cylinder. There is a rotary gate valve at the heart of the system, controlled by engine intake air to regulate fuel injection. As the volume of the air taken into the engine cylinder increases, the gate valve tips to increase the amount of fuel injected into the engine.

Incidentally, Robert Bosch Corporation, Germany, reported that this system helped in optimising fuel and distribution as the fuel is injected into the air stream at the point of highest velocity.

Figure 10.36 Rotary gate make fuel injection system


Unlike SI engines, the fuel in CI engines is supplied at a very high pressure, partly during the compression stroke and partly during the power stroke. The air is taken in during the suction stroke and compressed to a high pressure (30–70 bar) and high temperature (500°C–700°C) according to the compression ratio used (12–20). The high temperature of air at the end of the stroke is sufficient to ignite the fuel.

As the fuel is injected in a high pressure air, the pressure of fuel injected lies between 100–200 bar. During the process of injection, the fuel is broken into very fine droplets. The droplets vapourise, taking the heat from the hot air and form a combustible mixture and start burning. As the burning starts, the vapourisation of fuel is accelerated as more heat is available. As the combustion advances, the amount of oxygen available for burning reduces and therefore, the heat release rate is reduced.

The period between the injection and ignition of fuel is known as delay period (ignition delay). This lies between 0.001 and 0.002 seconds, according to the speed of the engine. This counts for nearly 25° crank rotation. The whole performance of the engine is totally dependent on the delay period. The less is the delay period, better is the performance of the engine.

The main functions of the injection system are as follows:

  1. To supply the correct quantity of fuel to be injected as per the load of the engine and increase in speed for automobile engines.
  2. To supply the fuel within a precisely defined period of the cycles.
  3. To control the rate of fuel injection such that it should result in the desired heat release pattern.
  4. To atomise the fuel into very fine particles.
  5. To distribute the fuel uniformly in the combustion chamber of the engine and results in rapid mixing of fuel and air.
  6. Injection starts and stops sharply. There should not be any after injection.

10.14.1 Types of Injection Systems

Injection system may be divided into two general types, as follows:

  1. Air injection system
  2. Airless or solid or mechanical injection

Air Injection System

It was first developed by Rudolf Diesel. The arrangement of the system is shown in Fig. 10.37. In this system, air and fuel are injected into the cylinder during the supply of fuel. The required pressure of the air for injecting the fuel is about 70 bar or higher.

A fuel pump is driven by the engine itself. A cam shaft operates the fuel pump through a cam and the power required to rotate the cam shaft is taken from the main shaft of the engine with the help of gears and discharges a definite quantity of fuel into the injection valve as shown in Fig. 10.37. The injection valve is mechanically opened and high pressure air drives the fuel charge and some air into the combustion chamber. The amount of fuel delivered is under the control of oil pump suction valve, which is operated by a governor.

The air pressure is raised to about 70 bar by a three-stage compressor (as shown in Fig. 10.37) providing intercooling. The compressor is also operated by the engine. The high-pressure air projects the fuel into the combustion chamber and atomises it. Nowadays, such systems are rarely used in diesel engines.

Figure 10.37 Airless injection system

The advantages and disadvantages of this system are as follows:


  1. It provides better atomisation and distribution of fuel.
  2. As the combustion is more complete, the brake mean effective power (BMEP) is higher than with other types of injection systems.
  3. It allows to use the inferior fuels.


  1. It requires a complicated mechanism to run the compressor.
  2. The weight of the engine increases.
  3. A part of the power is used to drive the compressor and the BHP of the engine is reduced.

Airless or Solid Injection

In this system, the fuel is supplied at a very high pressure (150 bar) from the fuel pump to the fuel injector from where it is injected to the combustion chamber with the help of an injector. The main parts of this system are the fuel pump and the fuel injector. The fuel pump is operated by a cam which is mounted on a cam shaft. The power required to operate the cam is taken from the engine crank shaft. Depending on the location of the fuel pumps and injectors, and the method used to meter the fuel, solid injection may further be classified as follows:

  1. Common rail system: In this system, there is a single high pressure pump, which supplies high pressure fuel to a common header (rail). The accumulator is connected to different cylinders by a separate fuel line through the fuel nozzle as shown in Fig. 10.38. The pressure in the accumulator is maintained constant with the help of a pressure relief valve. The fuel is supplied to each cylinder by operating the respective fuel valve with the help of a cam mechanism driven by the engine crank shaft.

    Figure 10.38 Common rail system

    The quantity of fuel injected and the timing of injection are controlled by fuel valve and not by the injection pump. The arrangement of the fuel valve (metering meter) and the fuel nozzle is shown in Fig. 10.38. This system uses spring loaded injection valves which open and close by mechanical means.

    The pressure used in this system ranges between 110–300 bar, according to the compression ratio used for the engine design.

    The advantages and disadvantages of the system are listed below.


    1. It fulfils the requirements of either the constant load with variable speed or constant speed with variable load.
    2. Only one pump is sufficient for a multi-cylinder engine.
    3. Variation of pump supply pressure will affect all the cylinders uniformly.
    4. The arrangement of the system is very simple and maintenance cost is less.


    1. Very accurate design and workmanship are required.
    2. There is tendency to develop leaks in the injection valve.
  2. Distributor system: This system, like the common rail system, employs a single high pressure pump as shown in Fig. 10.39. The high pressure pump in this system is used for metering and compressing the fuel and then the fuel is delivered to the common rotating distributor. The fuel is supplied to each cylinder by the distributor. In every cycle, the injection strokes of the pump are equal to the number of cylinders. The quantity of fuel supplied and the timing of fuel supply are done by single plunger (main pump). Therefore, equal amount of fuel is supplied to each cylinder and at the same point in the cycle. The function of the distributor is merely to select the cylinder to receive the fuel.

    Figure 10.39 Distributor system

    The distributor block selects a particular cylinder according to the cam coming in contact with the distributor as shown in Fig. 10.39. The appropriate valve opens just before the beginning of injection and oil is supplied to the required cylinder.

  3. Individual pump and nozzle system: This differs from constant pressure injection both in design and operation of the pump and fuel injector. Each injector has a separate pump and the injector contains a spring-loaded hydraulically operated automatic plunger valve. No separate mechanism is required to operate it.

    In this system, separate pumps, each (depending on the number of cylinder) driven individually or a single pump having four plungers in a common block may be used. In this case, the single pump is driven by the crank shaft through a single cam shaft having individual cam for each cylinder.

    The arrangement of the individual pump system is shown in Fig. 10.40 with all pumps in one block, four plungers in one barrel and a common cam shaft.

    The design of this type of pump must be accurate as the volume of fuel injected per cycle is 1/20,000 of the engine displacement at full load and 1/100,000 of the engine displacement during idling. The time allowed for injecting such a small quantity of fuel is very limited (about 1/450 second at 1500 rpm of the engine providing injection through a 20° angle). The pressure requirements vary from 100–300 bar.

    A comparison of various fuel injection systems is given in Table 10.6.

Figure 10.40 Individual pump system


Table 10.6 Comparison of fuel injection systems

10.14.2 Design of Fuel Nozzle

The fuel injection into the cylinder through the fuel nozzle is shown in Fig. 10.41.

Figure 10.41 Fuel injection through an injection nozzle


p1 = pinj = injection pressure, kN/m2

p2 = pcyl = combustion chamber pressure, kN/m2

ρf = density of fuel, kg/cm3

τ = period of injection, s

Q = fuel sprayed, cm3/cycle/cylinder

df = diameter of injector

Let uf = specific volume of fuel (assumed incompressible)

c1 = velocity at section 1 – 1′

c2 = velocity at section 2 – 2′


Neglecting c1, being very small compared with c2,

where Cdf = co-efficient of discharge for the fuel injector.

Duration of injection in second,

Mass of fuel supplied per second,

Volume of fuel injected per second

where d = orifice diameter

n = number of orifices

θ = crank angle during which the fuel is supplied

Ni = number of injections per minute

N = rpm

Fuel consumed per hour, mf = power in kW × SFC in kg/kWh.

Fuel consumed per cylinder per cycle (for 4-stroke engine),

Volume of fuel injected per cylinder per cycle,

Injection period

Example 10.10

Calculate the diameter of the injector orifice of a six-cylinder, 4-stroke CI engine using the following data:

Brake power = 250 kW; Engine speed = 1500 rpm; BSFC = 0.3 kg/kWh; Cylinder pressure = 35 bar; Injection pressure = 200 bar; Specific gravity of fuel = 0.88; Co-efficient of discharge of the fuel orifice = 0.92; Duration of injection = 36° of crank angle.

[IES, 2007]


Given data: BP = 250 kW, N = 1500 rpm, BSFC = 0.3 kg/kWh, p1 = 200 bar, p2 = 35 bar, Specific gravity of fuel = 0.88, Cdf = 0.92, Duration of injection = 36° of crank angle, Number of cylinders = 6, Number of strokes = 4

Fuel consumed per hour = BP × BSFC = 250 × 0.3 = 75 kg/h

Fuel consumed per cylinder per cycle

Volume of fuel injected per cylinder per cycle

Density of fuel, ρf = 1000 × 0.88 = 880 kg/m3

Nozzle hole area,

Diameter of cylinder orifice, df = 0.75 mm.

Example 10.11

A six-cylinder, four-stroke oil engine develops 200 kW at 1200 rpm and consumes 0.3 kg/kWh. Determine the diameter of a single orifice injector if the injection pressure is 200 bar and combustion chamber pressure is 40 bar. The injection is carried for 30° rotation of a crank. Take ρf = 900 kg/m3, and Cdf = 0.7. Each nozzle on a cylinder is provided with a single orifice.


or Af = 0.0476 × 10−4 m2 or 4.76 mm2

or df = 1.005 mm

Example 10.12

A 16-cylinder diesel engine has a power output of 800 kW at 900 revolutions per minute. The engine works on the four stroke cycle and has a fuel consumption of 0.238 kg/kWh. The pressure in the cylinder at the beginning of injection is 32.4 bar and the maximum cylinder pressure is 55 bar. The injector is set at 214 bar and maximum pressure at the injector is around 600 bar. The specific gravity of the fuel is 0.86. Calculate the orifice area required per injector, if the injection takes place over 10 degree crank angle.

[IES, 2001]


Given: i = 16, Pt = 800 kW, N = 900 rpm, n = 2, f = 0.238 kg/kWh, pc = 32.4 bar, pmax = 55 bar, pi = 214 bar, s = 0.86, θ = 10°, (pimax) = 600 bar

Power per cylinder, P =

Fuel consumption per cylinder, mf = f × P = 0.238 × 50 = 1.19 kg/h or 3.3056 × 10−3 kg/s

Fuel to be injected per cycle,

Fuel injection time,

Fuel mass rate per second,

Pressure difference in the beginning, ∆pi = pipc = 214 – 32.4 = 181.6 bar

Pressure difference at the end,

pe = (pi)maxpmax = 600 − 55 = 545 bar

Average pressure difference,

Now mf = CdAi[2ρfpa]1/2

or 0.238 = 0.6 × Ai [2 × 0.86 × 103 × 363.3 × 105]1/2

or Ai = 1.587 × 10–6 m2 or 1.587 mm2

Example 10.13

An 8-cylinder, 4-stroke diesel engine has a power output of 368 kW at 800 rpm. The fuel consumption is 0.238 kg/kW-hr. The pressure in the cylinder at the beginning of injection is 35 bar and the maximum cylinder pressure is 60 bar. The injector is adjusted to operate at 210 bar and the maximum pressure in the injector is set at 600 bar. Calculate the orifice area required per injector if the injection takes place over 12° crank angle. Assume the coefficient of discharge for the injector = 0.6, specific gravity of fuel = 0.85 and the atmospheric pressure = 1.013 bar. Take the effective pressure difference to be the average pressure difference over the injection period.

[IAS, 2004]


Given: i = 8, n = 2, BP = 368 kW, N = 800 rpm, f = 0.238 kg/kWh, p1= 35 bar, pmax = 60 bar, pi = 210 bar, (pi)max = 600 bar, θ = 12°, Cdf = 0.6, Sf = 0.85, patm = 1.013 bar, ∆pe = (∆pavg)i

Fuel consumed by the engine per cylinder =

Fuel consumed per cycle =

Duration of injection =

Rate of fuel consumption

Density of fuel ρf = 0.85 × 103 = 850 kg/m3

or Af = 2.49 × 10–6 m2 or 2.49 mm2

or df = 1.78 mm


The ignition of fuel is concerned only with starting the combustion and not with the behaviour of the combustion flame. For starting the burning of fuel, it is necessary to raise the temperature of the air-fuel mixture to its ignition temperature. The energy required for this purpose is supplied through an electric spark. Within the range of mixtures normally used (12:1 to 15:1) in SI engines, a spark energy of 10 kJ is sufficient to start the combustion process.

10.15.1 Requirement of Ignition System

The important requirements of a spark ignition system are as follows:

  1. The voltage across the spark plug electrodes should be sufficiently large to produce an arc required to initiate the combustion. The voltage necessary to overcome the resistance of the spark gap and to release enough energy to initiate the self-propagating flame front in the combustible mixture is about 10,000–20,000 volts.
  2. The intensity of the spark should lie in a specified limit because extremely high intensity may burn the electrodes and extremely low intensity may not ignite the mixture properly.
  3. The volume of the mixture (clearance volume) at the end of compression should not too large, otherwise the spark produced may not be sufficient to ignite the whole charge. There is definite relation between the size of the spark and clearance volume.
  4. There should be no missing cycle due to failure of spark.
  5. In a multi-cylinder engine, there must be arrangement (distributor) to carry this voltage to the right cylinder at the right time.

10.15.2 Ignition Systems

The basic ignition systems in use are as follows:

  1. Battery ignition system—conventional, transistor-assisted
  2. Magneto ignition system—low temperature, high temperature
  3. Electronic ignition system

Battery and magneto ignition systems differ only in the source of electrical energy. A battery ignition system uses a battery, whereas a magneto ignition system uses a magneto to supply low voltage, all other system components being similar.

Battery Ignition System

The function of battery ignition system is to produce high voltage spark and to deliver it to the spark plugs at regular intervals and at the correct time with respect to the crank position. The required components of the system are as follows:

  1. A battery of 6–12 volts
  2. Ignition coil
  3. Contact breaker
  4. Condenser
  5. Distributor
  6. Spark plug

The arrangement of all the components of battery ignition system for 4-cylinder engine is shown in Fig. 10.41.

The source of current is the storage battery and it is connected to the primary induction coil through the starting switch as shown in Fig. 10.42. The other end of the primary coil is connected to the breaker, which is connected to the ground, when the breaker contact points are closed. (In Fig. 10.42, the breaker contact points are shown in the open position). As one terminal of the battery is grounded, the circuit is closed by passing the current from the battery through the starting switch, primary coil, contact breaker, ground and back to the battery when contact points are closed.

The induction coil consists of primary winding, usually 100–200 turns and a secondary winding, usually 10,000 turns. Both windings are mounted on soft iron core.

The contact breaker consists of contact points, a camshaft on which a cam is mounted which is used to break and make the contacts between the contact points.

The distributor consists of a distributor arm, as shown in Fig. 10.42. The arm is mounted on a cam shaft and is rotated at half the speed of crankshaft. The function of the arm is to make the contact with each spark plug as shown in Fig. 10.42.

Figure 10.42 Battery ignition system for multi-cylinder engines

The distributor unit generally includes contact breaker to make the unit more compact, as both are driven by the same cam shaft.

A condenser is included in the circuit as shown in Fig. 10.42.

Principle of Induction

An EMF is produced in the coil due to the relative movement of magnetic lines and coil because the magnetic field lines are cut by the coil. The EMF produced depends on the relative movement between the two; higher the movement, greater the EMF.

The principle of induction from one coil into another is shown in Figs 10.43(a) to (c). When current is allowed to flow through the primary coil, a magnetic field is set up and this field passes through the secondary coil and induces EMF, sending a current through a closed circuit. The current in the primary coil quickly attains a steady value and a magnetic field is stabilised. The EMF is not induced in the secondary coil as there is no relative movement between field (established by primary) and secondary coil.

If the primary is switched off, the established magnetic field collapses and the EMF is induced again in the secondary coil in the opposite direction. The greater EMF is induced when the circuit is broken because the collapse is more rapid.

The EMF induced in the secondary coil can be further increased by collapsing the established field more rapidly. The magnetic effect in the secondary coil (EMF-produced) is intensified by winding the primary and secondary coils on a common soft iron bar or a ring. The lines of the magnetic field are concentrated around the bar or ring and the magnetic effects are intensified. A higher EMF at the secondary terminal can be obtained by suitable proportioning of primary and secondary coil around a common iron coil (1:100).

Figure 10.43 Principle of induction

Principle of Ignition

High voltage can be introduced at the terminals of the secondary coil by collapsing the field established by primary and proper proportioning of the turns of primary and secondary as mentioned earlier. If this is connected to the two points providing an air gap between them as shown in Fig. 10.44, a spark is produced. Ignition can take place in the compressed charge of petrol engine if the spark is produced in the charge at the end of compression.

Working of Battery Ignition System

When the primary circuit is closed by the contact breaker (shown in open position in Fig. 10.42) a current begins to flow through the primary coil and magnetise core of the coil. The EMF is induced in the secondary as the current in the primary increases. The EMF induced in the secondary coil is proportional to the rate at which the magnetic flux increases. The EMF produced in the secondary coil due to the growth of current in the primary coil is not sufficient to produce a spark at the spark plug because the primary circuit has to establish the magnetic flux.

When the primary circuit is opened by the contact breaker, the magnetic field collapses. Electromotive force is induced in the secondary which is directly proportional to the rate at which the magnetic field of the core collapses which in turn depends on the rate of decrease of the primary current. A condenser is connected across the contact breaker in the primary circuit as shown in Fig. 10.42. This helps to collapse the field very rapidly by absorbing part of the energy of the magnetic field which is thrown back into the primary winding and produces a very high voltage in the secondary. This EMF in the secondary is sufficient to ignite the charge by producing the spark.

Figure 10.44 Principle of ignition

One end of the secondary coil is connected to the ground and the other end is connected to the central terminal of the distributor. The distributor connects the secondary coil in turn to the different spark plugs of the engine in their firing order. The spark plug of a particular cylinder is placed in circuit of the secondary coil with the help of the distributor when the time comes for the charge in that cylinder to be ignited and at the same time the primary circuit is opened by contact breaker. A spark is produced between the points of the spark plug.

The distributor and contact breaker are generally mounted on the same cam-shaft which rotates at half speed of the crankshaft. The function of the distributor is to connect the secondary coil to each cylinder of a multi-cylinder engine at the time of ignition. The contact breaker also works simultaneously with the distributor and its function is to disconnect the primary circuit exactly at the same time when the spark in the particular cylinder is required. The distributor arm connects four spark plugs in one rotation of the cam shaft and therefore, four contact points are required in four cylinder engines. The contact breaker has to break the contacts four times in four cylinder engines and requires four cams as shown in Fig. 10.42. If there are n cylinders, the contact points and cams required are also n in number.

In a single cylinder engine, the distributor is not required as in scooter engine, and single cam is sufficient for giving the spark. An ignition system used in single cylinder petrol engine is shown in Fig. 10.45. Instead of the battery, the magneto is used in this system.

The cam is mounted on the crankshaft only as breaking of circuit during each rotation is required in two stroke engine and there is no necessity of cam-shaft.

Figure 10.45 Battery ignition system for single cylinder engine

Number of Sparks

The number of sparks produced must be equal to the number of working strokes in a single cylinder engine. If there are Nc, cylinders, the number of sparks produced for that engine are as follows:

Ns, (Number of spark) = n × Nc

where n = number of working strokes and Nc = number of cylinders

Further, n = for 4-stroke cycle engine

= N for 2-stroke cycle engine where n is the rpm of the engine

Thus, for 4-stroke engine, Ns = and for 2-stroke engine, Ns = NNc.

Advantages of Battery Ignition System

  1. Its initial cost is low compared with magneto. This is the main reason for the adoption of coil ignition on cars and commercial vehicles.
  2. It provides better spark at low speeds of the engine during starting and idling. This is because the maximum current is available throughout the engine speed range, including starting.
  3. The maintenance cost is negligible, except for the battery.
  4. The spark efficiency (intensity) remains unaffected by advance and retard positions of the timing control mechanism.
  5. The simplicity of the distributor drive is another factor in favour of coil ignition.

Disadvantages of Battery Ignition System

  1. The engine cannot be started if the battery runs down.
  2. The weight of the battery ignition system is greater than the magneto which is a major consideration in adopting the system in aero-engines.
  3. The wiring involved in the coil ignition is more complicated than the one used in a magneto ignition. This results in more likelihood of defects occurring in the system.
  4. The sparking voltage drops with increasing speed of the engine.

Components of a Battery Ignition System

  1. Battery and cut-out: A battery is an electro-chemical device which supplies current because of chemical reactions that occur in it. The common type of battery used in automobiles is a lead acid storage battery. A 12 V battery has six cells each generating two volts. Each cell consists of group of positive and negative plates. The positive plate has a grid of lead and antimony alloy filled with lead peroxide. The negative plate has a grid filled with spongy lead. The positive and negative plates are immersed in an electrolyte of dilute sulphuric acid. The plates are separated from each other by PVC or rubber separators. The positive and negative plate grids are connected to lead antimony strips. These strips separately connect the positive and negative sparks in series and the positive and negative terminals. The plates are placed in a battery container made of hard rubber which is acid-proof. The plate tops are supported by insulating and acid-resisting cell covers. Holes are provided in each cell to fill and unfill electrolytes or distilled water. The filter holes are covered with plugs containing small holds that allow gases to escape.

    The battery is charged continuously by a dynamo directly mounted on the crankshaft. To avoid over charging or discharging of the battery, an electric switch is introduced between the battery and dynamo, which is known as a cut-out.

  2. Ignition coil: The purpose of the ignition coil is to step up 6 V or 12 V battery to 5000 V, which is sufficient to generate the spark. This coil consists of two insulated conducting coils having primary and secondary windings. The primary winding is connected to the battery through an ignition switch and contact breaker and the secondary winding is connected to the spark plugs through the distributor.
  3. Contact breaker: As the number of sparks produced increases with an increase in engine speed, the time available for building up the magnetic field of the spark coil becomes shorter. As a result, the maximum value attained by the primary coil is lessened with increasing speed and the magnetic field produced becomes weaker until there is not enough voltage induced in the secondary winding to produce the spark. Contact breaker is a device to increase the time during which the primary circuit remains closed.
  4. Condenser: For obtaining the highest voltage in the secondary circuit, a quick collapse of the magnetic field is essential. In addition, it is also necessary to prevent the arcing and consequently burning of contact points.

    This is achieved by providing a condenser. The condenser is a device which will absorb and hold an electric charge when an electromotive force is applied to its terminal. In its simplest form, it consists of two sheets of conducting material separated by a layer of insulating material.

    The condenser’s capacity is given by C = A × x × p, where A is the area of conducting material, x is the distance between the sheets of conducting material and p is the specific inductive capacity of the insulating material. The parts of the condenser are shown in Fig. 10.46.

    In order to get large capacity with limited space, instead of two sheets of conducting material, two sets of such sheets are used and are separated by an insulating material. The conducting material used is ‘tinfoil’, whereas the insulating material is mica in condenser used with magneto and wax-impregnated paper for battery system. All conducting sheets of a set are electrically connected. The arrangement for the condenser is shown in Fig. 10.47.

    The operation of the condenser is described as follows. The condenser is connected across the contact breaker. When the contacts points open, instead of passing across the points in the form of an arc, the current flows into the condenser, is stored by it, and becomes charged. The charge in the condenser discharges back immediately in the primary circuit in the direction reverse to the flow of a battery current, thus assisting in a quicker collapse of magnetic field when the contact points open.

    Figure 10.46 Condenser

    Figure 10.47 Condenser arrangement

  5. Distributor: The main function of the distributor is to distribute the high voltage surge to different plugs of multi-cylinder engine at the right time. The high tension current first goes to the control electrode of the distributor and then to the rotor. As the rotor rotates, it passes this high tension current to metal electrodes which are connected to the spark plugs through high tension wiring according to the firing order of the engine.

    These are two types of distributors—brush type and gap type. The blow-up of distributor parts is shown in Fig. 10.48. In the brush type, a carbon brush carried by a rotor arm slides over a metallic segments embedded in the distributor cap of moulded insulating material, thus establishing electrical connection between the secondary winding of the coil and the spark plug.

    In the gap type, the electrode of the rotor arm passes close to, but does not actually contact the segments in the distributor cap. There is no appreciable wear and misfiring due to fouled spark plug.

    When the surface of the spark plug insulator inside the combustion chamber becomes fouled with conducting carbon and oxides, there is considerable leakage of current through this conducting layer which prevents the voltage in the secondary from building up.

    In actual practice, the distributor, contact breaker, rotating shaft with breaker cam, condenser, and the ignition advance mechanism are housed together.

  6. Spark plug: The function of the spark plug is to generate the spark in the combustion using a high voltage communicated by the secondary. The spark plug provides two electrodes with a proper gap across which high potential is discharged and spark is generated.

    A sectional view of a conventional spark plug is shown is Fig. 10.49. It consists of a steel shell, an insulator, and two electrodes. The high voltage supply from secondary is given to the central electrode which is insulated with porcelain. The other electrode is welded to the steel shell of the plug and thereby automatically grounded when the plug is fitted in the cylinder head of the engine. The electrodes are made of high nickel alloy to withstand severe corrosion and erosion to which they are subjected.

    Figure 10.48 Distributor, blow up

    Figure 10.49 Schematic of a typical spark plug

    The tips of central electrode and insulation are exposed to the burned gases. This results in high thermal stresses and the insulator may crack. As the tips are subject to high temperature (2000°C – 25.000°C), heat must flow from the insulator and tip to the surrounding shell in order to cool the electrodes and prevent pre-ignition.

    The spark plugs are classified as hot plug and cold plug, depending on the temperature at the tip of the electrodes. The operating temperature of the tip depends on the amount of heat transferred, which, in turn, depends on the path followed by the heat to flow. A cold plug has a short heat flow path, whereas a hot plug follows a long flow path for the heat to flow as shown in Fig. 10.50.

    The hot plug is used to avoid cold fouling where combustion chamber temperatures are relatively low such as during low power operation and continuous idling.

    The temperature may be less at idling speed and then tips will be fouled by unburned carbon deposits or excess lubricating oil. The carbon deposits burn at 350°C, whereas lubricating deposits burn at 550°C. If the spark plug runs hot at idling speed to prevent carbon deposits, it may run too hot at a high speed. This may cause undesirable pre-ignition. If the plug runs above 800°C, pre-ignition generally occurs.

    The insulator tip length is the most important parameter which controls the operating temperature. Therefore, the tip temperature is generally controlled by varying the insulator tip position and the electrode material.

    It is necessary in practice to compromise to obtain a proper spark plug which would operate satisfactorily throughout the engine operating range. An improper spark plug has remained a major source of engine trouble as misfiring and pre-ignition.

    The factors affecting the operation of spark plug are (i) compression ratio, (ii) electrode temperature, (iii) speed of the engine, (iv) mixture strength, and (v) run of the engine.

Magneto Ignition System

Some of the drawbacks of the battery ignition system are as follows:

  1. There are chances of discharging the battery.
  2. There are chances of misfiring at higher speed of the engine.
  3. It requires complicated wiring.
  4. There are chances of failure.
  5. There are many mechanical complications in the operation of the system.

These difficulties experienced with battery ignition system can be avoided by using magneto ignition system.

Figure 10.50 Heat transfer path of hot and cold spark plug

Magneto Ignition System—Working

The arrangement of the magneto ignition system is shown in Fig. 10.51. The only difference between the battery and the magneto system is that the battery is replaced by the rotating magnet. As the magneto rotates, the voltage and current are generated in the primary coil and the circuit is completed by passing the current through the contact breaking point and the ground. As the current passes through the primary coil through the contact breaker, the circuit is completed through the ground. As the camshaft rotates, the cam 1 opens the contact breaker and interrupts the flow of current in the primary. This causes the decay in the magnetic field lines and cuts the lines of magnetic field in the secondary coil, and a high voltage is generated in the secondary coil. The process of generating the high voltage in the secondary coil is known as the induction phenomenon. The voltage generation in the secondary depends on the ration of number of turns in the secondary and primary coils.

Further, consider the effect of a firing sequence on engine cooling. When the first cylinder is fired, its temperature increases. When the next cylinder fired is number 2, the portion of the engine between the cylinder number 1 and 2 gets overheated. If the third cylinder fired is cylinder number 3, this over-heating is shifted to the portion between the cylinders 2 and 3. Thus, the task of the cooling system becomes very difficult because it is required to cool more at a place than at other places and this place too changes its position with time. If we fire the third cylinder after the first, the overheating problem can be mitigated.

Next, consider the flow of exhaust gases in the exhaust pipe. After firing the first cylinder, exhaust gases flow out to the exhaust pipe. If the next cylinder fired is the cylinder number 2, we find that before the gases exhausted from the first cylinder go out to the exhaust pipe, the gases exhausted from the second cylinder overtake them. This would require that the exhaust pipe be made bigger. Otherwise, the back pressure in it would increase and the danger of back flow will arise. If, instead of firing cylinder number 2, cylinder number 4 is fired, by the time the gases exhausted by cylinder 4 come into the exhaust pipe, the gases from cylinder 1 will have sufficient time to travel the distance between 1 cylinder and 4 and thus, the development of a high back pressure would be avoided.

Figure 10.51 Magneto ignition system

It should be noted that to some extent, all the three requirements are conflicting and therefore, a trade-off is necessary. For four cylinder engines, the possible firing orders are 1-2-4-3 and 1-3-4-2. The latter is in common use. For a six-cylinder engine, the firing orders which can be used are 1-5-3-6-2-4, 1-5-4-6-2-3, 1-2-4-6-5-3, and 1-2-3-6-5-4. The first one is in common use.

Other Firing Orders

For three-cylinder engine 1-3-2

Eight-cylinder in-line engine 1-6-2-5-8-3-7-4

Eight-cylinder V-shape engine 1-5-4-8-6-3-7-2, 1-8-4-3-6-5-7-2, 1-6-2-5-8-3-7-4, 1-8-7-3-6-5-4-2, 1-5-4-2-6-3-7-8. Cylinder no. 1 is taken from front of the in-line engines, whereas in V-shape, front cylinder on right side-bank is considered cylinder no. 1 for fixing H.T. leads according to engine firing order.

Ignition Timing

In order to obtain maximum power from an engine, the compressed mixture must deliver its maximum pressure at a time when the piston is about to commence its downward stroke. Since there is a time lag between the occurrence of spark and the burning of the mixture, the spark must take place before the piston reaches TDC on its compression stroke. Usually, it should occur at about 15° bTDC. If the spark occurs too early, combustion will take place before combustion stroke is completed and the pressure so developed will oppose the piston motion and thereby, reduce the engine power. If the spark occurs too late, the piston will have already completed a certain part of the expansion stroke before the pressure rise occurs and a corresponding amount of engine power will be lost.

The correct instant for the introduction of spark is mainly determined by the ignition lag. The ignition lag depends on many factors such as compression ratio, mixture strength, throttle opening engine temperature, combustion chamber design, and speed. Some of the important factors affecting the ignition timings are as follows:

  1. Engine speed: When an engine is rotating at 2000 rpm, even one-thousandth of a second represents 12° of crank rotation. Therefore, the spark must occur, say, 20° before the TDC for maximum power and economy. As the speed of the engine increases, the combustion time increases in terms of crank angle degrees and the spark must advance accordingly, that is, the angle of the advance must increase as the speed increases (see Fig. 10.52).
  2. Mixture strength: In general, rich mixtures burn faster. Therefore, as the mixture is made richer, the optimum spark advance must be retarded, that is, the number of degrees of the crank angle before TDC is decreased and the spark occurs closer to the TDC.
  3. Part load operation: Part-load operation of a spark ignition engine is affected by throttling. Due to throttling, a smaller amount of charge enters the cylinder and the dilution that occurs due to residual gases is greater. In addition, higher air-fuel ratio required for the part-load operation causes the combustion time to increase. Therefore, at part load, the spark advance must be increased.
  4. Type of fuel: Ignition delay will depend on the type of fuel used in the engine. For maximum power and economy, a slow-burning fuel will need a higher spark advance than a fast-burning fuel.

    It is obvious from the above that the point in the cycle where the spark occurs must be regulated to ensure maximum power and economy at different speeds and loads and must be automatic. Most engines are fitted with a mechanism which is integral with the distributor and automatically regulate the optimum spark advance to account for change of speed and load.

Figure 10.52 Combustion time in terms of engine speed and degrees of crankshaft rotation


Combustion is defined as a rapid chemical reaction between H2 and C with oxygen in the air, and in the process, the reaction liberates energy in the form of heat.

It is absolutely essential to burn the fuel supplied completely for the economical working and safety of the engine. Therefore, the mixture supplied to the engine should possess the required A:F ratio; otherwise, combustion cannot be initiated or if initiated, it cannot be sustained. In addition, there must be some means to initiate the combustion and the generated flame should be able to propagate through the mixture and burn the mixture completely.

It is a known fact that the fuel with any A:F ratio cannot be burned. There is an ignition limit for any fuel to start the combustion and sustain it till the complete fuel burns by the flame generated with spark plug. In addition, the temperature of the mixture to initiate the ignition is equally important. It is also known that the flame will propagate if the temperature of the burnt gases exceeds 1500 K for SI engine (gasoline) fuels. The ignition limits of the hydrocarbon fuel when temperature of mixture reaches to 1500 K are shown in Fig. 10.53.

The upper and lower limits of A:F ratio for ignition depend on the temperature of a particular fuel. The limit becomes wider at higher temperatures because of higher reaction rate and higher thermal diffusivity coefficients of the mixture. Therefore, it is essential to see that the A:F ratio of the mixture supplied to the engine lies in the practical limit as shown in Fig. 10.53.

Figure 10.53 Ignition limit of hydro-carbon fuels at 1500 K

10.16.1 Stages of Combustion in SI Engines

It is assumed that the heat is added instantaneously at constant volume in the ideal air-cycle of SI engines. To achieve this, the burning of fuel in the SI engine should be instantaneous. In actual engines, combustion occurs over a finite period of time as the flame starting around the spark plug has to propagate through the entire mixture of air and fuel.

A high intensity spark is produced for a few degrees of crank angle before TDC at the end of the compression stroke. The function of the spark is to provide a source of heat to the combustible mixture. The temperature of the combustible mixture surrounding the spark is raised to a sufficient temperature level to start the combustion. The flame propagation is continued through the combustible mixture, provided the release of heat from the initial sources is sufficient to heat the adjoining portion of the mixture to a temperature at which the heat from the reaction is sufficient to overcome the heat losses. The heat liberated by the burning portion ‘flame front’ prepares the adjacent portion of the unburned charge for the combustion reaction.

It has been observed experimentally that there is a certain time interval between the instant of spark and the instant when there is a noticeable rise in pressure due to combustion. This means that the combustion starts some time after the spark. This time lag is known as ‘ignition lag’. Ignition lag is a time interval in the process of chemical reaction during which the molecules of the fuel get heated up to the self-ignition temperature, get ignited, and produce a self-propagating nucleus of flame.

The pressure variation in the SI engine combustion chamber during the crank rotation is shown in Fig. 10.54. This is an unfold p-v diagram.

The ignition is timed to take place at the point a but the burning commences only at the point ‘b’. The time interval between the points ‘a’ and ‘b’ is known as ‘ignition lag’. This is generally expressed in terms of crank angle and is given by θ1 as shown in Fig. 10.54. The time required for the crank to rotate through an angle θ2 is known as combustion period during which the propagation of flame takes place. Many times, the combustion is not complete at the point c (when the mixture is too rich) and the fuel may burn after the point c during the expansion stroke. This burning of fuel after the point of maximum pressure is known as ‘after burning’. This is not desirable.

The major disadvantages of the ignition lag is to reduce the power developed. If the ignition lag is long, the peak pressure occurs during the expansion stroke, and complete advantage of expansion is not achieved. The time lag may vary from 10°–50°, according to the type of the fuel used and several other factors affecting the time lag. To maintain the ‘time lag’ constant at variable speed of the engine, it is necessary to change the angle of advance. The angle of advance must increase with increase in speed. This is commonly done in automobile engines by an automatic advancing mechanism.

Figure 10.54 Crank angle v’s pressure variation in SI engine

All considerations should be taken into account in the design of the combustion chamber and selecting the fuel used to reduce the ignition lag.

The theoretical diagram of combustion is shown in Fig. 10.55(a).

However, the actual process differs from theory as instant combustion is not possible as shown by bc in Fig. 10.55(a). Some time (order of millisecond) is needed to start the combustion after giving the spark because of the fact that the surrounding mixture is to be heated up to ignition temperature and then formed nucleus of flame which starts propagating through the surrounding mixture.

The pressure variation in the engine with respect to crank-angle is shown in Fig. 10.55(b). There are mainly three stages of combustion in SI engines as shown in Fig. 10.55(b).

  1. First phase: This phase is considered between the point of ignition and point of combustion. As shown in Fig. 10.55(b), ignition is timed at point a and combustion starts at point b. The time interval between the points a and b is known as ignition lag. The ignition lag is generally expressed in terms of crank angle (θ1). This period of ignition lag is very small and lies between 0.00015 and 0.002 seconds. An ignition lag of 0.002 seconds corresponds to 35° crank rotation when the engine is running at 3000 rpm and crank angle required (which is known as angle of advance) increases with an increase in engine speed.

    Figure 10.55 Pressure vs crank angle diagram:
    (a) Theoretical, (b) Actual

  2. Second phase: Once the flame is formed at point b, it should be self-sustained and be able to propagate through the mixture. This is possible when the rate of heat generation (Qg) by burning the surrounding mixture of the flame nucleus is higher than the heat lost (Ql) by the flame to the surrounding. As the difference between (QgQl) is higher, the rate of flame propagation is higher and complete combustion will occur at the earliest, which is the most desirable requirement of combustion in SI engines. The propagation of flame also depends on the flame temperature and the temperature and density of the surrounding mixture as its propagation is directly proportional to these factors. A weak spark and a low compression ratio (as density of mixture is less) give low propagation of the flame.

    After point b, the flame propagation is abnormally low in the beginning as heat lost is more than heat generated. Therefore, the pressure rise is also slow as the mass of mixture burned is small. Therefore, it is necessary to provide an angle of advance of 30° to 35° if the peak pressure is to be attained at 5°–10° after TDC.

    After the point c, the pressure starts falling due to the fall in the rate of heat release when the flame reaches the wall in the last part of combustion and cannot compensate for its fall due to the gas expansion, and heat is transferred to the walls.

    The time required for the flame to travel 95% of the chamber length with respect to speed of the engine is shown in Fig. 10.56, when θ1 = 30° and A:F = 13.

    It is obvious that the crank angle required for 95% travel increases with an increasing engine speed (the time available is decreased), therefore, if the combustion is to be completed at point c, the angle of advance (θ) must be increased with increasing engine speed. The flame speed increases with increasing engine speed because of the increase in turbulence of the mixture.

    The time required for the crank to rotate through an angle θ2 is known as combustion period during which the propagation of the flame takes place.

    Stages I and II are not entirely distinct. The starting point of stage II is measurable as the rise in pressure can be seen on the p-θ diagram. This is the point where the line of combustion departs from the line of compression.

    Figure 10.56 Time required for the flame to travel 95% of the chamber length

  3. Third phase: Although the point c represents the end of the flame travel, it does not assure the complete combustion of fuel. In this case, the combustion still continues after attaining the peak pressure and is known as afterburning. This is continued throughout the expansion stroke. This generally happens when the rich mixture is supplied to the engine.

10.16.2 Ignition Lag (or Delay) in SI Engines

The ignition delay period is the phase during which some fuel has already been admitted but has not yet ignited. The fuel does not ignite immediately upon injection into the combustion chamber. There is a definite period of inactivity between the time when the first droplet of fuel hits the hot air in the combustion chamber and the time it starts through the actual burning phase. This period is known as the ignition delay period. In Fig. 10.57, the delay period is shown on the pressure crank angle (or time) diagram between points a and b. Point a represents the time of injection and point b represents the time at which the pressure curve first separates from the motoring curve.

The important factors affecting delay period are as follows:

  1. Compression ratio: With increase in compression ratio, the minimum auto ignition temperature of a fuel decreases due to increased density of the compressed air. This results in closer contact between the molecules of fuel and oxygen, reducing the time of reaction. The increase in the compression temperature as well as the decrease in the minimum auto ignition temperature decreases the delay period. The peak pressure during the combustion process is only marginally affected by the compression ratio (because delay period is shorter with higher compression ratio and hence the pressure rise is lower).

    One of the practical disadvantages of using a very high compression ratio is that the mechanical efficiency tends to decrease due to increase in weight of the reciprocating parts. Therefore, in practice, engine designers always try to use a lower compression ratio which helps in easy cold starting and light load running at high speeds.

  2. Engine speed: The delay period decreases with an increase in engine speed in a variable speed operation with a given fuel. With increase in engine speed, the loss of heat during compression decreases, resulting in the rise of both the temperature and pressure of the compressed air thus reducing the delay period.
  3. Power output: With an increase in engine output, the air–fuel ratio decreases, operating temperatures increase, and hence, delay period decreases.

    Figure 10.57 Pressure-time diagram illustrating ignition delay

  4. Quality of fuel: A lower self-ignition temperature results in a lower delay period. In addition, fuels with higher cetane number give lower delay period, and smoother engine operation. Other properties such as volatility, latent heat, viscosity, and surface tension also affect the delay period.

Table 10.7 shows the summary of the effects of variables on delay period.


Table 10.7 Effect of variables on delay period

10.16.3 Factors Affecting the Flame Propagation

After the ignition, the rate of flame propagation affects the combustion process in SI engines.

Better the propagation, higher is the combustion efficiency and higher the economy. The propagation depends on the flame velocity. Unfortunately, flame velocities for most fuels range between 10–30 m/s. Therefore, all steps should be taken in the design and operation of the engine so that the flame velocity is as maximum as possible.

The study of flame propagation rate and the factors affecting the same is important for the following two reasons:

  1. Its effect on the rate of pressure rise in the cylinder.
  2. Its effect in connection with abnormal combustion (knocking).

The factors which affect the flame propagation are discussed below:

  1. A:F ratio: The mixture strength influences the rate of combustion and the amount of heat generated. The maximum flame speed for all hydrocarbon fuels occurs at nearly 10% rich mixture. Flame speed is reduced for lean and rich mixture.

    Lean mixture releases less heat, resulting in lower flame temperature and lower flame speed. Very rich mixture results in incomplete combustion (C → CO instead of CO2) and in production of less heat; the flame speed remains low.

    The effects of the A:F ratio on the p-v diagram and p-θ diagram are shown in Fig. 10.58.

  2. Compression ratio (r): A higher compression ratio increases the pressure and temperature of the mixture and also decreases the concentration of residual gases. All these factors which are in favour reduces the ignition lag and helps to speed up the second phase of combustion. The maximum pressure of the cycle as well as the mean effective pressure of the cycle will increase with an increase in r. Figure 10.59 shows the effect of the compression ratio on pressure (indirectly on the speed of combustion) with respect to crank angle for same A:F ratio and same angle of advance. Higher compression ration increases the surface to volume ratio and thereby increases the part of the mixture which afterburns in the third phase.

    Figure 10.58 Indicator diagrams for stoichiometric and weak mixtures

    Figure 10.59 Effect of compression ratio on pressure rise

  3. Load on engine: With increase in load, the cycle pressures increase and the flame speed also increases.

    In an SI engine, the power developed by an engine is controlled by throttling. At lower load and higher throttle, the initial and final pressure of the mixture after compression decrease and the mixture is also diluted by more residual gases. This reduces the flame propagation and prolongs the ignition lag. Therefore, the advance mechanism is also provided with a change in load on the engine. This difficulty can be partly overcome by providing rich mixture at part loads but this definitely increases the chances of after-burning. The after-burning is prolonged with richer mixture.

    In fact, poor combustion at part loads and necessity of providing richer mixture are the main disadvantages of SI engines which cause wastage of fuel and discharge of a large amount of CO with exhaust gases.

  4. Turbulence and engine speed: Turbulence plays very important role in the combustion of fuel as the flame speed is directly proportional to the turbulence of the mixture. This is because turbulence increases the mixing and heat transfer coefficient or heat transfer rate between the burned and the unburned mixture. The turbulence of the mixture can be increased at the end of compression by a suitable design of the combustion chamber (geometry of cylinder head and piston crown).

    Insufficient turbulence provides low flame velocity and incomplete combustion and reduces the power output. However, excessive turbulence is also not desirable as it increases the combustion rapidly and leads to detonation. Excessive turbulence causes to cool the flame generated and flame propagation is reduced.

    Moderate turbulence is always desirable as it accelerates the chemical reaction, reduces ignition lag, increases flame propagation, and even allows the weak mixture to burn efficiently.

    The turbulence of the mixture increases with an increase in engine speed. Therefore, the effects of an increase in engine speed are similar to increase in turbulence.

  5. Other factors: Among other factors, the factors which increase flame speed include supercharging of the engine, spark timing, and residual gases left in the engine at the end of exhaust stroke.

    The air humidity also affects the flame velocity but its exact effect is not known. Anyhow, its effect is not large compared with A:F ratio and turbulence.

Example 10.14

The spark plug is fixed at 18° before top dead centre (TDC) in an SI engine running at 1800 rpm. It takes 8° of rotation to start combustion and get into flame propagation mode. Flame termination occurs at 12°C after TDC. Flame front can be approximated as a sphere moving out from the spark plug which is offset 8 mm from the centre line of the cylinder whose bore diameter is 8.4 cm. Calculate the effective flame front speed during flame propagation. The engine speed is increased to 34000 rpm and subsequently as a result of which the effective flame front speed increases at a rate such that it is directly proportional to 0.85 times of engine speed. Flame development after spark plug firing still takes 8° of engine rotation. Calculate how much engine rotation must be advanced such that the flame termination again occurs at 12° after TDC.

[IES, 2010]


Given that θ1 = 8°, θ2 = 10° + 12° = 22°, N1 = 1800 rpm, e = 8 mm, D = 8.4 cm, N2 = 34000 rpm,

The combustion process in p-v diagram is shown in Fig. 10.60.

Time taken,

Flame front speed,

When engine speed is increased to 34000 rpm, then

Figure 10.60 Combustion process in p-v diagram

10.16.4 Phenomena of Knocking/Detonation in SI Engines

In a SI engine, combustion is initiated between the spark plug electrodes and then spreads across the combustible mixture. A definite flame front which separates the fresh mixture from the products of combustion travels from the spark plug to the other end of the combustion chamber. Heat release due to combustion increases the temperature and pressure of the burned part of the mixture above those of the unburned mixture. In order to effect pressure equalisation, the burned part of the mixture will expand and compress the unburned part adiabatically, thereby increasing the pressure and temperature of the unburned part further. If the temperature of the unburned mixture exceeds the self-ignition temperature of the fuel and remains at or above this temperature during the ignition lag, spontaneous ignition (or auto-ignition) occurs at various pin-point locations. This phenomenon is called detonation or knocking.

The phenomenon of knocking may be explained by referring to Fig. 10.61(a) which shows the cross-section of the combustion chamber with flame advancing from the spark plug location A without knock, whereas Fig. 10.61(c) shows the combustion process with knock. In the normal combustion, the flame travels across the combustion chamber from A towards D. The advancing flame front compresses the end charge BBD farthest from the spark plug, thus raising its temperature. The temperature is also increased due to heat transfer from the hot advancing flame front and the pre-flame oxidation in the end charge. If the temperature of the end charge does not reach its self-ignition temperature, the charge would not auto-ignite and the flame will advance further and consume the charge BBD. This is the normal combustion process which is illustrated by means of the pressure-time diagram in Fig. 10.61(b). If the end charge BBD reaches its auto-ignition temperature and remains for some period of time equal to the time of pre-flame reactions, the charge will auto-ignite, leading to detonation. In Fig. 10.61(c),it is assumed that when the flame has reached the position BB′, the charge ahead of it has reached critical auto-ignition temperature. During the pre-flame reaction period, if the flame front could move from BB′ to only CC′, the charge ahead of CC ′ would auto-ignite. Due to auto-ignition, another flame front starts travelling in the opposite direction to the main flame front. When the two flame fronts collide, a severe pressure pulse is generated. The gas in the chamber is subject to compression and rarefaction along the pressure pulse until pressure equilibrium is restored. This disturbance can force the walls of the combustion chamber to vibrate at the same frequency as the gas (≈ 5000 Hz). The pressure-time trace of such a situation is shown in Fig. 10.61(d).

Figure 10.61 Normal and abnormal combustion: (a) Movement
of flame front for normal combustion, (b) P-θ diagram for normal combustion, (c) Movement of flame front for abnormal combustion, (d) P-θ diagram for abnormal combustion

10.16.5 Factors Influencing Detonation/Knocking

The factors influencing knocking are as follows:

  1. Density factors: Any factor which reduces the density of the charge tends to reduce knocking by providing lower energy release.
    1. Compression ratio: Increase in compression ratio increases the pressure and temperature of the gases at the end of compression stroke which decreases the ignition lag of the end gas and increases the tendency of knocking.
    2. Mass of induced charge: An increase in the mass of induced charge into the cylinder of an engine by throttling or by increasing the amount of supercharging increases both temperature and density of the charge at the time of ignition. This increases the tendency for knocking.
    3. Inlet temperature of mixture: Increase in inlet temperature of mixture makes the compression temperature higher thereby, increasing the tendency of knocking.
    4. Temperature of combustion chamber walls: Hot spots (such as spark plug and exhaust valve) in combustion chamber promote knocking.
    5. Retarding spark timing: Having the spark closer to the TDC reduces knocking. However, this affects the brake torque and power output of the engine.
    6. Power output of the engine: A decrease in the output of the engine reduces the tendency to knock.
  2. Time factors: The following factors reduce the possibility of knocking:
    1. Turbulence: Turbulence increases the engine speed and reduces the time available for the end charge to attain auto-ignition conditions thereby decreasing the tendency to knock.
    2. Engine speed: An increase in engine speed increases turbulence and decreases the tendency to knock.
    3. Flame travel distance: Shortening the time required for the flame front to traverse the combustion chamber reduces the tendency for knocking. Flame travel distance is governed by combustion chamber size and spark plug position.
    4. Engine size: A large engine (combustion chamber size) has a greater tendency for knocking as the flame requires a longer time to travel across the combustion chamber.
    5. Combustion chamber shape: The combustion chamber should be such that it promotes turbulence to reduce knocking. Spherical chambers minimise knocking tendency.
    6. Location of spark plug: A centrally located spark plug or multiple spark plugs minimise flame travel time and reduce knocking.
  3. Composition factors:
    1. Fuel: F/A ratio affects the reaction time or ignition delay. When the mixture is nearly 10% richer than stoichiometric (F/A ratio = 0.08), ignition lag of the end gas is minimum, velocity of flame propagation is maximum, and knocking tendency is maximum. A rich mixture is effective in decreasing the knocking tendency due to a longer delay period and lower temperature of compression.
    2. Octane rating of fuel: Higher the octane number, lesser is the tendency for knocking. Paraffin series have the maximum and aromatic series the minimum tendency to knock. The knocking characteristics of a fuel can be decreased by adding small amounts of additives called dopes.
    3. Humidity of air: Increasing the humidity of the atmospheric air decreases the tendency to knock.
  4. Temperature factors: Increasing the temperature of the unburned mixture by any factor in design or operation will increase the possibility of knocking in SI engine. The temperature of the unburned mixture is increased by the following factors:
    1. Raising the compression ratio: For a given engine setting and fuel, there will be a critical compression ratio above which knocking occurs. This compression ratio is called the highest useful compression ratio (HUCR).
    2. Supercharging
    3. Raising the inlet temperature
    4. Raising the coolant temperature
    5. Increasing the load (opening the throttle)
    6. Raising the temperature of the cylinder and combustion chamber walls
    7. Advancing the spark timing

A summary of variable affecting knocking in SI engines is given in Table 10.8.


Table 10.8 Summary of variables affecting knock in an SI engine


10.16.6 Methods for Suppressing Knocking

  1. Proper location of spark plug: A compact combustion chamber with proper central location of the spark plug reduces the path of the flame front travel from the spark plug to the remotest part of the combustion chamber, and thereby, reduces the tendency of knocking. The same result could be obtained by multi-fuel injection system.
  2. Proper selection of material for piston and cylinder head: The end charge cooling is much better when the piston and cylinder are made of aluminium alloys as they have good thermal conductivity. A better cooling of the end charge reduces the tendency for knocking.
  3. Injecting water into the intake manifold: Injection of water in the cylinder reduces the temperature of the end gas and increases the delay period, thereby decreasing the tendency to knock.
  4. Retarding the spark timing: By retarding the spark timing, the peak pressures are reached only during the power stroke and are of lower magnitude. This reduces knock.
  5. Extremely rich or lean mixture: By using extremely rich or lean mixture, the flame temperature can be kept low, thus considerably eliminating the tendency of knock.
  6. Squish recesses in combustion chamber: The provision of squish recess in combustion chamber cools the last position of the charge and reduces the tendency of knock.

10.16.7 Effects of Knocking/Detonation

  1. Mechanical damage: Knocking creates a high pressure wave of large amplitude. This increases the rate of wear of almost all mechanical parts such as piston, cylinder head, and valves. The engine parts are also subjected to very high temperatures due to auto-ignition and the piston is damaged by overheating.
  2. Noise: When the intensity of the knock is high, a loud pulsating noise is created due to high intensity pressure wave which vibrates back and forth across the cylinder. The high vibrating motion of the gases causes crankshaft vibrations and engine runs rough.
  3. Heat transfer rate: The heat transfer rate increases as more heat is lost to the coolant.
  4. Power output: The power output slightly decreases due to rapid burning of last part of the charge.
  5. Pre-ignition: Pre-ignition is the ignition of the charge in the absence of spark as it comes in contact with hot surface. This adds negative work by the piston during compression which increases high peak pressure and temperature in the cylinder.