Chapter 11 Performance of Internal Combustion Engines – Thermal Engineering

Chapter 11

Performance of Internal Combustion Engines

11.1 ❐ PERFORMANCE PARAMETERS

The performance of an engine is an indication of the degree of success with which the conversion of chemical energy contained in the fuel is done into useful mechanical work. The degree of success is compared on the basis of following parameters:

  1. Specific fuel consumption (SFC)
  2. Brake mean effective pressure (BMEP)
  3. Specific power output (SP)
  4. Specific weight (SW)
  5. Exhaust smoke and other emissions

However, in the evaluation of engine performance, the following performance parameters are chosen:

  1. Power and mechanical efficiency
  2. Mean effective pressure and torque
  3. Specific output
  4. Fuel-air ratio
  5. Volumetric efficiency
  6. Specific fuel consumption
  7. Thermal efficiency and heat balance
  8. Exhaust smoke and other emissions
  9. Specific weight

Indicated power,

where pim = indicated mean effective pressure, Pa

L = Length of stroke, m

A = Area of stroke, m2

N = rotational speed, rpm

n = number of strokes per revolution

= 2 for four-stroke engine

= 1 for two-stroke engine

where pbm = brake mean effective pressure, Pa

Specific output: It is defined as the brake output per unit of piston displacement.

Volumetric efficiency,

= va /vs

Brake specific fuel consumption,

where mf = mass of fuel consumed, kg

t = time in minutes

Indicated specific fuel consumption,

where LCV = lower calorific value of fuel, kJ/kg

where ηa = air standard efficiency

Exhaust smoke and other emissions: Oxides in nitrogen, unburned hydrocarbons, etc.

11.2 ❐ BASIC ENGINE MEASUREMENTS

The basic measurements required to be carried out to evaluate the performance of an engine are as follows:

  1. Speed
  2. Fuel consumption
  3. Air consumption
  4. Smoke density
  5. Brake power
  6. Indicated power and friction power
  7. Heat loss to cooling water
  8. Heat going to exhaust
  9. Exhaust gas analysis
  1. Speed Measurement: The angular speed of engine (crankshaft) may be measured by a tachometer. Tachometers may be classified into the following categories:
    1. Mechanical tachometers
    2. Electrical tachometers

    They can further be classified as contact type and non-contact type. Contact type tachometers are of the following types:

    1. Mechanical type: Revolution counter and timer, slipping clutch tachometers, and centrifugalfree tachometers.
    2. Electrical type: Drag cup tachometer, electromagnetic tachogenerator. Non-contact type tachometers are photoelectric tachometer, variable reluctance tachometer, stroboscope, and capacitance pick up tachometer.
  2. Fuel consumption measurement: The fuel consumption of an engine is measured by determining the volume flow in a given time interval and multiplying by the specific gravity of the fuel. Another method is to measure the time required for consumption of a given mass of fuel. Accurate methods for fuel measurement are as follows:
    1. Two glass-chambers method
    2. Rotameter
    3. Positive displacement meter
  3. Measurement of air consumption: Accurate measurements for air consumption can be carried out by the following methods:
    1. Orifice chamber method
    2. Viscous-flow air meter
  4. Measurement of exhaust smoke: A smoke meter is a sort density measuring device and may be used to measure exhaust smoke.
  5. Measurement of heat carried away by exhaust gases: An exhaust gas calorimeter is commonly used for the measurement of heat carried by exhaust gases.

    Mass of air supplied per kg of fuel,

    Heat carried away by exhaust gas per kg of fuel,

    where (ma + 1) = mass of exhaust gases formed per kg of fuel supplied

    cpg = specific heat of exhaust gases, kJ/kgk

    Tge = temperature of exhaust gases coming out from engine, °C

    Ta = temperature of ambient air, °C

  6. Measurement of heat carried away by cooling water: The heat carried away by cooling water is generally measured by measuring the water flow rate through the cooling jacket and the rise in temperature of water during flow through the engine. The inlet and outlet temperatures of water are measured by thermometers inserted in the pockets provided at the inlet and outlet from the engine. The quantity of water flowing is measured by collecting the water in a bucket for a specified period.

    Heat carried away by cooling water,

    where mw = mass of water collected /min

    cpw = specific heat of water, kJ/kg °C

    Twi = inlet temperature of water, °C

    Tw0 = outlet temperature of water, °C

11.3 ❐ HEAT BALANCE SHEET

A heat balance sheet is an account of heat supplied and heat utilised during engine trial.

  1. Heat supplied to the engine,
    Qs = ṁf × LCV

    where f = mass of fuel supplied per hour.

    LCV lower calorific value of fuel, kJ/kg

  2. Heat utilised in various ways:
    1. Heat equivalent of BP,
      Q1 = 3600 × BP in kW kJ/h
    2. Heat carried away by cooling water,
      Q2 = wcpw (Tw0Twi) kJ/h
    3. Heat carried away by exhaust gases,
      Q3 = gcpg (Tge − Ta) kJ/h
    4. Heat unaccounted for,
      Q4 = Qs − (Q1 + Q2 + Q3)
11.4 ❐ WILLAN’S LINE METHOD

It is a method to determine the frictional power of an engine. In this method, gross fuel consumption v’s brake power at a constant speed is plotted and the graph is extrapolated back to zero fuel consumption as shown in Fig. 11.1. The point where this graph cuts the BP axis is an indication of the frictional power of the engine at that speed. The test is applicable to only compression ignition engines.

Figure 11.1 Willan’s line method

11.5 ❐ MORSE TEST

This is a method to determine the indicated power of a cylinder in a multi-cylinder engine. In this method, the engine is first run at the required speed and the output is measured. One cylinder is then cut out by short circuiting the spark plug or by disconnecting the injector, as the case may be. The output is measured by keeping the speed constant at its original value. The difference in the output is a measure of the indicated power of the cutout cylinder. Thus for each cylinder the IP is obtain and is added together to find the total IP of the engine.

IP of n cylinders, IPn = BPn + FPn

IP of (n − 1) cylinders IPn1 = BPn1 + FPn

IP of nth cylinder, (IP)nth = BPn − BPn1 = IPn − IPn1

Total IP of engine, IPn = ∑(IP)nth

11.6 ❐ PERFORMANCE OF SI ENGINES

The performance of an engine is generally given by a heat balance sheet. The main components of a heat balance sheet are (i) heat equivalent to the effective (brake) work of the engine, (ii) heat rejected to the cooling medium, (iii) heat carried away from the engine with the exhaust gases, and (iv) unaccounted losses. The unaccounted losses include the radiation losses from the various parts of the engine and heat loss due to incomplete combustion. The friction loss is not shown as a separate item as it ultimately reappears as heat in cooling water, exhaust, and radiation.

Figure 11.2 shows the heat balance sheet for a petrol engine run at full throttle over its speed range. In SI engines, the loss due to incomplete combustion included in unaccounted losses can be rather high. For a rich mixture (A /F ratio 12.5:13) it could be 20%. Figure 11.3 shows the heat balance of an uncontrolled SI engine at different loads.

Figure 11.2 Heat balance diagram for a typical SI engine

Figure 11.3 Uncontrolled SI engine

Figure 11.4 Efficiency and specific fuel consumption vs speed for a SI engine at full throttle

Figure 11.4 shows the brake thermal efficiency, indicated thermal efficiency, mechanical efficiency, and specific fuel consumption for the above SI engine.

Figures 11.5 (a) and (b) show the indicated power (IP), brake power (BP), and friction power (FP) (by difference), brake torque, brake mean effective pressure, and brake specific fuel consumption of a high compression ratio (r) automotive SI engine at full or wide open throttle (WOT).

The following conclusions can be drawn from the above figures:

  1. At full throttle, the brake thermal efficiency at various speeds varies from 20%–27%, maximum efficiency being at the middle speed range.
  2. The percentage heat rejected to a coolant is more at lower speed (>> 35%) and reduces at higher speeds (>> 25%). Considerably more heat is carried by the exhaust at higher speeds.
  3. Torque and mean effective pressure (MEP) do not strongly depend on the speed of the engine but depend on volumetric efficiency and friction losses. Maximum torque position corresponds with the maximum air charge or maximum volumetric efficiency position.

    Torque and MEP curves peak at about half that of the brake power.

    Figure 11.5 Variable speed test of automotive SI engine at full throttle (CR = 9): (a) Power, mep vs engine speed, (b) ηmech, SFC vs engine speed

    Note: If size (displacement) of the engine were to be double, torque would also double, but MEP is a ‘specific’ torque—a variable independent of the size of the engine.

  4. High brake power arises from high speed. In the speed range before the maximum brake power is obtained, doubling the speed doubles the power.
  5. At low engine speed, the frictional power is relatively low and brake power is nearly as large as indicated power. As engine speed increases, however, friction power increases at continuously greater rate and therefore, brake power reaches a peak and starts reducing even though indicated power is increasing. At engine speeds above the usual operating range, friction power increases very rapidly. At these higher speeds, indicated power will reach a maximum and then fall off. At some point, indicated power and friction power will be equal, and brake power will then drop to zero.

11.6.1 Performance of SI Engine at Constant Speed and Variable Load

The performance of an SI engine at constant speed and variable load is different from the performance at full throttle and variable speed. Figure 11.6 shows the heat balance of an SI engine at constant speed and variable load. The load is varied by altering the throttle and the speed is kept constant by resetting the dynamometer.

Figure 11.6 Heat balance vs load for a SI engine

Closing the throttle reduces the pressure inside the cylinders but the temperature is affected very little because the air /fuel ratio is substantially constant, and the gas temperatures throughout the cycle are high. This results in high loss to coolant at low engine loads.

At low loads, the efficiency is about 10%, rising to about 25% at full load. The loss to the coolant is about 60% at low loads and 30% at full load. The exhaust temperature rises very slowly with load and as the mass flow rate of exhaust gas is reduced because the mass flow rate of fuel into the engine is reduced, the percentage loss to exhaust remains nearly constant (about 21% at low loads to 24% at full load). Percentage loss to radiation increases from about 7% at low loads to 20% at full load.

11.7 ❐ PERFORMANCE OF CI ENGINES

The performance of a CI engine at constant speed and variable load is shown in Fig. 11.7. As the efficiency of the CI engine is more than the SI engine, the total losses are less. The coolant loss is more at low loads and radiation and other losses are greater at high loads. The BMEP, BP, and torque directly increase with load, as shown in Figs. 11.8(a) and (b). Unlike the SI engine, the BP and BMEP are continuously rising curves and are limited only by the smoke. The exhaust temperature is also nearly proportional to the load. The lowest BSFC and hence the maximum efficiency occurs at about 80% of the full load.

Figure 11.7 Heat balance diagram for a typical CI engine

Figure 11.8 IP, BP, IMEP, BMEP, and SFC for a CI engine

Figure 11.9 Performance curves of a diesel engine

Figure 11.9 shows the performance curves of a variable speed GM 7850 cc four cycle V-6 Toro-flow diesel engine. The maximum torque value is at about 70% of maximum speed compared to about 50% in the SI engine. In addition, the BSFC is low through most of the speed range for the diesel engine and is better than the SI engine.

11.8 ❐ PERFORMANCE MAPS

For critical analysis, the performance of an IC engine under all conditions of load and speed is shown by a performance map. Figure 11.10 shows the performance map of an automotive SI engine and Fig. 11.11 shows the performance map of a four-stroke pre-chamber CI engine. Figure 11.11 also includes a typical curve of BMEP versus piston speed for level road operation in high gear. Note that these maps can be used for comparing engines of different sizes as performance parameters have been generalised by converting rpm into piston speed and brake power per time of piston area.

Figure 11.10 Form of performance map for a SI engine

Figure 11.11 Form of performance map for a diesel engine

Generally, all engines show a region of the lowest specific fuel consumption (highest efficiency) at a relatively low piston speed with a relatively high BMEP.

SI Engine: Constant speed line: Increased BSFC is obtained by moving upward along the constant speed line because of mixture enrichment at high load which offsets an increase in mechanical efficiency. Moving to lower BMEP, the BSFC increases because of the reduced mechanical efficiency (IMEP decreases, whereas FMEP remains constant).

Constant BMEP line: Moving from the region of the highest efficiency along a line of constant BMEP, the BSFC increases due to increased friction at higher piston speeds. Moving to the left towards lower piston speed, although friction MEP decreases, indicated efficiency falls off owing to poor fuel distribution and increased relative heat losses.

CI Engine: In the CI engine, the BSFC increases at high loads owing to increased fuel waste (smoke) associated with high fuel-air ratios. At lower load, BSFC increases due to decrease in mechanical efficiency (same in the SI engine).

As speed is reduced from the point of the best economy along a line of constant BMEP, the product of mechanical and indicated thermal efficiency appears to remain about constant down to the lowest operating speed. The reduction in FMEP with speed is apparently balanced by a reduction in indicated thermal efficiency due to poor spray characteristics at very low speed.

An interesting feature of performance curves is that they show that power at maximum economy is about half of the maximum power.

Example 11.1

A six-cylinder, four-stroke spark-ignition engine of 10 cm × 2 cm (bore × stroke) with a compression ratio of 6 is tested at 4800 rpm on a dynamometer of arm 55 cm. During a 10 minutes test, the dynamometer reads 45 kg and the engine consumed 5 kg of petrol of calorific value 45 MJ/kg. The carburettor receives air at 9°C and 1 bar at the rate of 10 kg /min. Calculate (a) the brake power, (b) the brake mean effective pressure, (c) the brake specific fuel consumption, (d) the brake specific air consumption, (e) the brake thermal efficiency, and (f) the air-fuel-ratio.

Solution

Given that i = 6, n = 2, ∆ = 10 cm, L = 12 cm, r = 6, N = 4800 rpm, l = 55 cm, t = 10 min, w = 45 kg, mf = 5 kg, CV = 45 MJ/kg, a = 10 kg /min

  1. Brake power,
    T = w × g × l = 45 × 9.81 × 55 × 10−2 = 242.79 N-m
  2. BP = pbm [(LA)N/n] × Number of cylinders
    pbm = 5.39 bar
  3. a = 10 kg /min or 600 kg /h

Example 11.2

A sharp-edged circular orifice of diameter 3.8 cm and coefficient of discharge as 0.6 is used to measure air-consumption of a four-stroke petrol engine. Pressure drop through the orifice is 145 mm of water and barometer reads 75.5 cm of Hg. The compression ratio of the engine is 6 and the piston displacement volume is 2000 cm3.

The temperature of air is taken to be 26°C. At 2600 rpm, the engine brake power recorded is 29.5 kW. The fuel consumption is 0.14 kg /min and the calorific value of fuel used is 43960 kJ/kg. Determine (a) the volumetric efficiency, (b) the air-fuel ratio, (c) the brake mean effective pressure, (d) the brake thermal efficiency, (e) the air standard efficiency, and (f) the relative efficiency.

Solution

Given that d = 3.8 cm, Cd = 0.6, hw = 14.5 cm of water, pa = 75.5 cm of Hg, r = 6, Vs = 2000 cm3, N = 2600 rpm, BP = 29.5 kW, f = 0.14 kg /min, CV = 43960 kJ/kg

Density of air at atmospheric conditions,

Head in m of air =

Velocity of air passing through orifice,

 

ca = [2gH]0.5 = [2 × 9.81 × 124.43]0.5 = 49.41 m /s

 

Volume of air passing through orifice,

= 0.033622 m3/s or 2.0173 m3/min
  1. Volumetric efficiency
  2. Mass flow rate of air,
    a = Vaρa = 2.0173 × 1.1653 = 2.35 kg /min
  3. BMEP =
  4. Brake thermal efficiency
  5. Air standard efficiency,
  6. Relative efficiency =

Example 11.3

A four-stroke SI engine has six single-acting cylinders of 7.5 cm bore and 9 cm stroke. The engine is coupled to a brake having a torque arm radius of 38 m. At 3300 rpm, with all cylinders operating the net brake load is 324 N. When each cylinder, in turn, is rendered in operative, the average net brake load produced at the same speed of the remaining five cylinders is 245 N. Estimate the indicated mean effective pressure engine.

Solution

BP when all cylinders are working =

= 42.55 kW

BP when each cylinder is cut-off in turn

IP of cylinder cut off = 42.55 − 32.17 = 10.38 kW

Total IP of engine = 6 × 10.38 = 62.3 kW

Indicated mean effective pressure

Example 11.4

A six-cylinder SI engine operates in a four-stroke cycle. The bore of each cylinder is 70 mm and the stroke 100 mm. The clearance volume per cylinder is 67 cm3. At a speed of 3960 rpm the fuel consumption is 19.5 kg /h and the torque developed is 140 N-m. Calculate (a) brake power, (b) brake mean effective pressure, (c) the brake thermal efficiency if LCV of fuel is 44000 kJ/kg, and (d) the relative efficiency on brake power basis. Assume γ = 1.4 for air.

Solution

  1. Brake power
  2. Brake mean effective pressure
  3. Brake thermal efficiency
  4. Swept volume per cylinder,

    Compression ratio,

    Air standard efficiency,

    Relative efficiency,

11.9 ❐ MEASUREMENT OF AIR CONSUMPTION BY AIR-BOX METHOD

The arrangement to measure the consumption of air by air-box method is shown in Fig. 11.12. It consists of an air-tight box fitted with a sharp-edged orifice of known coefficient of discharge. Due to the suction of engine, there is a pressure depression in the box which causes the flow through orifice for obtaining a steady flow. The pressure difference causing the flow through the orifice is measured with the help of a water manometer.

Let

A0 = area of orifice, m2

hw = head of water causing flow, cm

Cd = coefficient of discharge for orifice

d0 = diameter of orifice, cm

ρa = density of air, kg /m3

Head of air,

Velocity of air through the orifice,

Figure 11.12 Measurement of air by air box method

Volume of air passing through the orifice,

Mass flow rate of air passing through the orifice,

Density of air,

where hw is in cm, pa in bar and Ta in K

Mass of air supplied per kg of fuel used,

where N = % of N2 by volume in exhaust gases

C = % of carbon in fuel

C1, C2 = % of CO2 and CO by volume in exhaust gases.

11.10 ❐ MEASUREMENT OF BRAKE POWER

Brake power is measured with the following ways:

  1. Rope brake dynamometer: A rope brake dynamometer is shown in Fig. 11.13. A rope is wound around the brake drum, whose one end is connected to the spring balance S suspended from overhead and the other end carries the load W. In this arrangement, the whole power developed by the engine is absorbed by the friction produced at the rim of the brake drum. The rim of the brake drum is generally water-cooled to absorb the heat generated due to rubbing action of rope on rim.

    Let

    W = dead weight on the rope, N

    S = spring pull, N

    D = outer diameter of brake drum, m

    dr = diameter of rope, m

    N = speed of engine, rpm

    Net load acting on brake drum = (WS) N

    Effective radius of brake drum,

    Frictional torque rating on brake drum,

    Tf = (WS) × R Nm

    Figure 11.13 Rope brake dynamometer

  2. Prony brake: The arrangement of this braking system is shown in Fig. 11.14. It consists of brake shoes which are clamped on the rim of the brake drum by means of bolts. The pressure on the rim is adjusted with the help of nuts and springs. A load lever extends from top of the brake and a load carrier is attached to the end of the load lever. The weights kept on this load carrier are balanced by the reaction torque in the shoes. The load lever is kept horizontal to keep its length constant.

    Figure 11.14 Prony-brake arrangement

where

W = load on load carrier, N

L = distances from centre of shaft to load, m.

Example 11.5

A four-cylinder engine running at 1200 rpm gave 18.6 kW brake power. The average torque when one cylinder was cut off was 105 N.m. Determine the indicated thermal efficiency if the calorific value of the fuel is 42,000 kJ/kg and the engine uses 0.34 kg of petrol per kWh of brake power.

Solution

Given that i = 4, N = 1200 rpm, (BP)n = 18.6 kW, T3 = 105 N.m, CV = 42,000 kJ/kg

BSFC = 0.34 kg /kWh

(BP)n − (BP)n1 = (IP)n − (IP)n1 = (IP)1

∴ (IP)1 = 18.6 − 13.2 = 5.4 kW

Indicated power of 4 cylinders,

(IP)n = i × (IP)1 = 4 × 5.4 = 21.6 kW

Fuel consumption,

Indicated thermal efficiency,

Example 11.6

A single cylinder four-stroke oil engine works on a diesel cycle. The following readings correspond to full load conditions:

Area of indicator diagram = 3 cm2, Length of the diagram = 5 cm, Spring constant = 12 bar/cm2. cm, Speed of the engine = 500 rpm, Load on the brake = 400 N, Spring reading = 50 N

Diameter of brake drum = 120 cm, Fuel consumption = 2.75 kg /h, Calorific value of fuel = 42000 kJ/kg, Diameter of cylinder = 16 cm, Stroke length of piston = 20 cm. Determine (a) the friction power, (b) the mechanical efficiency, (c) the brake thermal efficiency, and (d) the brake mean effective pressure.

Solution

Given that n = 2, Ai = 3 cm2, Li = 5 cm, C = 12 bar/cm2. cm, N = 500 rpm, W = 400 N, S = 50 N, Db = 120 cm, f = 2.75 kg /h, CV = 42,000 kJ/kg, d = 16 cm, L = 20 cm

  1. Frictional power,
    FP = IP − BP = 12.06 − 10.99 = 1.07 kW
  2. Mechanical efficiency
  3. Brake thermal efficiency
  4. Brake mean effective pressure

Example 11.7

The following data are known for a four-cylinder four-stroke petrol engine: cylinder dimensions: 11 cm bore and 13 cm stroke, engine speed: 2250 rpm, brake power: 50 kW, friction power: 15 kW, fuel consumption rate: 10.5 kg /h, calorific value of fuel: 50,000 kJ/kg, air inhalation rate: 300 kg /h, and ambient condition: 15°C, 1.03 bar. Estimate (a) brake mean effective pressure, (b) volumetric efficiency, (c) brake thermal efficiency, and (d) mechanical efficiency.

[IES, 2008]

Solution

Given that D = 11 cm, L = 13 cm, N = 2250 rpm, BP = 50 kW, FP = 15 kW, f = 10.5 kg /h. CV = 50,000 kJ/kg, a = 300 kg /h, ta = 15°C, pa = 1.03 bar, i = 4

  1. or

  2. Volumetric efficiency

    Swept mass of air per second

  3. Brake thermal efficiency,
  4. Indicated power,
    IP = BP + FP = 50 + 15 = 65 kW

    Mechanical efficiency

Example 11.8

The following data relates to a four-cylinder, four-stroke petrol engine:

Diameter of piston = 80 mm, length of stroke = 120 mm

Clearance volume = 100 × 103 mm3

Fuel supply = 4.8 kg /h. calorific value = 44100 kJ/kg

When the Morse test was performed the following data were obtained:

BP with all cylinders working = 14.5 kW

BP with cylinder 1 cut-off = 9.8 kW

BP with cylinder 2 cut-off = 10.3 kW

BP with cylinder 3 cut-off = 10.14 kW

BP with cylinder 4 cut-off = 10 kW

Find the IP of the engine and calculate indicated thermal efficiency, brake thermal efficiency, and relative efficiency.

Solution

Given that i = 4, n = 2, d = 80 mm, L = 120 mm, Vc = 100 × 103 mm3, f = 4.8 kg /k, CV = 44100 kJ/kg, (BP)n = 14.5 kW, (BP)1 = 9.8 kW, (BP)2 = 10.3 kW, (BP)3 = 10.14 kW, (BP)4 = 10 kW

(IP)1 = (BP)n − (BP)n1

∴     (IP)1 = 14.5 − 9.8 = 4.7 kW

(IP)2 = 14.5 − 10.3 = 4.2 kW
(IP)3 = 14.5 − 10.14 = 4.36 kW
(IP)4 = 14.5 − 10 = 4.5 kW

Indicated power,

Indicated thermal efficiency,

Brake thermal efficiency,

Swept volume,

Compression ratio,

Air standard efficiency,

Relative efficiency,

Example 11.9

Find the air-fuel ratio of a four-stroke, single-cylinder, air-cooled engine with a fuel consumption time for 10 cc as 20.4 s and air consumption time for 0.1 m3 as 16.3 s. The load is 17 kg at the speed of 3000 rpm. Find the brake specific fuel consumption and brake thermal efficiency. Assume the density of air as 1.175 kg /m3 and specific gravity of fuel to be 0.7. The lower heating value of fuel is 43MJ/kg and dynamometer constant is 5000.

Solution

Given that W = 17 kg, N = 3000 rpm, ρa = 1.175 kg /m3, sf = 0.7, LCV = 43 MJ/kg, C = 5000

f = V̇f ρf = V̇f × sf × ρw = 0.49 × 10−6 × 0.7 × 103 = 0.343 × 10−3 kg /s

a = V̇a ρa = 6.135 × 10−3 × 1.175 = 7.208 × 10−3 kg /s

A /F ratio =

Brake power,

Brake specific fuel consumption,

Brake thermal efficiency,

Example 11.10

The following data refers to a test on a single cylinder oil engine working on four-stroke cycle:

Diameter of brake wheel = 60 cm; rope diameter = 3 cm; dead load is 25 kg, spring balance reading = 5 kg, and the engine is running at 400 rev/min. The indicator diagram has area = 4 cm2, length = 6 cm, and spring stiffness is 12 bar/cm. The fuel consumption is 0.23 kg /kWh and the fuel used has a calorific value 43963.5 kJ/kg. Taking cylinder bore 10 cm and piston stroke 15 cm, calculate the brake power, indicated power, mechanical and indicated thermal efficiencies of the engine.

Solution

Given that Db = 0.6 m, dr = 0.03 m, W = 25 × 9.81 = 245.25 N, S = 5 × 9.81 = 49.05 N, N = 400 rpm, n = 2, a2 = 4 cm2, li = 6 cm, k = 12 bar/cm, f = 0.23 kg /kWh, CV = 43963.5 kJ/kg, D = 10 cm, L = 15 cm

Indicated mean effective pressure,

Indicated power,

Mechanical efficiency,

Total fuel consumption, = BP × f

= 2.589 × 0.23 = 0.5955 kg /h

Indicated thermal efficiency,

Example 11.11

A four-stroke, single cylinder petrol engine mounted on a motor cycle was put to load test. The load measured on dynamometer was 30 kg with drum diameter and speed, respectively, at 900 mm and 2000 rpm. The engine consumed 0.15 kg of fuel in one minute, the calorific value of fuel being 43.5 MJ/kg. The fuel supply to the engine was stopped and was driven by a motor which needed 5 kW of power to keep it running at the same speed, the efficiency of the motor being 80%. The engine cylinder bore and stroke are respectively at 150 mm and 200 mm. Calculate (a) the brake power, (b) the indicated power, (c) the mechanical efficiency, (d) the brake thermal efficiency, (e) the indicated thermal efficiency, (f) the brake mean effective pressure and (vii) the indicated mean effective pressure.

[IES, 2012]

Solution

Given that four-stroke single cylinder petrol engine, W = 30 kg, Dd = 900 mm, N = 2000 rpm, f = 0.15 kg /min, CV = 43.5 MJ/kg, Pm = 5 kW, ηm = 80%, D = 150 mm, L = 200 mm, n = 2.

  1. Torque applied,

    Break power,

  2. Indicated power,
    IP = BP + FP

    Frictional power,

    Indicated power, IP = 27.737 + 6.25 = 33.987 kW

  3. Mechanical efficiency,
  4. Brake thermal efficiency,
  5. Indicated thermal efficiency,
  6. BMEP, Pbm = 4.7 bar

Example 11.12

During the trial of a single oil engine, cylinder diameter 20 cm, stroke 28 cm, working on the two-stroke cycle, and firing every cycle, the following observations were made:

Duration of trial = 1 h, total fuel used = 4.22 kg, calorific value of fuel = 44670 kJ/kg, proportion of hydrogen in fuel = 15%, total number of revolutions = 21,000, mean effective pressure = 2.74 bar, net brake load applied to a drum of 100 cm diameter = 600 N, total mass of cooling water circulated = 495 kg, temperature of cooling water: inlet 13°C, outlet 38°C, air used = 135 kg, temperature of air in test room = 20°C, temperature of exhaust gases = 370°C. Assume cp of gases = 1.005 kJ/kg.K, cp of steam = 2.093 kJ/kg.K at atmospheric pressure. Calculate the thermal efficiency and draw up the heat balance sheet.

[IES, 1997]

Solution

Given that D = 20 cm, L = 28 cm, n = 1, t = 1 h, mf = 4.22 kg, CV = 44670 kJ/kg, H2 = 15%, Nt = 21000 rev, pm = 2.74 bar, W = 600 N, d = 100 cm, mw = 495 kg, twi = 13°C, two = 38°C, me = 13.5 kg, tr = 20°C, tg = 370°C, cpg = 1.005 kJ/kg. K, cps = 2.093 kJ/kg.K

Torque, T = W × d/2 = 600 × 0.5 = 300 N.m

Brake power,

Indicated power,

Thermal efficiency,

Swept volume,

Heat input,

Heat equivalent of BP, Q1 = BP × 60 = 10.996 × 60 = 659.76 kJ/min

Heat lost to cooling water,

Heat carried away by exhaust gases:

Mass of flue gases =

Steam in exhaust gases

Mass of dry exhaust gases g = 2.32 − 0.095 = 2.225 kg /min

Heat in steam in exhaust gases, Q3 = s [hg + cps (tgtsat) − hfw] [∴ tsup = tg] at patm = 1.013 bar

= 0.095 [2676 + 2.093 (370 − 100) − 4.187 × 20] [ ∵ hfw = cpw × tr]
= 300 kJ/min

Heat in dry exhaust gases = g × cpg (tgtr)

= 2.225 × 1.005 × (370 − 20) = 782.6 kJ/min

Unaccounted heat loss = 3141.79 − [659.76 + 862.125 + 300 + 782.6]

= 537.305 kJ/min

Heat Balance Sheet:

Example 11.13

During a trial of a single cylinder, a four-stroke diesel engine, the following observations were recorded:

Bore = 340 mm, stroke = 440 mm, rpm = 400, area of indicator diagram = 465 mm2, length of diagram = 60 mm, spring constant = 0.6 bar/mm, load on hydraulic dynamometer = 950 N, dynamometer constant = 7460, fuel used = 10.6 kg /h, calorific value of fuel (CV) = 49500 kJ/kg, cooling water circulated = 25 kg /min, rise in temp. of cooling water = 25°C, mass analysis of fuel: carbon = 84%, hydrogen = 15%, incombustible = 1%, volume analysis of exhaust gas: carbon dioxide = 9%, oxygen = 10%, temperature of exhaust gases = 400°C, specific heat of exhaust gas = 1.05 kJ/kg°C, partial pressure of steam in exhaust gas = 0.030 bar, specific heat of superheated steam = 2.1 kJ/kg°C, saturation temp. of steam at 0.030 bar = 24.1°C. Draw up heat balance sheet on minute basis.

[IAS, 2011]

Solution

Given that n = 2, D = 0.34 m, L = 0.44 m, N = 400 rpm, Ai = 465 mm2, Ld = 60 mm, k = 0.6 bar/mm, W = 950 N, C = 7460, CV = 49500 kJ/kg, f = 10.6 kg /h, w = 25 kg /min, ∆tw = 25°C, C = 84%, H2 = 15%, incombustible = 1%, CO2 = 9%, O2 = 10%, teg = 400°C, cpg = 1.05 kJ/kg /°C, ps = 0.030 bar, cps = 2.1 kJ/kg°C, ts at 0.030 bar = 24.1°C

Brake power,

Indicated mean effective pressure, pim =

Indicated power, IP =

= 61.92 kW
    1. Heat equivalent of BP, Qb = BP × 60 = 50.938 × 60 = 3056.28 kJ/min
    2. Heat lost to cooling water, Qw = wcpw ⋅ Δtw = 25 × 4.187 × 25 = 2616.87 kJ/min

      Percentage of N2 in exhaust gas = 100 − (9 + 10) = 81%

      Mass of air supplied per kg of fuel,

      Mass of exhaust gases formed per kg of fuel = 1 + 22.91 = 23.91 kg

      Mass of exhaust gases formed per minute =

      Mass of steam formed and carried with the exhaust gases per minute due to the combustion of hydrogen in the fuel,

      Mass of dry exhaust gases formed per minute, g = 4.224 − 0.2385 = 3.9855 kg /min

    3. Heat carried away by exhaust gases per minute, Qg = g cpg (tg − ta)

      Let ambient air temperature, ta = 27°C

      Qg = 3.9855 × 1.05 (400 − 27) = 1560.92 kJ/min
    4. Heat carried away by steam with exhaust gases per minute,
      Qs = s [hg+ cps (Tsup − 100)]

      At p = 0.030 bar, ts = 24.10°C, hg = 2545.5 kJ/kg

      Qs = 0.2385 [2545.5 + 2.1 (400 − 100)] = 757.36 kJ/min
  1. Heat unaccounted for = Qi − (Qb + Qw + Qg + Qs)
    = 8745 − (3056.28 + 2616.87 + 1560.92 × 757.36) = 753.57 kJ/min

Heat Balance Sheet:

Example 11.14

During a test on a two-stroke engine on full load, the following observations were recorded:

Speed =350 rpm, net brake load = 590 N, mean effective pressure = 2.8 bar, fuel oil consumption = 4.3 kg /h, cooling water required = 500 kg /h, rise in cooling water temperature = 25°C, air used per kg of fuel = 33 kg, room temperature = 25°C, exhaust gas temperature = 400°C, cylinder diameter = 220 mm, stroke length = 280 mm, effective brake diameter = 1 m, C.V. of fuel oil = 43900 kJ/kg, proportion of hydrogen in fuel = 15%, mean specific heat of exhaust gases = 1.0 kJ/kg-K, specific heat of steam = 2.09 kJ/kg-K

Calculate the following:

  1. Indicated power
  2. Brake power
  3. Draw the heat balance sheet on the basis of kJ/min

    [IAS, 2010]

Solution

Given that N = 350 rpm, Wb = 590 N, pmi = 2.8 bar, f = 4.3 kg /h, (∆t)w = 25°C, w = 500 kg /h, ma /mf = 33, t0 = 25°C, tg = 400°C, D = 220 mm, L = 280 mm, db = 1 m, CV = 43900 kJ/kg, H2 = 15%, cpg = 1.0 kJ/kg.K, cps = 2.09 kJ/kg.K

  1. Indicated power.
  2. Brake power,
  3. Heat balance sheet:

    Heat input =

    Heat equivalent of BP = 10.81 × 60 = 648.6 kJ/min

    Heat lost to cooling water = w × cpw × (Δt)w =

    Heat carried away by exhaust gases:

    Mass of flue gases =

    Steam in exhaust gases,

    Mass of dry exhaust gases, g = 2.437 − 0.09675 = 2.340 kg /min

    hf = cpw (100 − t0)

    Heat in steam in exhaust gases = s [hf + hfg + cps (tg − 100)]

    = 0.09675 [4.187 (100 − 25) + 2256.9 + 2.09 (400 − 100)]
    = 309.4 kJ/min

    Heat in dry exhaust gases = g × cpg (tg − 100) = 2.340 × 1.0 (400 − 100) = 702 kJ/min

    Unaccounted heat loss = 3146.2 − (648.6 + 872.3 + 309.4 + 702) = 613.9 kJ/min

Example 11.15

A four-stroke petrol engine develops 30 kW at 2600 rpm. The compression ratio of the engine is 8 and its fuel consumption is 8.4 kg /h with calorific value of 44 MJ/kg. The air consumption of the engine as measured by means of a sharp-edged orifice is 2 m3 per min. If the piston displacement volume is 2 litres, calculate (a) the volumetric efficiency, (b) the air-fuel ratio, (c) the brake mean effective pressure, (d) the brake thermal efficiency, and (e) the relative efficiency.

The ambient temperature air can be taken as 27°C, R for air as 287 J/kgK, and g = 1.4. The barometer reads 755 mm of mercury.

[IAS, 2006]

Solution

Given that n = 2, BP = 30 kW, N = 2600 rpm, r = 8, f = 8.4 kg /h, CV = 44 MJ/kg, a= 2m3/min, Vs = 2 litres = 2 × 10−3 m3, T0 = 27 + 273 = 300 K, R = 287 J/kg.K, γ = 1.4, pb = 755 mm of Hg

  1. Swept volume per minute,

    Volumetric efficiency

  2. a = a × 60 × ρa = 2 × 60 × 1.169 = 140.28 kg /h
  3. Brake power BP =

    or

    or ps = 0.6923 MPa

  4. Brake thermal efficiency = 0.2922 or 29.22%
    ηa = 1− (1/rγ −1) = 1 − 1/(0.8)0.4 = 0.5647 or 56.47%
  5. Relative efficiency

Example 11.16

A six-cylinder, four-stroke petrol engine has a swept volume of 3.0 litres with a compression ratio of 9.5. Brake output torque is 205 N-m at 3600 rpm. Air enters at 105 N/m2 and 60°C. The mechanical efficiency of the engine is 85% and air-fuel ratio is 15:1. The heating value of fuel is 44,000 kJ/kg and the combustion efficiency is 97%. Calculate (a) the rate of fuel flow, (b) the brake thermal efficiency, (c) the indicated thermal efficiency, (d) the volumetric efficiency, and (e) the brake specific fuel consumption.

[IAS, 2005]

Solution

Given that i = 6, n = 2, Vs = 3 litre = 3 × 10−3 m3, r = 9.5, Tb = 205 N.m, N = 3600 rpm, p = 105 N/m2, T = 60 + 273 = 333 K, ηmech = 85%, A /F = 15, CV = 44000 kJ/kg, ηcomb = 97%

Brake power,

Indicated power,

Air standard efficiency,

Density of air,

Air volume flow rate,

Mass of air taken in per second = 0.09 × 1.046 = 0.09414 kg /s

  1. Mass flow rate of fuel for the engine,
  2. Brake thermal efficiency,
    = 0.2885 or 28.85%
  3. Indicated thermal efficiency,
  4. Relative efficiency,

Example 11.17

The air flow to a four cylinder, four-stroke engine is measured by means of a 4.5 cm diameter orifice, having Cd = 0.65. During a test the following data was recorded:

Bore = 10 cm, stroke = 15 cm, engine speed = 1000 rpm, brake torque = 135 Nm, fuel consumption = 5.0 kg /h, CVfuel = 42600 kJ/kg, head across orifice = 6 cm of water.

Ambient temperature and pressure are 300 K and 1.0 bar, respectively.

Calculate (a) the brake thermal efficiency, (b) the brake mean effective pressure, and (c) the volumetric efficiency.

[IAS, 2004]

Solution

Given that i = 4, n = 2, d0 = 4.5 cm, Cd = 0.65, D = 10 cm, L = 15 cm, N = 1000 rpm, Tb = 135 Nm, f = 5 kg /h, CV = 42600 kJ/kg, h = 6 cm of H2O, T0 = 300 K, p0 = 1 bar, R = 287 J/kg.K

  1. Brake power,

    Brake thermal efficiency,

  2. Brake mean effective pressure,

  3. Δp = ρgh = 103 × 9.81 × 6 × 10−2 = 588.6 N/m2

    Air inhaled per second,

    Swept volume/s,

    Volumetric efficiency,

Example 11.18

An eight-cylinder automobile engine of 80 mm diameter and 90 mm stroke with a compression ratio of 7, is tested at 4000 rpm on a dynamometer of 600 mm arm length. During a 10-minute test period at a dynamometer scale reading of 450 N, 4.8 kg of gasoline having a calorific value of 45000 kJ/kg was burnt and air at 27°C and 1.0 bar was supplied to the carburettor at the rate of 6.6 kg /min. Find (a) the brake power delivered, (b) the brake mean effective pressure, (c) the brake specific fuel consumption, (d) the brake thermal efficiency, (e) the volumetric efficiency, and (f) the air fuel ratio.

[IAS, 2003]

Solution

Given that i = 8, D = 80 mm, L = 90 mm, r = 7, N = 4000 rpm, l = 600 mm, t = 10 min, W = 450 N, mf = 4.8 kg, CV = 45000 kJ/kg, T1 = 27 + 273 = 300 K, p1 = 1 bar, a = 6.6 kg /min

Density of air,

  1. Torque on the dynamometer, T = Wl = 450 × 0.6 = 270 N.m

    Brake power,

  2. or

    or pbm = 9.375 bar

  3. Brake specific fuel consumption
  4. Brake thermal efficiency,
  5. Swept volume,

    Swept volume/s,

    Air inhaled: p1 = aRT1

    Volumetric efficiency,

Example 11.19

A six-cylinder, four-stroke petrol engine, with a bore of 120 mm and stroke of 180 mm under test, is supplied petrol of composition: C = 82% and H2 = 18% by mass. The Orsat gas analysis indicated that CO2 = 12%, O2 = 4% and and N2 = 84% by volume. Determine (a) the air-fuel ratio and (b) the percentage of excess air.

Calculate the volumetric efficiency of engine based on intake conditions when the mass flow rate of petrol is 32 kg /min at 1600 rpm. Intake condition are 1 bar and 17°C. Consider the density of petrol vapour to be 3.5 times that of air at same temperature and pressure. Air contains 23% oxygen by mass.

[IAS, 2003]

Solution

Given that i = 6, n = 2, D = 120 mm, L = 180 mm, C = 82%, H2 = 18%, CO2 = 12%, O2 = 4%, N2 = 84%, f = 32 kg /h, N = 1600 rpm, p0 = 1 bar, T0 = 17 + 273 = 290 K, ρfa = 3.5

Minimum mass of air required for complete combustion

Mass of air supplied,

Excess air supplied = 17.394 − 15.768 = 1.626 kg /kg of fuel

Percentage of excess air =

Actual A/F ratio = 17.394:1 kg air/kg of fuel

Mass flow rate of petrol = 32 kg /h

Swept volume per second =

Volume of air,

Volume of petrol vapour,

Total volume = 14.477 + 2.913 = 17.39 m3/kg fuel

Mixture aspirated per minute =

Volumetric efficiency,

Example 11.20

Two identical petrol engines having the following specifications are used in vehicles:

Engine 1: Swept volume = 3300 cc. normally aspirated, BMEP = 9.3 bar, rpm ≃ 4500, compression ratio = 8.2, efficiency ratio = 0.5, mechanical efficiency ≃ 0.9, mass of the engine ≃ 200 kg

Engine 2: Supercharged, swept volume = 3300 cc, BMEP ≃ 12.0 bar, rpm = 4500, compression ratio = 5.5, efficiency ratio = 0.5, mechanical efficiency = 0.92, mass of the engine ≃ 220 kg

If both the engines are supplied with just adequate quantity of petrol for the test run, determine the duration of test run so that the specific mass per kW of brake power is same for both the engines. Calorific value of petrol = 44000 kJ/kg.

Assume both the engines operate on four stroke cycle.

[IAS, 2010]

Solution

Engine 1

Vs = 3300 cc or 33 × 10−4 m3, pbm = 9.3 bar, N = 4500 rpm, n = 2, r = 8.2, ηr = 0.5, ηmech = 0.9, M1 = 200 kg, LCV = 4400 kJ/kg

Brake power,

Indicated power, IP = BP/ηmech = 115/0.9 = 127.8 kW

Air standard efficiency,

Efficiency ratio,

or    ηth = 0.5 × 0.569 = 0.2845

Again,     IP = mf 1 × CV × (ηth)i

or    127.8 = mf 1 × 44000 × 0.2865

or    mf 1 = 10.96 × 10−3 kg /s

= 10.96 ×10−3 × t × 3600 = 39.46 × t kg /h

Specific mass of engine per kW of

Engine 2:

Vs = 3300 cc or 33 × 10−4 m3, pbm = 12 bar, N = 450 rpm, n = 2, r = 5.5, ηr = 0.5, ηmech = 0.92, M2 = 220 kg

Air standard efficiency,

(ηth)i = 0.494 × 0.5 = 0.247
IP = mf 2 × CV × (ηth)i

or    161.3 = mf 2 × 44000 × 0.247

or    mf 2 = 14.84 × 10−3 kg /s

= 14.84 × 10−3 × t × 3600 = 53.43 × t kg /h

Specific mass of engine per kW of

Now

or    1.2904 (39.46 × t + 200) = 53.43 × t + 220

or   38.08 = 2.511 × t

or   t = 15.1 hours

Example 11.21

The following data relates to a test trial of a single-cylinder four-stroke engine:

Cylinder dia. = 24 cm, stroke length = 48 cm, compression ratio = 5.9, number of explosions/minute = 77, gas used/min at 771 mm of mercury and 15°C = 0.172 m3, lower calorific value of gas at NTP = 49350 kJ/m3, mean effective pressure from indicator card = 7.5 bar, weight of Jacket cooling water/min = 11 kg, temperature rise of cooling water = 34.2°C, specific heat of water = 4.2 kJ/kg°C.

Net brake load applied at brake wheel having an effective circumference of 3.86 m is 1260 N at average speed of 227 rpm.

Estimate (a) the mechanical efficiency, (b) the indicated thermal efficiency and (c) the efficiency ratio, and draw a heat balance sheet for the engine assuming that exhaust gases carry away 24% of heat.

[IAS, 2000]

Solution

Given that D = 24 cm, L = 48 cm, r = 5.9, Wbn = 1260 N, = 0.6143 m, Nm = 227 rpm, Ne = 77/min, g = 0.217 m3/min at 15°C and 771 mm of Hg, LCV = 49350 kJ/m3 at NTP, pmi = 7.5 bar, w = 11 kg /min, (Δt)w = 34.2°C, cpw = 4.2 kJ/kg.°C

  1. Brake torque, Tb = Wbn × rb = 1260 × 0.6143 = 774 N.m

    Brake power,

    Indicated power,

    Mechanical efficiency,

  2. Indicated thermal efficiency,

    Volume of gas at NTP = 0.20873 m3/min

    Indicated thermal efficiency

  3. Air standard efficiency, ηa =

    Heat balance sheet:

    Heat input = 0 × LCV = 0.20873 × 49350 = 10300.82 kJ/min

    Heat equivalent of BP = 18.4 × 60 = 1104 kJ/min

    Heat lost to cooling water = w × cpw × (Δt)w

    = 11 × 4.2 × 34.2 = 1580.04 kJ/min

    Heat lost to exhaust gases = 0.24 × 10,300.82 = 2472.20 kJ/s

    Heat unaccounted for = 10,300.82 − (1104 + 1580.04 + 2472.20) = 5144.58 kJ/min

11.11 ❐ SUPERCHARGING OF IC ENGINES

The power output of an engine depends on the amount of air inducted into the cylinder per unit time, the degree of utilisation of this air, and the thermal efficiency of the engine. The amount of air inducted per unit time can be increased by increasing the engine speed or by increasing the density of air at intake. The volumetric efficiency decreases as the speed is increased. The method of increasing the inlet air density is called supercharging. It is generally used to increase the power output of the engine. It is done by supplying air at a pressure higher than the pressure at which the engine naturally aspirates air from the atmosphere by using a pressure boosting device called a supercharger.

11.11.1 Thermodynamic Cycle

The p-v diagram for an ideal Otto-cycle supercharged engine is shown in Fig. 11.15(a). The pressure p1 represents the supercharging pressure and p6 is the exhaust pressure. Area 8-6-7-0-1-8 represents the work done by the supercharger in supplying air at a pressure p1, whereas the area 1-2-3-4-1 is the output of the engine. Area 0-1-6-7-0 represents the gain in work during the gas exchange process due to supercharging. Thus, a part of the supercharger work is recovered. However, the area 1-6-8-1 cannot be recovered and represents a loss of work. This loss of work causes the ideal thermal efficiency of the supercharged engine to decrease with an increase in supercharging pressure.

Figure 11.15(b) shows an ideal dual combustion cycle supercharged engine. The pressure p1 represents the supercharging pressure and p6 = p7, is the exhaust pressure. The engine is supercharged by the compressor. Area 9-10-1-11-9 represents the work done on the supercharger in supplying air at pressure p1. Thus, for the engine, 8-1 represents induction process, 1-2 as the compression process, 2-3-4 as the heat addition process, 4-5 as the expansion process, 5-6 as the blow down to atmosphere or heat rejection process and 6-7 as the exhaust process.

Figure 11.15 p-v diagram for supercharged engine: (a) Ideal Otto cycle, (b) Dual combustion cycle

Work done my mass m by the supercharged engine,

W = Area (1-2-3-4-5-1) + Area (8-1-6-7-8) — Area (9-10-1-11-9).

Thus, a part of the supercharger work is recovered; however, the curve equivalent to Area (9-10-1-11-9) — Area (8-1-6-7-8) is not recoverable and represents loss of work. This loss of work causes the ideal thermal efficiency of supercharged engine to decrease with an increase in supercharging pressure.

11.11.2 Supercharging of SI Engines

As far as SI engines are concerned, supercharging is employed only for aircraft and racing car engines. This is because the increase in supercharging pressure increases the tendency to detonate and pre-ignite.

Apart from increasing the volumetric efficiency of the engine, supercharging results in an increase in the intake temperature and pressure of the engine. These reduce the ignition delay and increase the flame speed and results in greater tendency to detonate or pre-ignite. For this reason, the supercharged SI engines use a lower compression ratio which results in lower thermal efficiency. The fuel consumption is also greater than naturally aspirated engines.

Due to its poor fuel economy, supercharging of petrol engines is not very popular and is used only when more power is needed or when more power is need to compensate altitude loss.

11.11.3 Supercharging of CI Engines

Supercharging of CI engines does not result in any combustion problems. Increase in pressure and temperature of the intake air reduces ignition delay and hence the rate of pressure rise results in a better, quieter, and smoother combustion. This allows the use of poor quality fuel. The increase in intake air temperature reduces volumetric and thermal efficiencies but the increase in density due to pressure compensates for this and intercooling is not required except for highly supercharged engines. It is possible to use lower fuel-air ratios in a supercharged engine, resulting in lower temperature and reduced smoke from the engine. This results in an increased life of the engine. Thus, a CI engine is more suitable for supercharging.

The apparatus used for increasing air density is known as a supercharger. A supercharger is an air compressor which may be positive displacement, reciprocating (piston-cylinder), rotary type (roots blower), or rotodynamic compressors (centrifugal or axial flow type). The centrifugal compressor is widely used as a supercharger for IC engines. The function of a supercharger is either to produce more power from an engine of a given cylinder size, or to compensate the power loss at high altitudes due to rarefied atmosphere.

11.11.4 Effects of Supercharging

The effect of supercharging on the power and efficiency is shown in Fig. 11.16.

  1. Power output: Supercharging produces more power because the supercharger supplies air at a pressure and density higher than atmospheric with has the effect of increasing volumetric efficiency.
  2. Mechanical efficiency: The mechanical efficiency of a supercharged engine is higher than that of one not supercharged.
  3. Fuel consumption: It provides better mixing of fuel and air, which results in a specific reduction of fuel consumption and the thermal efficiency increases.

Figure 11.16 Effect of supercharging ratio on power and efficiency

11.11.5 Objectives of Supercharging

The objectives of supercharging are as follows:

  1. To overcome the effect of high altitudes, as in the case of aircrafts and stationary engines in mountains.
  2. To reduce the weight of an engine per kW or power developed, as in the case of racing cars.
  3. To reduce the size of the engine to fit into a limited space, as in the case of locomotives or marine engines.
  4. To increase the power output of an existing engine to meet greater power demands.

11.11.6 Configurations of a Supercharger

  1. Compressor Coupled of Engine Shaft: The compressor is operated directly by the engine with set-up gearing to increase the speed of centrifugal compressor. A part of the engine output is used to drive the supercharger and the net output is calculated by deducing this power from the gross power of the engine.
  2. Turbo-charger: The energy of the exhaust gases from the engine is used to develop power in a turbine which directly runs the supercharger. There is no mechanical connection between the engine and the supercharger.
  3. Direct Coupling between the Engine, Compressor, and Turbine: The advantage of this arrangement is that when the turbine output is insufficient to run the compressor, additional power required is taken from the engine. Additional power from the turbine can also be fed to the engine.
  4. Compressor geared with Engine and Free turbine: The engine drives the compressor but the energy in the exhaust is utilised to develop the power from a separate turbine.

The various configurations are shown in Fig. 11.17(a) to (d).

Figure 11.17 Arrangements for supercharging: (a) Mechanical supercharging, (b) Turbocharging, (c) Engine-driven compressor and turbocharger, (d) Engine-driven compressor and free turbine

11.11.7 Supercharging of Single Cylinder Engines

Supercharging of single cylinder engines is not carried out because the thermal efficiency drops to zero when the ratio of supercharger is about 6. Further, the power is maximum when the ratio of the supercharger is nearly 2, 5, and then drops to zero at a ratio of about 6.

Example 11.22

A four-stroke diesel engine of 3000 cc capacity develops 14 kW per m3 of free air induced per minute. When running at 3500 rev/min, it has a volumetric efficiency of 85% referred to free air-conditions of 1.013 bar and 27°C. It is proposed to boost the power of the engine by supercharging by a blower (driven mechanically from the engine) of pressure ratio 1.7 and isentropic efficiency of 80%. Assuming that at the end of induction, the cylinders contain a volume of charge equal to the swept volume, at the pressure and temperature of the delivery from the blower, estimate the increase in brake power to be expected from the engine. Take overall mechanical efficiency as 80%, γ for air = 1.4, R = 0.287 kJ/kg K.

[IES, 2009]

Solution

The schematic arrangement is shown in Fig. 11.18(a).

Swept volume of engine per minute,

Unsupercharged inducted volume

= Vs × ηvol = 5.25 × 0.85 = 4.4625 m3/min

The actual and isentropic compression processes of supercharger are shown in Fig. 11.18(b).

Isentropic efficiency of supercharger.

Figure 11.18 Supercharger: (a) Schematic arrangement, (b) Processes on T-s diagram

The supercharger delivers 5.25 m3/min at p2 = 1.013 × 1.7 = 1.72 bar and 361.4 K to the engine. This volume referred to 1.013 bar and 300 K is,

Increase in inducted volume = VVs = 7.41 − 4.4625 = 2.95 m3/min

As indicated power (I.P.) is directly proportional to induced volume, therefore, increase in I.P. due to increase in inducted volume of air

= 14 × 2.95 = 41.3 kW

Increase in IP due to increase in inducted air pressure because of supercharger

Total increase in IP because of supercharger

= 41.3 + 6.186 = 47.486 kW

Increase in BP = Increase in IP × ηmech

= 47.486 × 0.8 = 38 kW

Power required to run the supercharger.

Psup = a ⋅ cpa (T2′ − T1)

where a = mass of air delivered by the supercharger per second.

Psup = 0.145 × 1.005 × (361.4 − 300) = 8.95 kW

Net increase in BP = 38 − 8.95 = 29.05 kW

Percentage increase in

11.12 ❐ SI ENGINE EMISSIONS

Spark ignition engine emissions are divided into three categories as exhaust emission, evaporative emission, and crank case emission.

Figure 11.19 Emissions from SI engines

The major constituents which contribute to air pollution are CO, NOx, and hydrocarbons (HC) coming from SI engine exhaust. The percentages of different constituents coming out from the above three mentioed sources are shown in Fig. 11.19.

The relative amounts depend on the engine design and operating conditions but are of order, NOx → 500 to 1000 ppm (20 g /kg of fuel), CO → 1 to 2% (200 g /kg of fuel) and HC → 3000 ppm(25 g /kg of fuel). Fuel evaporation from the fuel tank and the carburettor exists even after engine shut down and these are unburned HCs. However, in most modern engines, these non-exhaust unburned HCs are effectively controlled by returning the blow-by gases from the crank case to the engine intake system by venting the fuel tank and through a vapour-absorbing carbon cannister which is purged as some parts of the engine intake air during normal engine operation.

The other constituents include SO2 and lead compounds. Petrol rarely contains sulphur; therefore, SO2 is not a pollutant from the SI engine exhaust. Petrol contains lead in small percentages but its effect is more serious on human health.

The processes by which pollutants form within the cylinder in a conventional SI engine are qualitatively illustrated in Fig. 11.19. It shows the formation of pollutants during four strokes of the cycle. NO forms throughout the high temperature burned gases behind the flame through chemical reactions. NO formation rate increases with an increase in gas temperature. As the burned gases cool, during expansion stroke, the reactions involving NO freeze and leave NO concentrations far in excess of levels corresponding to equilibrium temperature at exhaust conditions.

CO also forms during combustion process with lean A:F mixtures, and there is sufficient O2 to burn all the carbon in the fuel to CO2. However, in high temperature products even with lean mixtures, there is sufficient CO in exhaust because of dissociation of CO2. Later, in expansion stroke, the CO oxidation process also freezes as the gas temperature falls.

The unburned hydrocarbon emission comes from different sources. During compression and combustion, the increasing cylinder pressure forces some of the gases in the cylinder into crevices connected to combustion chamber, the volumes between the piston rings and cylinder wall are the largest of these. Most of this gas entering into crevices is unburned air fuel mixture escaped from primary combustion zone. This happens because the flame cannot enter these narrow crevices. The gas which leaves these crevices later in the expansion and exhaust processes is one source of unburned HC emissions. The combustion chamber walls are another possible source. A quench layer containing unburned and partially burned A:F mixture is left at the wall when the flame dies as it approaches the wall. This unburned HC in this layer (0.1 mm) burns rapidly if the combustion chamber walls are clean. The next source of HC is the thin layer of lubricating oil on cylinder wall, and the piston which absorbs HC before and after combustion. A final source of HC in engines is incomplete combustion due to bulk quenching of the flame in that fraction of engine cycle where the combustion is especially slow. This unburned HC near the cylinder wall is exhausted during exhaust stroke as the piston pushes the gases out.

11.12.1 Exhaust Emissions

The major exhaust emissions are as follows:

  1. Unburnt hydrocarbons, (HC)
  2. Oxides of carbon, (CO and CO2)
  3. Oxides of nitrogen, (NO and NO2)
  4. Oxides of sulphur, (SO2 and SO3)
  5. Particulates
  6. Soot and smoke

The various exhaust emissions from petrol engine are as follows:

  1. Unburnt hydrocarbons (HC): The causes for the emissions of HC are incomplete combustion, crevice volumes and flow in crevices, leakage past the exhaust valve, valve overlap, deposits on walls, and oil on combustion chamber walls.

    The reasons for incomplete combustion are improper mixing of air and fuel, and flame quenching at the walls of the cylinder. Low load and idle conditions increase HC.

  2. Oxides of Carbon (CO and CO2): Carbon monoxide is generated with a rich fuel-air ratio mixture. This happens during starting and accelerating under load. Poor mixing, local rich regions, and incomplete combustion are the sources of CO emissions.
  3. Oxides of Nitrogen (NOx): NOx are created mostly from nitrogen in the air and fuel blends.In addition to temperature, the formation of NOx depends on pressure, A/F ratio, combustion duration, and location of spark plug.
  4. Oxides of Sulphur (SOx): These are generated due to the presence of sulphur in the fuel. SO2 and SO3 react with water to give rise to H2SO3 and H2SO4, which causes acid rain.

The variation of emissions from a petrol engine with A/F ratio is shown in Fig. 11.20.

Figure 11.20 Variation of emissions with A/F

11.12.2 Evaporative Emission

As mentioned earlier, there are two main sources of evaporative emissions—the fuel tank and the carburettor. The main factors governing the tank emissions are fuel volatility and ambient temperature but the tank design and location can also influence the emissions as location affects the temperature. Insulation of the fuel tank and vapour collection systems have all been explored with a view to reduce the tank emissions.

Carburettor emission may be divided into two categories as running losses and parking losses. Most internally vented carburettors have an external vent which opens at idle throttle position. The existing pressure forces prevent outflow of vapours to the atmosphere. Internally vented carburettor may enrich the mixture which, in turn, increases exhaust emission. Carburettor losses are significant only during hot conditions when the vehicle is in operation. Fuel volatility also affects the carburettor emissions.

11.12.3 Crankcase Emission

It consists of engine blow by-gases and crank case lubricant fumes. From the point of view of pollution, blow-by gases are the most important. The blow-by is the phenomenon of leakage past the piston from the cylinder to the crankcase because of pressure difference. The blow of HC emissions is about 20% of the total HC emission from the engine. This is further increased to 30% if the piston-rings are worn.

11.12.4 Lead Emission

Lead emissions come only from SI engines. The lead is present in the fuel as lead tetraethyl or tetramethyl, to control the self-ignition tendency of fuel-air mixtures that is responsible for knock.

11.13 ❐ CONTROL OF EMISSIONS IN SI ENGINE

An emission control programme aims at reducing the concentration of CO, HC, and NO. The main approaches adopted are as follows:

  1. Engine design modification: The following steps reduce the exhaust emission by engine design modification:
    1. Use leaner air-fuel ratio
    2. Retard ignition timing
    3. Avoid flame quenching time by reducing the surface to volume ratio of the combustion chamber
    4. Lower compression ratio
    5. Reduce valve overlap
    6. Modify the induction system by using high velocity carburettors or multi-chock carburettors
  2. Exhaust gas oxidation: The devices used to reduce HC/CO emissions are as follows:
    1. Use of after-burner
    2. Use of exhaust manifold reactor
    3. Use of catalytic converter
  3. Fuel modification: Air-fuel ratios leaner than stoichiometric result in almost insignificant amount of CO and reduce HC with reduced specific fuel consumption.
  4. Blow-by control: The crankcase blow-by control is the recirculation of the vapours back to the intake air cleaner.

To reduce atmospheric pollution, two different approaches are followed:

  1. Reduction of formation of pollutants in the emission by redesigning the engine system, fuel system, cooling system, and ignition system.
  2. Destroying the pollutants after these have been formed.

In petrol engines, the main pollutants which are objectionable and are to be reduced are HC, CO, and NOx. The methods used are as follows:

  1. Modifications in the engine design: Engine modifications improve emission quality. A few parameters which improve an emission are as follows:
    1. Combustion chamber configuration: Modification in the combustion chamber as reducing surface/volume ratio can reduce quenching zone and reduce HC emission. This can also be achieved by reducing dead space around piston ring.
    2. Lower compression ratio: Lower compression ratio also reduces the quenching area and thus reduces HC emission. Lower compression ratio also reduces NOx emission due to lower maximum temperature. However, lowering compression ratio reduces thermal efficiency and increases fuel consumption. By using petrol of lower octane number, it is possible to phase the lead out of petrol, that is, use of unleaded petrol.
    3. Induction system: The supply of designed A:F ratio mixture to multi-cylinder engine is always difficult under all operating conditions of load and power. This can be achieved by proper designing of induction system or using high velocity or multiple-choke carburettors.
    4. Ignition timing: Retarding spark ignition allows increased time for the fuel to burn. Retarding the spark reduces NOx formation by decreasing NOx emission. It also reduces HC emission by causing higher exhaust temperature. However, retarding the ignition results in loss of power and consumption of fuel. The controls are designed to retard the spark timing during idling and provide normal spark advance during acceleration.
    5. Reduced valve overlap: Increased overlap carries fresh mixture with the exhaust and increases emission level. This can be avoided by reducing the valve overlap.
  2. Modifying the fuel used: To reduce the pollution from the exhaust of these engines, the emission of olefins should be obviated as far as possible. For this, we can change the fuel itself. LPG and CNG be used instead of gasoline as they produce less pollution than present petrol engines.
  3. Exhaust gas treatment: Exhaust Gas Oxidation—The exhaust gases coming out of exhaust manifold are treated to reduce HC and CO emission. A few devices are discussed below.
    1. Use of after-burner: An after-burner is a burner where air is supplied to the exhaust gases and the mixture is burned with the help of an ignition system. The HC and CO which are formed in the engine combustion chamber because of inadequate O2 and inadequate time to burn are further burned by providing air in a separate box, known as an after-burner. The after-burner is located close to the exhaust manifold with an intention that the temperature of the exhaust should not fall. The oxidation of HC in the after-burner depends on the temperature of the exhaust and the mixing provided in the after-burner. Air injection does nothing to NOx emission. A simple arrangement of an after-burner is shown in Fig. 11.21.

      The performance of this system was not satisfactory as combustion was not sustained during low HC emission.

      Figure 11.21 Typical after-burner

    2. Exhaust manifold reactor: This is a further development of the after-burner where high temperature exhaust gases and secondary air are mixed properly and burnt. Here HC carried with exhaust combines with O2 and forms non-objectionable gases.

      There are different types of after-burners where heat losses are minimised and sufficient time and mixing of exhaust and secondary air are provided.

      A special after-burner designed by Du-Point, where the entry of exhaust gases is radial and air flow is peripheral, is shown in Fig. 11.22.

    3. Catalytic converters: Catalytic oxidation of the exhausted HC and CO is accomplished by placing a common device called catalytic converter in the vehicle exhaust system. The catalytic converter is filled with catalytic material. Exhaust gas hydrocarbons and CO are oxidised while passing through the bed. The catalytic material itself does not enter into the reaction but only promotes the oxidation process at a lower temperature. Usually, air compressor is used to supply additional oxygen necessary for complete oxidation of the exhaust gas stream.

      A catalyst is an agent that aids or speeds a process or a chemical reaction without becoming a part of the reaction during the process—it is a sort of chemical middleman. In a modern car’s emissions control system, the so-called three-way catalyst helps the three major evil elements of exhaust—HC, CO, and NOx—react with oxygen and each other. The catalyst helps the HC and CO become non-poisonous CO2 and water vapour, whereas the NOx is converted into CO2, nitrogen, and water vapour.

    Figure 11.22 A typical after-burner

    Figure 11.23 Fuel-system evaporation loss control device: (a) Hot soak condition, (b) Purging condition

  4. Evaporation Emission Control Device: The purpose of this device is to collect all evaporative emissions (vapours) and recirculating them at a proper time.

    The device is shown in Fig. 11.23. It consists of an absorbent chamber, pressure balancing valve, and purge control valve. The absorbent chamber contains charcoal which can hold the hydrocarbon vapour before it escapes into the atmosphere. The fuel tank and carburettor float, which are main sources of HC emission in the form of vapour, are directly connected to the absorbent chamber when the engine is turned off, that is, under hot soak-condition. This causes the petrol to boil from the carburettor float and a large amount of petrol vapour comes out. All these vapours during stopping or running the engine are absorbed in the absorber chamber.

    When the absorber bed becomes saturated, the air coming out from the air-cleaner is passed through the absorber bed and the air with vapour is passed to the inlet manifold through the purge valve. Here, the seat of the pressure balancing valve is so located that there is direct pressure communication between the internal vent and the top of the carburettor float, maintaining the designed carburettor metering forces.

    The operation of the purge control valve is controlled by the exhaust back pressure as shown in Fig. 11.23. The fuel supply is cut off under idling condition and the level of HC is reduced.

11.14 ❐ CRANK CASE EMISSION CONTROL

The basic principle of a crank case blow-by control system is the recirculation of vapours back to the inlet manifold. Figure 11.24 shows a typical closed type known as a positive crank case ventilation (PCV) system. The gases escaping past the piston and entering into the crankcase are returned to the inlet manifold and then to the engine. During compression, the HC enters into the crankcase due to pressure difference between the engine cylinder and crank case (small vacuum) and the HC from the crank case is taken back into the intake manifold during the expansion stroke and back side of the piston compresses the gases in the crank cases.

11.15 ❐ CI ENGINE EMISSIONS

The diesel engine is used more than any other type of engine for transportation, thermal power generation, and many other industrial and agricultural applications. The exhaust emissions from combustion in diesel engine are no different from those of combustion processes in petrol engine, the difference being only in the level of concentration of individual pollutants. The sample of a diesel exhaust may be free from smoke, odour, and HC or may be heavily smoke-laden, highly malodorous, and can have heavy concentration of unburned HC.

Figure 11.24 Positive crank case ventilation system

The pollutants from a diesel engine can be classified into two types as visible and invisible emissions. Visible emission is the smoke which is objected more by the public. The invisible emissions include CO, HC, NO, SO2, partially oxidised organics (as aldehydes and ketones), and odours. An unpleasant odour is also heavily objected by the public. Smoke and odour are not harmful to public health but are objectionable because of their unsightliness, unbearable smell, and possible reduction in visibility. Other invisible emissions mentioned above have similar effect on health as they are also emitted by petrol engines.

11.15.1 Effect of Engine Type on Diesel Emission

The type of the engine and the speed of the engine are two main factors which influence the exhaust emission from a diesel engine. It has been observed that there is a significant difference in emission levels from different engines except the odour level.

The following observations can be summarised:

  1. A two-stroke air-scavenged engine produces high HC and intermediate NOx. The smoke level remains low at all load conditions.
  2. A four-stroke medium speed engine has the lowest emissions of all constituents except high smoke intensity.
  3. A four-stroke, high speed engine has high HC emissions.
  4. A turbo-charged, four-stroke engine is notably low in HC but high in NOx. The smoke level is also considerably low compared with other engines.

Formation of Smoke and Affecting Factors

Engine exhaust smoke is the result of incomplete combustion. Smoke from exhaust is a visible indicator of the combustion process within the engine. It is generated at any volume in the engine where the mixture is rich. The fuel-air ratio greater than 1.5 and at pressures developed in diesel engine produces soot. Once soot is formed, it can burn if it finds sufficient O2; otherwise it comes out with exhaust. It becomes visible if it is dense. The size of the soot particles affects the appearance of smoke. The soot particles agglomerate into bigger particles which have an objectionable darkening effect on diesel exhaust.

The turbo-charged engine emits less smoke with increasing fuel air ratio as the mixing is much better in this engine as well as sufficient O2 is also available in the engine cylinder at all times.

Diesel Odour

It has been observed from the experiments that the products of partial oxidation are the main cause of odour in diesel exhaust. This partial oxidation may due to a very lean mixture during idling or due to wall-quenching effect. The effect of fuel-air ratio and odour is shown in Fig. 11.25.

The members of the aldehyde family are considered responsible for the pungent odours of diesel exhaust. The aldehydes in the exhaust are found at maximum 30 ppm but it is observed that even 1 ppm can cause irritation to the nose and eyes. There is no standard method developed yet for measuring the odour. However, several odour producing components such as nepthaldehyde, n-butylbenzene, and so on, are given standard rating and trained personnel can give odour ratings for the diesel exhaust sample by comparison.

The factors which affect the odour formation in the diesel engine are as follows:

  1. Fuel-air Ratio: It is already mentioned that lean mixtures produce odours.
  2. Mode of engine operation: The mode of operation of the engine affects the exhaust odour significantly. Maximum odour occurs when the engine is accelerated from idling.
  3. Engine type: The odour intensity does not change with the type of engine—two-stroke or four-stroke engine.

It is also claimed by a few researchers that the intensity of odour is reduced by additive compounds.

Figure 11.25 Effect of fuel air ratio on odour in diesel exhaust

Unburned Hydrocarbons

The concentration of hydrocarbons in diesel exhaust varies from a few ppm to a several thousand ppm, depending on the load on the engine and its speed. The hydrocarbons in diesel exhaust are composed of a mixture of many individual hydrocarbons in the fuel supplied to the engine as well as partly burned hydrocarbons produced during combustion process.

During the normal operation of the engine, the relatively cold wall quenches the fuel-mixture and inhibits the combustion, leaving a thick layer of unburned fuel air mixture over the entire surface of the combustion chamber. The thickness of this layer depends on the combustion pressure, temperature, mixture ratio, turbulence, and residual gases in the engine at the end of the exhaust stroke. A greater surface to volume ratio of the combustion chamber leads to the formation of greater fraction of hydro-carbon from the quenched zone.

Carbon Monoxide

CO is formed when there is insufficient O2 to completely oxidise the fuel during combustion of fuel. The amount of CO formed in a diesel engine is considerably lower than a petrol engine because of supply of continuous excess air to the engine. Theoretically, diesel engine should not emit CO at all as it always operates with excess air. However, CO is present in small quantities in diesel exhaust and this is, possibly, due to the fact that the fuel injected during later part of injection does not find sufficient O2 as a result of local depletion in certain parts of the combustion chamber.

The percentage of CO in exhaust varies from 0.1%−0.75%, which is easily acceptable level.

Oxides of Nitrogen

Among the gaseous pollutants emitted by the diesel engine, NOx are the most significant. In this respect, the diesel engine is not very much behind the gasoline engine. NOx being most hazardous, the limit is set to 350 ppm in many countries. In many diesel engines, NOx varies from a few hundreds to 1000 ppm.

The mechanism of formation of NOx is the same as discussed in the petrol engine. The conditions which create the highest local temperature (2000 K) and have sufficient O2 give the highest NOx concentration in diesel engine too.

The pre-combustion chamber diesel engines produce less NOx than a direct injection engines because of low peak temperature. High fuel-air ratio (rich mixture), an additional fuel tends to cool the charge, and localised peak temperature falls and reduces the NOx emission.

In addition, injection pattern, injection period, cetane number of the fuel, viscosity, and rate of burning also affect NOx formation significantly. The effect of load on NOx emission rates for a four-stroke normally aspirated engine and four-stroke turbo-charged engines are shown in Fig. 11.26. The NOx emission for turbocharged engine is considerably high compared with a normally aspirated engine at all load conditions.

Smog

The increase in the levels of SO2, SO3, NOx, and suspended particulate matter creates havoc with the surrounding atmosphere. Out of these, SO2 is the most potent as it causes bronchitis spasms. O3 causes inflammations of the inner lining of air passages. A cocktail of these lethal substances narrows and inflames the air passages leads to smelling in the lining of the throat and finally wheezing and difficulty in breathing.

Figure 11.26 Effect of load on NOx emission for four-stroke normally and turbo-charged engines

11.15.2 Control of Emission from Diesel Engine

The main pollutants from the diesel engine as mentioned earlier are HC, CO, NOx smoke, odour, and SO2. The methods of reducing HC, CO, and NOx from petrol engine are already discussed in detail and the same methods are also used for reducing the pollutants from a diesel engine. Therefore, the methods used to reduce smoke and odour which are additional pollutants from diesel engine are discussed here.

Smoke and Control of Smoke

Formation of smoke is basically a process of conversion of molecules of hydrocarbon fuels into particles of soot. It should be noted that soot is not carbon but simply an agglomeration of very large polybenzenoid free radicals. It is also observed that soot formation during the early part of the actual combustion process is common to all diesel engines but is consumed during the latter part of combustion.

Pyrolysis of fuel molecules is thought to be responsible for soot formation. Fuel heated with insufficient O2 will give carbonaceous deposits. It is believed that the ‘heavy ends’ of diesel fuel may pyrolyse to yield the type of smoke that is observed from a diesel engine. This is believed to be the path of formation of polycyclic aromatic hydrocarbons (benzo-pyrene) found in soot.

Many theories have been put forward for the formation of smoke but the basic reactions leading to the formation of smoke are not fully known.

There is hardly any successful method to control the formation of soot except the engine has to run at lower load and maintain the engine at best possible condition.

Some methods suggested for the control of smoke are as follows.

  1. Smoke-suppressing additives: It has been found that some barium compounds added in fuel reduce the temperature of combustion and avoid soot formation. It is further observed that if the soot is found, the barium compounds break them in very fine particles and reduce the smoke. However, berium salts added in fuel form the deposit on engine parts and reduce the filtering capacity of the filter.
  2. Fumigation: Fumigation is a method of injecting a small amount of fuel in the intake manifold. This helps pre-combustion reactions during compression stroke and reduces the chemical delay because the intermediate products such as peroxides and aldehydes react more rapidly with O2 than hydrocarbons. Reducing the chemical delay curbs thermal cracking which is responsible for soot formation.

    The cracking may not even happen which is mainly responsible for soot formation when fumigation is used because it requires about 350 kJ/mole to break C-C bond and 425 kJ/mole to break C—H bond. The energy required may not be available due to easy oxidation during pre-combustion reaction.

  3. Catalytic convertors: Catalytic convertors are not effective like in a petrol engine because of large soot formation which interferes in the oxidation of HC, CO, and NOx. Therefore, these catalysts have a very small effect on engine smoke. Extensive research is being carried on to use catalysts for effective removal of soot as well as for removal of other emissions.

Odour Control

It is claimed by many manufacturers that odour additive compounds can reduce odour intensity. However, it is observed that by using additives, there is hardly any effect on odour formation and is carried by exhaust gases.

The control of odours by using catalysts are under development and experiments have revealed that a few oxidation catalysts reduce odour intensity.

Unfortunately, the study of exhaust odour is hampered by lack of standard tests and standard units to measure the intensity of odour and the type of odour.

Several nations have been undertaking studies on suppressing odours by different methods.

Cleaning Up Diesel Emissions with Plasma and a Converter

While diesel engines are more economical, they produce NOx during combustion and put engine designers in a catch-22 situation. Presently, the permitted emission of NOx in Europe is 0.7 g /km and it will be further reduced to 0.57 g /km. This limit is already in force in USA.

The major problem faced to reduce NOx by catalytic converter which has proved successful in conventional engines cannot be applied to lean burn S.I. engines or diesel engines because these engines burn their fuel with high excess air (A:F = 30:1) and O2 in the exhaust prevents the catalytic decomposition of NOx.

Siemens and Partners have developed an efficient exhaust gas purification process for diesel engine (SINOx). This system comprises a catalytic converter, a control system, and a dosing device for urea. The urea undergoes hydrolysis into CO2 and NH3, which act as a reducing agent, transforming NOx into environmentally compatible N2 and water. However, in order to chemically reduce NOx effectively in a catalytic converter, a minimum temperature of 200°C is essential—a condition that is usually met in trucks. In passenger cars, on other hand, the low efficiency of the process at temperatures below 200°C causes problems typical of urban driving and cold start phase.

11.15.3 NOx−Emission Control

The concentration NOx in the exhaust is closely related to the peak cycle temperature. There are different methods by which peak cycle temperature can be reduced and NOx emission can be controlled.

There are mainly three methods which are commonly used as follows: catalyst (which is already discussed), water injection (rarely used), and exhaust gas recirculation (EGR) method. EGR is commonly used to reduce NOx. This method is used in petrol as well as diesel engines. In SI engines, about 10% recirculation reduces NOx emission by 50%. Unfortunately, the consequently poorer combustion directly increases hydrocarbon emission and calls for mixture enrichment to restore combustion regularity which gives a further indirect increase of both HC and CO.

Figure 11.27 shows the arrangement of an EGR system. A portion of the exhaust gases is recirculated to the cylinder intake charge. This reduces the quantity of O2 available for combustion.

The exhaust gas for recirculation is taken as shown in Fig. 11.27 through an orifice and passed through the control valve for the regulation of the quantity of recirculation.

Figure 11.27 EGR-system

The effect of A:F ratio on NOx emission takes EGR as a parameter as shown in Fig. 11.28. It can be seen that maximum emission of NOx occurs during lean mixture when gas recirculation is the least effective. On the other hand, for less emission of CO and HC, a lean mixture is preferred. About 15% recycling reduces NOx by 80% but increases HC and CO by 50%−80%. These are conflicting requirements of this emission control system, and can be solved by adopting a package system to control all emissions.

Figure 11.28 Effect of recycling of gas on NOx concentration

11.16 ❐ THREE-WAY CATALYTIC CONVERTER

A catalyst is a substance that accelerates a chemical reaction by lowering the energy needed for it to proceed. It is not consumed in the reaction. A three-way catalytic converter reduces the concentration of CO, HC, and NOx in the exhaust. A catalytic converter is usually a stainless steel container mounted along the exhaust pipe of the engine. There is a porous ceramic (Al2O3) structure inside the container through which the exhaust gas flows. The ceramic structure is a single honey-comb structure with many flow passages. The ceramic passages contain small embedded particles of catalytic material such as platinum, palladium, rhodium, and so on that promote oxidation or reduction reaction in the exhaust gas. Platinum and palladium promote the oxidation of CO and HC, whereas rhodium promotes the reaction of NOx by reduction process.

The principle of working of a three-way catalytic converter is shown in Fig. 11.29(a) and (b).

11.16.1 Function of a Catalyst in a Catalytic Converter

The catalyst controls the level of various exhaust pollutants from the engine by changing the chemical characteristics of the exhaust gases. Catalyst materials such as platinum or platinum, palladium, and rhodium are used in the converter. CO and HC oxidise the CO2 and H2O by palladium and platinum and NOx is reduced by rhodium. An oxidation catalyst is placed downstream of the reduction catalyst.

Figure 11.29 Three-way catalytic converter: (a) Catalytic converter package, (b) Three way catalytic converter

Oxidation Reactions
2CO + O2 ⇒ 2CO2
CO + H2O ⇒ CO2 + H2
CxHy + zO2 xCO2 + yH2O

where z = x + 0.25 y

Reduction Reactions
2NO + 2CO ⇒ N2 + 2CO2
2NO + 5CO + 3H2O ⇒ 2NH3 + 5CO2
2NO + CO ⇒ N2O + CO2
2NO + 2H2 ⇒ N2 + 2H2O
2NO + 5H2 ⇒ 2NH3 + 2H2O
2NO + H2 ⇒ N2O + H2O
11.17 ❐ ENVIRONMENTAL PROBLEMS CREATED BY EXHAUST EMISSION FROM IC ENGINES

The emissions exhausted into the surroundings pollute the atmosphere and cause the following problems:

  1. Global warming: The earth surrounding the atmosphere contains a 3 mm-thick layer of ozone (O3) at 50 km from its surface in stratosphere. This layer of O3 has the specific property to absorb ultraviolet (UV) rays emitted by the sun. If UV rays enter into the atmosphere and touch the earth, they will destroy all human, animal, and crop life. It has been observed that the O3 layer is slowly getting destroyed due to the use of chlorofluorocarbons (CFCs) refrigerants used for refrigeration and air-conditioning purposes. The CFCs have verifying degree of ozone depletion potential (GWP) as well. The use of fully CFCs that are considered to have high ODP have been banned.
  2. Acid rain: The diesel engine emissions contain oxides of sulphur, SOx (SO2 and SO3) in varying amounts. These emissions are dissolved in water and give rise to sulphuric acid (H2SO4). They fall on the earth and are very harmful to human and plant life.
  3. Smog: Smog means an increase in ‘morbidity’. The increase in the levels of SO2, SO3, NOx, and suspended particulate matter create huge problem in the surrounding atmosphere. SO2 causes bronchitis spasms. SO3 causes inflammations of the inner lining of air passages, swelling in the lining of the throat, and difficulty in breathing.
  4. Odours: The members of the aldehyde family are considered responsible for the pungent odours of diesel exhaust. Even 1 ppm aldehydes in exhaust can cause irritation to the nose and eyes. The intensity of the odour is reduced by additive compounds.
  5. Respiratory and other health hazards
Pollutant Health hazards
(a) Carbon monoxide, CO Interferes with transfer of oxygen through the body, causes headache, nausea, dizziness, permanent visionary damage, coma, and even death.
(b) Nitrogen oxides (NOx) Tissue changes in lungs, heart, liver, and kidney.
(c) Hydrocarbons (HC) Reduces visibility.
(d) Smoke Creates irritation and reduces visibility.
(e) Sulphur oxides (SOx) Thickens the blood and reduces life. When it combines with water, it forms H2SO4, which is highly corrosive and destroys the crops, equipment, and buildings.
(f) Lead (Pb) It is poisonous.
(g) Odours Give unpleasant smell and cause unpleasant atmosphere.
11.18 ❐ USE OF UNLEADED PETROL

It is desirable that a catalytic converter has an effective life time equal to half to the car life or at least 2 lakh km. Converters lose their efficiency with age due to thermal degradation and poisoning of the active catalyst material. Just a small amount of lead on a catalyst site reduces HC emission reduction by a factor of two of three. Use of leaded gasoline filled two times (full tank) would completely poison a converter and make it totally useless. Therefore, leaded gasoline cannot be used in engines equipped with catalytic converters.

11.18.1 Use of Additives

To improve the combustion performance of fuels, some compounds called additives or dopes are used. The requirements of an additive are as follows:

  1. It must be knock-resistant, surface ignition-resistant, or both.
  2. It should be stable in storage and have no adverse effect on fuel stability.
  3. It should be soluble in fuel under all conditions.
  4. It should be in liquid phase at normal temperature and volatile to give rapid vapourisation in manifold.
  5. It must not produce harmful deposits.
  6. Its water solubility must be minimum to minimise handling losses.

The most commonly used additives are as follows:

  1. Tetraethyl lead (TEL), ethylene dibromide (EDB), or ethylene dichloride (EDC) are added to TEL to avoid lead deposits.
  2. Tetra-methyl lead (TML).

Summary for Quick Revision

  1. Performance parameters
    1. Indicated power,
      1. pim = indicated mean effective pressure (imep)
    2. Brake power,
    3. Frictional power, FP = IP − BP
    4. Mechanical efficiency,
    5. Specific output =
    6. Volumetric Efficiency,
    7. Brake specific fuel consumption,
    8. Indicated specific consumption,
    9. Brake thermal efficiency,
    10. Indicated thermal efficiency,
    11. Relative efficiency, =
    12. Fuel-air ratio,
    13. Relative fuel-air ratio,
    14. Equivalence ratio,
  2. Willans line method is used to determine the FP of an engine. In this method, gross fuel consumption v’s BP at constant speed is plotted and the line is extrapolated back to zero fuel consumption. The point where the line cuts the BP axis gives the FP of the engine at that speed. This test is applicable to CI engines only.
  3. Morse test

    This is a method to determine the IP of a cylinder in a multi-cylinder engine

    IP of n cylinders, IPn = BPn + FP

    IP of (n − 1) cylinders, IPn−1 = BPn–1 + FP

    IP of nth cylinder, IPnth = BPn − BPn–1

    Total IP of engine, IPn = ∑ IPnth

  4. Air standard efficiency:

  5. Heat balance sheet on minute basis:
    1. Heat supplied by fuel, Qs = f × CV
    2. Heat consumed in the system:
      1. Heat equivalent of BP, Q1 = 60 × BP in kW, kJ/min
      2. Heat carried away by cooling water, Q2 = cpw × w (Two − Twi), kJ/min
      3. Heat carried away by dry exhaust gases, Q3 = g × cpg (Tge − Ta), kJ/min

        where = a + ṁf = mass of exhaust gases, kg/min

        Steam in exhaust gases, s = 9 × H2 × f kg/min

        Mass of dry flue gases, g = s

      4. Heat carried away by steam in exhaust gases,

        Q4 = s [cpw (100 − ta) + hfg + cps (tge − ts)]

      5. Heat unaccounted for, Q5 = Qs − (Q1 + Q2 + Q3 + Q4)

Multiple-choice Questions

  1. Besides mean effective pressure, the data needed for determining the indicated power of an engine would include
    1. piston diameter, length of stroke, and calorific value of fuel
    2. piston diameter, specific fuel consumption, and calorific value of fuel
    3. piston diameter, length of stroke, and speed of rotation
    4. specific fuel consumption, speed of rotation, and torque
  2. For a typical automobile CI engine, for conditions of increasing engine speed, match List I with List II and select the correct answer using codes given below the lists:
    List I (Performance parameter) List II (Tendency, qualitatively)
    A. Power output 1. Increasing and then decreasing
    B. Torque 2. Decreasing and then increasing
    C. Brake specific fuel consumption 3. Increasing throughout the range
    4. Decreasing throughout the range

    Codes:

              A B C

    1.   1  2  3
    2.   1  4  3
    3.   2  3  4
    4.   3  1  2
  3. If the approximate average mean pressures during induction, compression, power and exhaust strokes of an internal combustion engine are, respectively, 15 kN/m2 below atmosphere, 200 kN/m2 above atmosphere, 1000 kN/m2 above atmosphere and 20 kN/m2 above atmosphere, then the resultant mean effective pressure, in kN/m2, is
    1. 765
    2. 795
    3. 800
    4. 805
  4. Match List I (Performance curves, labelled A, B, C and D, for a constant speed diesel engine) with List II (performance parameter) and select the correct answer using the codes given below the List:

    Codes:

              A B C D

    1.   3  4  1  2
    2.   3  4  2  1
    3.   4  3  1  2
    4.   4  3  2  2
  5. Which one of following quantities is assumed constant for an internal combustion engine while estimating its friction power by extrapolation through Willans line?
    1. Brake thermal efficiency
    2. Indicated thermal efficiency
    3. Mechanical efficiency
    4. Volumetric efficiency
  6. A gas engine has a swept volume of 300 cc and clearance volume of 25 cc. Its volumetric efficiency is 0.88 and mechanical efficiency is 0.90. What is the volume of the mixture taken in per stroke?
    1. 248 cc
    2. 252 cc
    3. 264 cc
    4. 286 cc
  7. The curve shown in the given Fig. 11.30 is characteristic of diesel engines.

    What does the Y-axis represent?

    1. Efficiency
    2. Specific fuel consumption
    3. Air-fuel ratio
    4. Total fuel consumption

    Figure 11.30 Characteristic of diesel engines

  8. The correct sequence of the decreasing order of brake thermal efficiency of the three given basic type of IC engines is
    1. four-stroke CI engine, four-stroke SI engine, two-stroke SI engine
    2. four-stroke SI engine, four-stroke CI engine, two-stroke SI engine
    3. four-stroke CI engine, two-stroke SI engine, four-stroke SI engine
    4. two-stroke SI engine, four-stroke SI engine, four-stroke CI engine
  9. Keeping other parameters constant brake power diesel engine can be increased by
    1. decreasing the density of intake air
    2. increasing the temperature of intake air
    3. increasing the pressure of intake air
    4. decreasing the pressure of intake air
  10. The method of determination of indicated power of multi-cylinder SI engine is by the use of
    1. Morse test
    2. Prony brake test
    3. motorist test
    4. heat balance test
  11. In the context of performance evaluation of IC engine, match List I with List II and select the correct answer.
    List I
    (Parameter)
    List II
    (Equipment for measurement)
    A. Brake power (BHP) 1. Bomb calorimeter
    B. Engine speed 2. Electrical tachometer
    C. Calorific value of fuel 3. Hydraulic dynamometer
    D. Exhaust emissions 4. Flame ionisation detector

    Codes:

              A B C D

    1.   3  1  2  4
    2.   4  2  1  3
    3.   3  2  1  4
    4.   2  3  4  1
  12. The presence of nitrogen in the products of combustion ensures that
    1. complete combustion of fuel takes place
    2. incomplete combustion of fuel occurs
    3. dry products of combustion are analysed
    4. air is used for combustion
  13. A two-stroke engine has a speed of 750 rpm. A four-stroke engine having an identical cylinder size runs at 1500 rpm. The theoretical output of the two-stroke engine will
    1. be twice that of the four-stroke engine
    2. be half that of the four-stroke engine
    3. be the same as that of the four-stroke
    4. depend upon whether it is a CI or SI engine
  14. For same power output and same compression ratio, as compared to two-stroke engines, four-stroke SI engines have
    1. higher fuel consumption
    2. lower thermal efficiency
    3. higher exhaust temperature
    4. higher thermal efficiency
  15. Which one of the following plots correctly represents the variation of thermal efficiency (y-axis) with mixture strength (x-axis)?
  16. Match List I with the performance curves and select the correct answer using the codes given below the List:

    Codes:

              A B C D

    1.   1  3  2  5
    2.   1  3  2  4
    3.   1  2  3  5
    4.   2  1  4  3
  17. Consider the following statements:
    1. Volumetric efficiency of diesel engines is higher than that of SI engines.
    2. When a SI engine is throttled; its mechanical efficiency decreases.
    3. Specific fuel consumption increases as the power capacity of the engine increases.
    4. In spite of higher compression ratios, the exhaust temperature in diesel engines is much lower than that in SI engines.

    Of these statements,

    1. I, II, III, and IV are correct
    2. I, II, and III are correct
    3. III and IV are correct
    4. I, II, and IV are correct
  18. In a variable speed SI engine, the maximum torque occurs at the maximum
    1. speed
    2. brake power
    3. indicated power
    4. volumetric efficiency
  19. In a Morse test for a two-cylinder, two-stroke, spark ignition engine, the brake power was 9 kW, whereas the brake powers of individual cylinders with spark cut off were 4.25 kW and 3.75 kW, respectively. The mechanical efficiency of the engine is
    1. 90%
    2. 80%
    3. 45.5%
    4. 52.5%
  20. Match List I (performance Parameter Y) with List II (Curves labelled 1, 2, 3, 4, and 5 BHP vs. Y) regarding a CI engine run at constant speed and select the correct answer using the codes given below the lists:

    Codes:

              A B C D

    1.   5  3  4  2
    2.   1  3  4  2
    3.   5  4  2  3
    4.   1  4  2  3
  21. Match List I with List II and select the correct answer using the codes given below the lists:
    List I List II
    A. Supercharging 1. Multi-cylinder engine
    B. Morse test 2. CI engine
    C. Heterogeneous combustion 3. Calorific value
    D. Ignition quality of petrol 4. Aircraft engine
    5. Octane number
    6. Single cylinder SI engine

    Codes:

              A B C D

    1.   4  1  2  5
    2.   6  3  2  5
    3.   6  1  5  2
    4.   4  3  5  2
  22. With respect to IC engine emissions, consider the following statements:
    1. Evaporative emissions have no carbon monoxide and oxides of nitrogen
    2. Blow-by emissions are essentially carbon monoxide and suspended particulate matter
    3. Exhaust emissions contain 100% of carbon monoxide, 100% of oxide of nitrogen, and around 50%–55% of hydrocarbons emitted by the engine
    4. There are no suspended particulate in the exhaust

    Of these statements,

    1. I and IV are correct
    2. I and III are correct
    3. II and III are correct
    4. I, II, III, and IV are correct
  23. A hydrocarbon fuel was burnt with air and the Orsat analysis of the dry products of combustion yielded the following data:
    Initial volume of dry gas sample 100 cc
    1. Volume after absorption in pipette containing potassium hydroxide solution 89 cc
    2. Volume after absorption in pipette containing solution of pyrogallic acid and potassium hydroxide 84 cc
    3. Volume after absorption in pipette containing cuprous chloride solution 82 cc

    The percentage (by volume) of CO2 in the dry products was

    1. 2%
    2. 5%
    3. 11%
    4. 18%
  24. The volumetric efficiency of a well-designed SI engine is in the range of
    1. 40%–50%
    2. 50%–60%
    3. 60%–70%
    4. 70%–90%
  25. Variation of specific fuel consumption with fuel-air ratio for spark ignition engine is represented by which of the curves shown in Fig. 11.31?

    Figure 11.31

    1. Curve 1
    2. Curve 2
    3. Curve 3
    4. Curve 4
  26. Exhaust emissions versus air-fuel ratio curves for a petrol engine are shown in Fig. 11.32.

    Figure 11.32

    The curve C represents

    1. hydrocarbon
    2. carbon dioxide
    3. carbon monoxide
    4. oxides of nitrogen
  27. If the performance of diesel engines of different sizes, cylinder dimensions, and power rating are to be compared, which of the following parameters can be used for such comparison?
    1. Swept volume
    2. Air fuel ratio
    3. Specific brake fuel consumption
    4. Volumetric efficiency
  28. Consider the following statements for NOx emissions from IC engines:
    1. Formation of NOx depends upon combustion temperature
    2. Formation of NOx depends upon type of coolant used
    3. Exhaust gas recirculation is an effective means for control of NOx
    4. Activated Platinum is used for reduction of NOx

    Which of the statements given above are correct?

    1. I and II
    2. I, II, and III
    3. II and IV
    4. I and III
  29. Consider the following statements:

    Exhaust emission of carbon monoxide from spark ignition engine is

    1. Mainly fuel-air mixture strength dependent
    2. In the range of zero to 10%
    3. Measured with the help of an instrument working on the principle of non-dispersive infra-red analysis.
    4. Controlled by the use of a two way catalytic convertor

    Which of the statements given above are correct?

    1. I and IV
    2. II and III
    3. I and III
    4. I, II, III, and IV
  30. An engine using octane-air mixture has N2, O2, CO2, CO, and H2O as constituents in the exhaust gas. Which one of the following can be concluded?
    1. Supply mixture is stoichiometric
    2. Supply mixture has incomplete combustion
    3. Supply mixture is rich
    4. Supply mixture is lean
  31. An engine produces 10 kW brake power while working with a brake thermal efficiency of 30%.If the calorific value of the fuel used is 40,000kJ/kg, then what is the fuel consumption?
    1. 1.5 kg/h
    2. 3.0 kg/h
    3. 0.3 kg/h
    4. .0 kg/h
  32. A 40 kW engine has a mechanical efficiency of 80%. If the frictional power is assumed to be constant with load, what is the approximate value of the mechanical efficiency at 50% of the rated load?
    1. 45%
    2. 55%
    3. 65%
    4. 75%
  33. Consider the following statements:
    1. Supercharging increases the power output and increases the volumetric efficiency
    2. Supercharging is more suitable for SI engines than CI engines
    3. The limit of supercharging for an SI engine is set by knock while that for a CI engine is set by thermal loading

    Which of the statements given above are correct?

    1. I and III
    2. I, II, and III
    3. II and III
    4. I and II
  34. Which one of the following cannot be controlled by a three-way catalytic converter?
    1. HC emission
    2. CO emission
    3. NOx emission
    4. SPM emission
  35. The discharge of hydrocarbons from petrol automobile exhaust is minimum when the vehicle is
    1. idling
    2. cruising
    3. accelerating
    4. decelerating
  36. What is the purpose of employing supercharging for an engine?
    1. To provide forced cooling air
    2. To raise exhaust pressure
    3. To inject excess fuel for coping with higher load
    4. To supply an intake of air at a density greater than the density of the surrounding atmosphere
  37. Consider the following statements:
    1. Supercharging increases the power output of an engine
    2. Supercharging increases the brake thermal efficiency considerably
    3. Supercharging helps scavenging of cylinders

    Which of the statements given above are correct?

    1. Only I and II
    2. Only II and III
    3. Only I and III
    4. I, II, and III

Explanatory Notes

  1. 3. (a) Resultant MEP = 1000 − 200 − (15 + 20) = 765 KN/m2
  2. 6. (c) Volumetric efficiency,

    Volume of mixture per stroke = 300 × 0.88 = 264 cc

  1. 19. (a) IP2 = 9 − 4.25 = 4.75 kW

    IP1 = 9 − 3.75 = 5.25 kW

    IP = IP1 + IP2 = 4.75 + 5.25 = 10 kW

    Mechanical efficiency =

  2. 23. (c) CO2 = 100 – 89 = 11 cc, O2 = 89 − 84 = 5 cc

    CO = 84 − 82 = 2 cc

  3. 32. (c)

    At 50% rated load, ηmech =

Review Questions

  1. Explain the method to measure the brake power of a small engine.
  2. Describe the method to measure the heat lost in exhaust gases of an IC engine.
  3. Explain the Morse test to measure the indicated power of a multi-cylinder engine.
  4. Describe the method commonly used in laboratory to measure the air supplied to an IC engine.
  5. Derive the formula used for finding the mass of air supplied to an engine using an orifice meter and tank.
  6. Draw the following curves for a single-cylinder, four-stroke petrol, and diesel engines:
    1. BP v’s fuel consumption
    2. BP v’s SFC
    3. BP v’s brake thermal efficiency
  7. Discuss the various performance parameters of IC engines.
  8. What are the effects of load on the following?
    1. ηvol
    2. ηmech
    3. ηbt
    4. BSFC of SI engines
  9. Define the following:
    1. BP
    2. IP
    3. FP
    4. BSFC
    5. ηr
  10. Define the following:
    1. Equivalence ratio
    2. ISFC
    3. Specific weight
    4. pim
  11. What are the categories of SI engine emissions?
  12. Show the emissions from a four-wheeler having an SI engine on a neat sketch.
  13. List the major exhaust emissions from a SI engine.
  14. What is evaporative emission?
  15. Name four methods for control of emissions in SI engines.
  16. What is an after-burner? What is it used for?
  17. What is a catalytic converter?
  18. What is an evaporation emission control device?
  19. Name the emissions from a CI engine.
  20. Define smog and fumigation.
  21. Explain SOx and NOx.
  22. Name the catalyst materials used in a catalytic converter.
  23. What are the environmental problems created by exhaust emission from IC engines?
  24. What are additives used in IC engine fuels?
  25. Name four alternative fuels for IC engines.
  26. Define performance number.
  27. Define HUCR.
  28. List four desirable properties of IC engine fuels.
  29. Define octane number and cetane number.
  30. What do you understand by rating of fuels?
  31. What are the main pollutants emitted by SI engine?
  32. What are the various types of exhaust emissions from SI engine? Discuss briefly.
  33. How exhaust emissions in SI engines can be controlled? Explain briefly.
  34. What are the various types of exhaust emissions from CI engine? Explain briefly.
  35. Describe the methods for control of emission from diesel engine.
  36. Write a short note on Nox emission control.
  37. What is the function of a catalytic converter? Explain the working of a three-way catalytic converter with the help of a neat sketch.
  38. Write short notes on the following:
    1. Environmental problems created by exhaust emissions from IC engines.
    2. Use of Unleaded petrol.
    3. Use of additives in IC engine fuels.
  39. Explain the various alternative fuels for IC engines.

Exercises

11.1 A single cylinder, four-stroke cycle oil engine works on diesel cycle. The following data is available from a test:

Area of indicator diagram = 3 cm3, length of indicator diagram = 4 cm, spring constant = 10 bar/cm2-cm, engine speed = 400 rpm, load on the brake = 380 N, spring reading = 50 N, diameter of brake drum = 1.2 m, fuel consumption = 2.8 kg/h, LCV = 42 MJ/kg, cylinder diameter = 16 cm, piston stroke = 20 cm.

Calculate (a) the friction power (b) the mechanical efficiency (c) the brake thermal efficiency, and (d) the brake mean effective pressure.

[Ans. (a) 1.71 kW, (b) 83%, (c) 25%, (d) 6.15 bar]

11.2 A four-cylinder, four-stroke petrol engine 6 cm bore and 9 cm stroke was tested at constant speed. The fuel supply was fixed to 0.13 kg/min and spark plugs of 4-cylinders were successively short-circuited without change of speed.

The power measurements were as follows:

With all cylinder firing = 16.25 kW, with first cylinder cut-off = 11.55 kW,

With 2nd cylinder cut-off = 11.65 kW, with 3rd cylinder cut-off = 11.70 kW,

With 4th cylinder cut-off = 11.50 kW.

Determine (a) the indicated power of the engine, (b) the mechanical efficiency, (c) the indicated thermal efficiency if LCV of fuel used is 42 kJ/kg, and (d) the relative efficiency on IP basis assuming clearance volume 65 cm3.

[Ans. 18.5 kW, 88%, 20.4%]

11.3 The following readings were taken during the test on a single cylinder, four-stroke IC engine:

Speed of engine = 240 rpm, orifice diameter of tank = 20 mm, pressure causing flow through the orifice = 10 cm of H2O, ambient temperature = 30°C, barometer reading = 760 mm of Hg, gas consumption = 3.8 m3/h at 30°C and 10 cm of H2O above atmosphere.

Find (a) the air consumption per hour, and (b) the A:F ratio by volume and weight. Take Cd = 0.7 for air orifice, R = 287 J/kg.K for air and 280 J/kg.K for gas.

[Ans. (a) 0.544 m3/min, (b) 85:1, 8:1]

11.4 The following data was obtained from a test on a single-cylinder, four-stroke oil engine.

Cylinder bore = 15 cm, stroke = 25 cm, area of indicator diagram = 540 mm2, length of indicator diagram = 50 mm, indicator spring rating = 1.2 mm for a pressure of 8.8 N/cm2, engine speed = 400 rpm, brake torque = 225 N-m, fuel consumption = 3 kg/h, calorific value of fuel = 44200 kJ/kg, cooling water flow rate = 4 kg/min, cooling water temperature rise = 42°C, specific heat of water = 4.1868 kJ/kg.°C.

Calculate (a) the mechanical efficiency, (b) the brake thermal efficiency, (c) the specific fuel consumption, and (d) draw heat balance sheet.

[Ans. (a) 87.07%, (b) 25.6%, (c) 0.31 kg/kWh, (d) Qs = 36.83 kW, Qw = 11.72 kW, Qge = 15.68 kW, Qb = 9.43 kW]

11.5 The following particulars were obtained in a trial for 1 h on a 4-stroke gas engine:

Speed = 16000 rpm, missed cycles = 600 rpm, net brake load = 1600 N, brake circumference = 4 m, IMEP = 8 bar, gas consumption = 2200 litres, CV of gas = 20 kJ/litre, d = 25 cm, L = 40 cm, r = 6.5.

Find (a) the indicated power and brake power (b) the brake specific fuel consumption, and (c) the brake thermal efficiency and relative efficiency.

[Ans. (a) 32.3 kW, 28.4 kW, (b) 773.4 litres/kWh, (c) 23.3%, 44%]

11.6 A 8-cylinder, 4-stroke petrol engine of 80 mm bore and 90 mm stroke has a compression ratio of 7. It is tested at 4000 rpm on a dynamometer which has 54 cm arm. During a 10 minute test, the dynamometer reads 400 N and the engine consumes 4.75 kg of fuel. Air is supplied at 1 bar and 27°C at the rate of 6.5 kg/min. CV of fuel = 44 MJ/kg.

Determine (a) the brake power (b) the brake mean effective pressure (c) the volumetric efficiency (d) the A:F ratio, and (e) the relative efficiency.

[Ans. (a) 90.8 kW, (b) 6 bar, (c) 77.32%, (d) 13.68:1, (e) 48%]

11.7 A six-cylinder gasoline engine operates on four stroke cycle. The bore of each cylinder is 80 mm and the stroke is 100 mm. At speed of 4000 rpm the fuel consumption is 20 kg/h and the torque developed is 150 Nm. Calculate (i) the brake power (ii) the brake mean effective pressure (iii) brake thermal efficiency if calorific value of the fuel is 43000 kJ/kg.

[Ans. (a) 62.83 kW, (b) 6.25 bar, (c) 26.3%]

11.8 An eight-cylinder, four-stroke automobile engine develops 76.25 kW brake power when tested at 4000 rpm. The compression ratio of the engine is 7. The testing is carried out for 10 minutes and fuel consumption was 4.5 kg. CV of fuel used = 45000 kJ/kg. The air passing through the carburettor has 20°C temperature and 1.03 bar pressure, Air measured was 5.45 kg/min. The diameter and stroke are both equal to 8.5 cm. Determine (a) the brake mean effective pressure, (b) the brake specific fuel consumption, (c) the brake thermal efficiency, and (d) the A:F ratio.

[Ans. (a) 5.93 bar, (b) 0.354 kg/kWh, (c) 57.6%, (d) 12.25:1]

11.9 A six-cylinder, four-stroke petrol engine develops 40 kW when running at 3000 rpm. The volumetric efficiency at NTP is 85% and the indicated thermal efficiency is 25%. The A:F ratio of the mixture supplied is 15:1. The calorific value of fuel used is 41000 kJ/kg. Assume the diameter to be equal to stroke of the engine, calculate the bore and stroke.

[Ans. 76 mm, 76 mm]

11.10 A single cylinder, four-stroke cycle oil engine works on diesel cycle. The following readings were taken when the engine was running at full load:

Area of indicator diagram = 3 cm, length of diagram = 4 cm,

spring constant = 10 bar/cm2. cm, speed of engine = 400 rpm,

Load on the brake = 380 N, Spring reading = 50 N, diameter of brake drum = 1.2 m, fuel consumption = 2.8 kg/h, calorific value of fuel = 42000 kJ/kg, cylinder diameter = 16 cm, stroke = 20 cm. Calculate (a) the friction power of the engine, (b) the mechanical efficiency, (c) the brake thermal efficiency, and (d) the brake mean effective pressure.

[Ans. (a) 1.71 kW, (b) 83%, (c) 25%, (d) 6.15 bar]

11.11 A four stroke petrol engine 80 mm bore 100 mm stroke is tested at full throttle at constant speed. The fuel supply is fixed at 0.08 kg/min and the plugs of the four cylinders are successively short circuited without change of speed, brake torque being correspondingly adjusted. The brake power measurements are the following:

With all cylinder firing = 12.5 kW

With cylinder No. 1 cut-off = 9 kW

With cylinder No. 2 cut-off = 9.15 kW

With cylinder No. 3 cut-off = 9.2 kW

With cylinder No. 4 cut-off = 9.1 kW

Determine the indicated power of the engine under these conditions. Also determine the indicated thermal efficiency. Calorific value of the fuel is 44100 kJ/kg. Compare this efficiency with the air standard value, clearance volume of one cylinder is 70 × 103 mm3.

[Ans. 13.55 kW, 23.04%, 40.52%]

11.12 A four-cylinder, four-stroke petrol engine, 60 mm bore and 90 mm stroke was tested at constant speed. The fuel supply was fixed at 0.13 kg/min and plugs of 4-cylinders were successively short-circuited without change of speed. The power measurements were as follows:

With all cylinders working = 16.25 kW

With 1st-cylinder cut-off = 11.55 kW (BP)

With 2nd-cylinder cut-off = 11.65 kW (BP)

With 3rd-cylinder cut-off = 11.70 kW (BP)

With 4th-cylinder cut-off = 11.50 kW (BP)

Find the IP of engine, the mechanical efficiency, the indicated thermal efficiency if CV of fuel used is 42,000 kJ/kg, and the relative efficiency on IP basis assuming clearance volume 65 m3.

[Ans. 1850 kW, 88%, 20.4%, 42%]

11.13 The air flow to a four cylinder, four-stroke petrol engine is measured by means of a 7.5 cm sharp-edged orifice, Cd = 0.6. During a test on the engine, following data was recorded.

Bore = 12 cm, stroke = 14 cm, engine speed = 2200 rpm, brake power = 35 kW, fuel consumption = 10 kg/h, LCV = 42000 kJ/kg, pressure drop across the orifice = 4 cm of water. Atmospheric temperature = 20°C, and atmospheric pressure = 1.01325 bar.

Calculate (a) the thermal efficiency on BP basis, (b) the brake mean effective pressure, and (c) the volumetric efficiency based on free air conditions.

[Ans. (a) 30%, (b) 1.507 bar, (c) 70.2%]

11.14 The following readings were taken during test on a single cylinder, four-stroke gas engine: speed of engine = 240 rpm, diameter of orifice of air tank = 20 mm, pressure causing the air flow through orifice = 10 cm of water, ambient temperature = 30°C, barometer reading = 76 cm of Hg, gas consumption = 3.8 m3/h measured at 30°C and 10 cm of water above atmosphere.

Find (a) air consumption per hour and (b) the A:F ratio by weight and by volume. Take Cda = 0.7, R = 287 J/kg. K for air, and R = 280 J/kg.K for gas

[Ans. (a) 0.544 m3/min, (b) 8.5:1, 8.1]

11.15 The following particulars were obtained in a trial on a 4-stroke gas engine:

Duration of trial = 1 h, revolutions = 16000, missed cycles = 600, net brake load = 1600 N, brake circumference = 4 m, MEP = 8 bar, gas consumption = 2200 litres, CV of gas = 20 kJ/litre, D = 25 cm, L = 40 cm compression ratio = 6.5. Calculate the indicated power, the brake power, the brake specific fuel consumption, the brake thermal efficiency, and the relative efficiency.

[Ans. 32.3 kW, 28.4 kW, 773.4 litres/kWh, 23.3%, 44%]

11.16 The following data is given for a four-stroke, four-cylinder, diesel engine:

Cylinder diameter = 35 cm, piston stroke = 40 cm, engine speed = 315 rpm, IMEP = 7 bar, BP of engine = 260 kW, fuel consumption = 80 kg/h, CV of fuel used = 43000 kJ/kg, hydrogen content in fuel = 13% and remaining is carbon, air consumption = 30 kg/min, cooling water circulated = 90 kg/min, rise in temperature of cooling water = 38°C, piston cooling oil used = 45 kg/min, rise in temperature of cooling oil = 23°C, cp for cooling oil = 2.2 kJ/kg.K, exhaust gas temperature = 322°C, cp for exhaust gases = 1.1 kJ/kg.K, ambient temperature = 22°C, cp of superheated steam = 2 kJ/kg.K, latent heat of steam = 2520 kJ/kg.

Calculate (a) the mechanical and indicated thermal efficiency, (b) find bsfc when load on the engine is 50% of full load assuming same indicated thermal efficiency, and (c) draw up heat balance sheet on minute and percentage basis.

[Ans. (a) 92%, 29.6%, (b) 0.308 kg/kWh, (c) Qs = 57333 kJ/min, Qb = 15600 kJ/min, Qw = 14364 kJ/min, Qoil = 2277 kJ/min, Qge = 9858 kJ/min, Qsteam = 4540 kJ/min, Qunacc = 10694 kJ/min]

ANSWERS TO MULTIPLE-CHOICE QUESTIONS
  1. c
  2. d
  3. a
  4. a
  5. b
  6. c
  7. d
  8. a
  9. c
  10. a
  11. c
  12. d
  13. c
  14. d
  15. c
  16. a
  17. a
  18. c
  19. a
  20. a
  21. a
  22. c
  23. c
  24. d
  25. c
  26. d
  27. c
  28. c
  29. c
  30. b
  31. b
  32. c
  33. b
  34. a
  35. c
  36. d
  37. c