Internal Combustion Engine Performance Characteristics Final Compressed for email

Internal Combustion Engine Performance Characteristics, Combustion Chamber Design, and Gas

Exchange Process

Thermofluids and TurbomachineryAssignment 2015-16

James Goddings

Student Number: 3131147

05/03/2016

Module Coordinator: Dr. Mark EllisENG_6_452

TABLE OF CONTENTS1. Introduction..............................................................................................................................................................1

2. Calculations..............................................................................................................................................................2

Power Correction Factor.....................................................................................................................................5

Energy Balance Diagram Calculations....................................................................................................6

3. Data Tables...................................................................................................................................................................8

Engine Data.................................................................................................................................................................. 8

Other Data..................................................................................................................................................................... 8

Perfomance Testing Data..................................................................................................................................9

4. Graphs....................................................................................................................................................................... 10

5. Energy Balance Diagram............................................................................................................................15

Discussion....................................................................................................................................................................16

6. Ignition and Combustion in Spark Ignition Engines..............................................................17

7. Intake Valve Flow Investigation............................................................................................................19

Mach Number................................................................................................................................................................21

Flow Coefficient.......................................................................................................................................................21

Calculation of Suitable Intake Valve Port Diameter...................................................................23

8. Harmful Pollutants.................................................................................................................................................25

References.......................................................................................................................................................................27

James Goddings - 3131147

1.INTRODUCTIONInternal Combustion Engines are evaluated based on a number of performance characteristics; these give a benchmark for comparing engines and a means of assessing an engines capabilities and suitable applications.

During this investigation a test bed comprising of a TQ Education and Training TD115 Hydraulic Dynamometer and TD114 Instrumentation Unit was used to test a Tecumseh TD110 172cc 4 Stroke Engine (TQ Education and Training Ltd., 2004).

The TD115 Dynamometer is a Hydraulic Dynamometer which measures the torque produced by an engine. A paddle wheel within a sealed vaned casing acts as a rotor and stator, separated by a volume of water. The engines crankshaft is connected to the paddle wheel via an extension shaft with a rubber coupling bush to isolate the engines vibrations from the dynamometer. The engine rotates the paddle wheel, churning the water within the vaned casing, which resists the waters motion. The casing rotates as the torque is transferred from the rotor to the stator, however it is constrained by a strap with a pair of springs attached and connected via a chain to a rotary potentiometer. Having been calibrated, the dynamometer reports a Torque value proportional to its angular displacement. An optical encoder is attached to the extremity of the dynamometer shaft to measure rotational engine speed.

The TD114 Instrumentation Unit consists of a number of instruments to enable other engine characteristics to be measured, including an air volumetric flow rate of the engine intake, a volumetric fuel flow meter measuring the flow rate of fuel passing through the carburettor into the combustion chamber and a thermocouple measuring exhaust temperature.

The Tecumseh TD110 172cc 4 Stroke engine is a small capacity, single cylinder, air-cooled, carburettor fed, L-Head side valve engine (see Figures 7.1-7.3). An L-Head side valve configuration was a design widespread in the automotive industry between the 1920s and 1950s. It places the valve ports in the cylinder block of the engine next to the cylinder, actuated by tappets driven by a camshaft mounted adjacent to the crankshaft, geared to a 2:1 ratio. This ensures the valves are timed to open and close twice for every rotation of the camshaft (Tecumseh, 1995).

Placing all the moving parts in the cylinder block makes packaging the engine less complex. Driving the camshaft directly off the crankshaft without the use of belts or chains means oil can be pulled up from the sump to lubricate the moving parts without the need for a pump, or integrated seals within the cylinder head. Overall the design is inexpensive to manufacture, easy to maintain, and relatively lightweight. This makes it ideal for use in applications such as lawnmowers and snow blowers, which this engine is designed for.

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Due to difficulties encountered on the day of testing a series of data performed by a previous group of students at LSBU was used for the investigation and calculations.

2.CALCULATIONSThe following calculations show a fully worked example of the calculations undertaken to produce Table 3.2, specifically the column for Test 3 at 3300rpm, highlighted grey:

Power=Engine Speed×Torque

P [kW ]= 2π [rad ]×N [rpm]×τ [Nm]60[s /min]

P [kW ]=2π [rad ]×3300[rpm]×10.9[Nm ]60 [s /min ]×1000[W /kW ]

=3.77 [kW ]

Calculation 2.1

SweptVolume= π×Bore Diameter2×Stroke4

V swept [mm3]=π ×(b[mm])2×s [mm]

4

V swept [mm3 ]= π × (66.69 [mm ] )2×49.23 [mm ]4

=1.72×105[mm3]

Calculation 2.2

BrakeMeanEffective Pressure= PV swept×N

= τV swept

BMEP [kPa ]=2[Cycles ]×2π [rad ]×τ [Nm]×109 [mm3/m3 ]V swept [mm3 ]×103 [Pa /kPa ]

BMEP [kPa ]= 2 [Cycles ]×2π [rad ]×10.9 [Nm ]×109 [mm3/m3 ]1.72×105 [mm3 ]×103 [ Pa /kPa ]

=796.5 [kPa ]

Calculation 2.3

FuelVolumetric Flow Rate=Volumeof FuelTime

V fuel [m3/hr ]=V fuel [ml ]×3600 [s /hr ]t [s ]×106 [ml /m3 ]

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V fuel [m3/hr ]= 8 [ml ]×3600 [s /hr ]11.35 [s ]×106 [ml /m3 ]

=2.54×10−3 [m3/hr ]

Calculation 2.4

FuelMass Flow Rate=V fuel× Density of Fuel=V fuel× Fuel SpecificGravity

mfuel [kg/hr ]=V fuel× ρfuel=V fuel [m3/hr ]×SG [ g/ml ]×106 [ml /m3 ]

103 [g /kg ]

mfuel [kg/hr ]=2.54×10−3 [m3/hr ]×0.74 [g /ml ]×106 [ml/m3 ]

103 [ g/kg ]=1.88 [kg/hr ]

Calculation 2.5

Specific FuelConsumption=mfuel

P

SFC [g /kWh]=mfuel [kg /hr ]×103 ¿¿

SFC [g /kWh ]=1.88 [kg /hr ]×103¿¿

Calculation 2.6

Air MassFlow Rate→UsingChart 2.8

mairNTP [kg/hr ]=(1.0223× p¿¿ vf [mmH 2O ])+1.0466 ¿

mairNTP [kg/hr ]=(1.0223×14 [mmH 2O ])+1.0466=15.37 [kg /hr ]

Calculation 2.7

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Figure 2.8 TD114 Instrumentation unit viscous flow meter calibration chart, converting pressure drop across laminar viscous flow element to air mass flow.

(TQ Education and Training Ltd., 2004)

(This Chart has been transposed from the Test Bed and Instrumentation Manual and a best fit curve calculated for the data in Excel.)

Air MassFlow Rate at NTP→Air MassFlow Rate at Experimental ATP

This formula is derived from the combination of Poiseuille’s Equation, Sutherland’s Formula for the viscosity of air and a calibration value for air pressure and temperature:

mair [kg /hr ]=3564× pair¿

mair [kg /hr ]=3564×1.020¿

Calculation 2.9 (TQ Education and Training Ltd., 2004)

Density of Air= Air PressureAir SpecificGasConstant× Air Temperature

ρair [kg /m3 ]= pair [kPa]×103[Pa/kPa]Rair [J /kgK ]×T air [ K ]

ρair [kg /m3 ]=102[kPa]×103[Pa/kPa]287 [J /kgK ] ×293.15 [K ]

=1.21235 [kg /m3 ]

Calculation 2.10

Air Volumetric Flow Rate=mair× ρair

V air [m3 /hr ]=mair [kg /hr ]× ρair [kg/m3 ]

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V air [m3 /hr ]=15.46 [kg/hr ]×1.21235 [kg /m3 ]=18.75 [m3/hr ]

Calculation 2.11

Air Fuel Ratio=mair

mfuel

AFR=mair [kg/hr ]mfuel[kg /hr ]

AFR=15.46 [kg/hr ]1.88 [kg/hr ]

=8.24 :1

Calculation 2.12

Lambda= AFRStoichiometric Air Fuel Ratio

λ= AFRSAFR

λ=8.2414.7

=0.50

Calculation 2.13

Volumetric Efficiency=V air

2×V swept× N

ηVolumetric=V air [m3/hr ]×109 [mm3/m3 ]

2[Cycles ]×V swept [mm3 ]×N [rpm]×60¿¿

ηVolumetric=18.75 [m3/hr ]×109 [mm3/m3 ]

2[Cycles ]×1.72×105 [mm3 ]×3300[rpm]×60 ¿¿

Calculation 2.14

BrakeThermal Efficiency= Pmfuel×FuelCa lorificValue

ηBTE=P [kW ]×3600[s /hr ]

mfuel [kg/hr ]×FCV [kJ ¿¿kg ]¿

ηBTE=3.77 [kW ]×3600[s /hr ]

1.88 [kg /hr ]×42000[kJ¿¿kg]=0.172=17.2%¿

Calculation 2.15

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Mean PistonSpeed=2×N ×st

Sppiston [m /s ]=2×N [rpm]×s [mm ]60[s ]×103[mm/m ]

Sppiston [m /s ]=2×3300[rpm]×49.23 [mm ]60[ s]×103 [mm /m]

=5.42 [m /s ]

Calculation 2.16

POWER CORRECTION FACTOR‘”The performance of SI and CI engines is affected by the density of the inlet combustion air as well as by the characteristics of the test fuel. Whenever possible, tests should be run at the standard conditions with reference fuels. When this is not possible, in order to provide a common basis of comparison, correction factors should be applied to the observed net power and torque to account for differences between reference air and fuel conditions and those at which the test data were acquired” (SAE J1349 JUN90).

This standard document has been produced by the Society of Automotive Engineers to provide a standard method of obtaining repeatable measurements that accurately reflect real world engine performance, in order that engines may be compared by their relative performance in a fair and unbiased manner. SAE J1349 JUN90 lists the correction factor formula for power from observed conditions to standard conditions to be:

Pcorrected [kW ]=Pobserved [kW ] [1.18 ( 99pair[kPa] )({T air [° C ]+273.15

298.15 }[K ]−0.18)]This formula was revised in the August 2004 revision assuming an 85% mechanical efficiency to:

Pcorrected [kW ]=Pobserved [kW ] [1.176 ( 99pair [kPa ])({T air [° C ]+273.15

298.15 }[K ]−0.176)]Pcorrected [kW ]=3.77 [kW ] [1.176 ( 99

102 [kPa])({20 [° C ]+273.15298.15 }[K ]−0.176)]=4.26 [kW ]

Calculation 2.17

τ corrected=P corrected

N

τ corrected [Nm]=Pcorrected [kW ] ×60¿¿

τ corrected [Nm]=Pcorrected [kW ] ×60¿¿

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Calculation 2.18

ENERGY BALANCE DIAGRAM CALCULATIONS

Fuel Energy∈¿ mfuel× FCV

Qfuel [kJ ]=mfuel[kJ¿¿kg]×FCV [kg /hr ]

3600[s /hr ]¿

Qfuel [kJ ]=1.88 [kg /hr ]×42000 [kJ ¿¿kg]3600 [s /hr ]

=21.07 [kJ ]¿

Calculation 2.19

IncompleteCombustion Losses=(1−λ ) mfuel×FCV

Qlost [kJ ]=(1−λ )×mfuel [kg/hr ]×FCV [kg /hr ]

3600[s /hr ]

Qlost [kJ ]=(1−0.56 )×1.88 [kg /hr ]×42000 [kJ ¿¿kg ]3600 [s /hr ]

=9.27[kJ ]¿

Calculation 2.20

Heat Loss¿Exhaust Gasses=Specific Heat Capacity× mair×TemperatureDifference

Qexhaust [kJ ]=cp exhaust [kJ /kgK ]×mair [kg /hr ]¿¿

Qexhaust [kJ ]=1 [kJ /kgK ]×15.46 [kg/hr ]¿¿

Calculation 2.21

Heat Loss¿Unburnt Fuel=(1−λ )mfuel× {[S pecific Heat Capacity Fuel× (T exhaust−T air ) ]+ [Enthalpy of Vaporisation ] }

Qunburnt [kJ ]=(1− λ ) mfuel [kg /hr ]3600[ s/hr ]

¿

Qunburnt [kJ ]= (1−0.56 )1.88 [kg /hr ]3600[s /hr ]

¿

Calculation 2.22

An estimated value for characteristic total heat losses to the engine coolant for an air cooled spark ignition combustion engine from of 30% (Pulkrabek W.,2004) has been assumed:

Heat Loss¿ EngineCoolant=0.30× (Q fuel−Qlost )

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Qcoolant [kJ ]=0.30× (Qfuel [kJ ]−Qlost [kJ ])

Qcoolant [kJ ]=0.30× (21.07 [kJ ]−9.27[kJ ])=3.50 [kJ ]

Calculation 2.23

An estimated value for characteristic mechanical losses for a spark ignition combustion engine from (SAE J1349 AUG04) of 15% has been assumed:

Mechanical Losses=0.15× (Qfuel−Qlost )

f mechanical[kJ ]=0.15× (Qfuel [kJ ]−Qlost[kJ ])

f mechanical [kJ ]=0.15× (21.07 [kJ ]−9.27 [kJ ] )=1.77 [kJ ]

Calculation 2.24

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3. DATA TABLES

ENGINE DATA OTHER DATAManufacturer(

1) Tecumseh Fuel Specific Gravity(3) [g/cm3] 0.74

Capacity(1) [cm3] 171.9 Fuel Calorific Value(4) [kJ/kg] 42000

Stroke(1) [mm] 49.23Specific Heat Capacity of

Exhaust Gases(5)

[kJ/kgK] 1

Bore(1) [mm] 66.69Specific Heat Capacity of Octane(5)

[kJ/kgK] 2.35(300−700[K ]¿

Compression Ratio(1) 6

Enthalpy of Vaporisation of Octane(6)

[kJ/kg] 41.5

Swept Volume(2) [mm3] 1.719E+05

Fuel Stoichiometric Fuel Ratio(7)

14.7

Table 3.1 Manufacturers Engine Data

(1 – Tecumseh, 1995) , (2-Calculation 2.2)

Table 3.2 Other Data used in Calculations

(3–Paul S. 1995), (4-Ganesan V. 2012), (5-Rodgers, G. and Mayhew, Y. 2013), (6-Majer, V. and Svoboda, V. 1985) (7- Crolla, D. 2009)

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PERFOMANCE TESTING DATATest No. 1 2 3 4 5 6 7 8Engine Speed [rpm] 3750 3600 3300 3000 2700 2000 1500 1400Torque [Nm] 6.4 7.5 10.9 10.9 11.1 10.7 9 9Fuel Flow Measurement Time for 8ml [s] 13 13.5 11.35 11.8 13.6 19.9 20.6 21.87Air Flow Measurement [mmH2O] 10.5 12 14 13 11 7 5 5Exhaust Gas Temperature [°C] 450 460 450 440 430 400 350 300Barometric Pressure [kPa] 102 102 102 102 102 102 102 102Ambient Temperature [°C] 20 20 20 20 20 20 20 20Power [kW] 2.51 2.83 3.77 3.42 3.14 2.24 1.41 1.32Brake Mean Effective Pressure [kPa] 467.7 548.1 796.5 796.5 811.1 781.9 657.7 657.7Fuel Volumetric Flow Rate [m3/hr] 2.22E-

032.13E-

032.54E-

032.44E-

032.12E-

031.45E-

031.40E-

031.32E-

03Fuel Mass Flow Rate [kg/hr] 1.64 1.58 1.88 1.81 1.57 1.07 1.03 0.97Specific Fuel Consumption [g/kWh] 652.3 558.3 498.5 527.4 499.3 477.9 731.8 738.5Air Volumetric Flow Rate [m3/hr] 14.38 16.25 18.75 17.50 15.00 10.01 7.52 7.52Air Mass Flow Rate [kg/hr] 11.86 13.41 15.46 14.43 12.38 8.26 6.20 6.20Lambda 0.49 0.58 0.56 0.54 0.54 0.52 0.41 0.43Air Fuel Ratio 7.24 8.49 8.24 7.99 7.90 7.71 5.99 6.36Volumetric Efficiency [%] 18.6 21.9 27.5 28.3 26.9 24.3 24.3 26.0Brake Thermal Effeciency [%] 13.1 15.4 17.2 16.3 17.2 17.9 11.7 11.6

Mean Piston Speed [m/s] 6.15 5.91 5.42 4.92 4.43 3.28 2.46 2.30Corrected Power [kW] 2.84 3.20 4.26 3.88 3.55 2.54 1.60 1.49Corrected Torque [Nm] 7.24 8.49 12.34 12.34 12.56 12.11 10.19 10.19

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Table 3.3 Performance Characteristics for Tecumseh TD110 Engine from a Series of Tests Performed at LSBU

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4.GRAPHS

Figure 4.1 Characteristic Torque response for the Tecumseh TD110 single cylinder 4-stroke spark ignition engine obtained by experiment at LSBU.

Figure 4.1 shows a characteristic curve for a normally aspirated spark ignition small engine. Torque rises with engine speed to the middle range of engine speeds, where it forms a broad peak; in this series of tests between 2000-3300[rpm]. As engine speed increases beyond this into the upper range of engine speeds the engine becomes less volumetrically efficient and the torque drops off.

Figure 4.2 Characteristic Power response for the Tecumseh TD110 single cylinder 4-stroke spark ignition engine obtained by experiment at LSBU.

Figure 4.2 shows a typical response for a normally aspirated spark ignition small engine. Power rises with engine speed, until the upper range of engine speeds, where it comes to a sharp peak; in this series of tests at 3300[rpm]. As engine speed increases beyond this the engine becomes less volumetrically efficient and generates less torque, therefore the power drops off.

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Figure 4.3 Characteristic BMEP response for the Tecumseh TD110 single cylinder 4-stroke spark ignition engine obtained by experiment at LSBU.

Due to the proportionality between BMEP and Torque produced at a specific engine speed, Figure 4.3 closely resembles Figure 4.1.

Figure 4.4 Characteristic Mean Piston Speed for the Tecumseh TD110 single cylinder 4-stroke spark ignition engine obtained by experiment at

LSBU.

Due to the stroke of the engine being a fixed dimension, the mean piston speed is directly proportional to the engine speed; hence the relationship shown by Figure 4.4 of a straight line, which if regressed would pass through the origin.

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Figure 4.5 Characteristic Specific Fuel Consumption for the Tecumseh TD110 single cylinder 4-stroke spark ignition engine obtained by

experiment at LSBU.

The specific fuel consumption curve illustrated by Figure 4.5 resembles an inverse of the torque curve, with a flattened trough between 2000-3300[rpm]. As engine speed increases beyond this the engine becomes less volumetrically efficient and therefore less thermally efficient. Below 2000rpm the air fuel ratio is too rich and a lot of fuel is wasted by being unburnt.

Figure 4.6 Characteristic Air Fuel Ratio and resulting Lambda for the Tecumseh TD110 single cylinder 4-stroke spark ignition engine obtained

by experiment at LSBU.

There is a direct correlation between the two sets of data plotted in Figure 4.6, with lambda representing the air fuel ratios proportion to the stoichiometric air fuel ratio for the specific fuel. The resulting graph shows the engine to be running very rich with a low air fuel ratio throughout the entire range of engine speeds.

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Figure 4.7 Characteristic Brake Thermal Efficiency for the Tecumseh TD110 single cylinder 4-stroke spark ignition engine obtained by

experiment at LSBU.

Figure 4.7 illustrates the brake thermal efficiency for this engine over this series of tests to be very low. This is an inexpensive, low specification, small capacity single cylinder engine. It is tuned to operate most efficiently between 2000-3300[rpm] indicated by the broad flat peak; due to the very rich AFR much of the energy input is wasted as unburnt fuel. Typical passenger car spark ignition engines achieve efficiencies greater than 20% and even approaching 30% with forced induction.

Figure 4.8 Characteristic Volumetric Efficiency for the Tecumseh TD110 single cylinder 4-stroke spark ignition engine obtained by experiment at

LSBU.

Volumetric efficiency describes how the engine performs when exchanging gases within the cylinder. This is a critical performance characteristic for an engine, as it can only burn fuel in the presence of oxygen in the air. In Figure 4.8 it is

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evident the volumetric efficiency is fairly constant rising steadily until 3000-3300[rpm]. Beyond this, as piston speed becomes greater and gas velocities increase the engine becomes less efficient at exchanging gases within the cylinder.

Figure 4.9 Characteristic Exhaust Gas Temperature for the Tecumseh TD110 single cylinder 4-stroke spark ignition engine obtained by

experiment at LSBU.

As described by Figure 4.9, this engine shows a typical relationship between engine speed and exhaust temperature. The ability of this air cooled engine to dissipate the heat energy produced as more fuel is burnt per minute is fixed; therefore the exhaust gases must carry more heat away from the combustion chamber. The drop in temperature at 3750[rpm] may be due to the increased proportion of unburnt fuel in the exhaust; which due to its high specific heat capacity would dissipate more heat energy.

Figure 4.10 Comparison of Indicated Power and Torque with Power and Torque corrected for Barometric Pressure.

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As barometric pressure increases, the density of air, and therefore amount of oxygen available for combustion increases. This means the engine can burn more fuel per cycle and therefore produce more torque per cycle. This increases the power output of the engine. Figure 4.10 shows this increase can be significant for a relatively small increase in barometric pressure.

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5.ENERGY BALANCE DIAGRAM

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Qin = 21.07[kJ]

9.27[kJ] = Uncombusted Fuel

0.42[kJ] = Heat Loss to Uncombusted Fuel

1.77[kJ] = Mechanical Loses

1.85[kJ] = Heat Loss to Exhaust Gases

3.5[kJ] = Heat Loss to Coolant

4.26[kJ] = Power Output


DISCUSSIONThe Tecumseh TD110 172cc 4 Stroke engine is a small capacity, single cylinder, air-cooled, L-Head side valve engine. It is designed for use in machinery such as snow blowers and lawnmowers (Tecumseh, 1995). This places design emphasis on the engine being inexpensive and capable of running on poor quality fuel. It is therefore poorly tuned, and very inefficient.

It runs on a very rich air fuel ratio which allows it to run on poor quality or slightly contaminated fuel without causing premature combustion or knocking, and enables the engine to run cooler. This improves the longevity of the engine, however much of the inducted fuel passes through the cylinder uncombusted and the chemical energy is lost to the environment.

Due to the design and application of this engine this makes the Energy Balance Diagram atypical of most data in published literature, such as (Pulkrabek W., 2004) which states that a modern passenger car spark ignition engine will combust 95-98% of the inducted fuel, many engine designs will run at AFRs above the stoichiometric ratio in their low and mid-range to ensure complete combustion. At full load typical AFR data would be 13-14 to 1, slightly below the stoichiometric ratio to prevent knocking and overheating (Heywood, 1998).

The air cooled nature of the engine also makes it atypical of most engines for which published data is available. Most published data applies to water-cooled multi-cylinder engines, whereby as an estimated 35-40% of Q in does work resulting in 20-30% being converted to useful power, with mechanical and pumping losses accounting for the remainder, 30-35% is lost as heat to the engine coolant, and 25-30% is lost as heat in the exhaust gases (Pulkrabek W., 2004 and Crolla D., 2009).

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6.IGNITION AND COMBUSTION IN SPARK IGNITION ENGINES

Figure 6.1 Typical pressure versus crank angle diagram showing the main 4-stroke processes, indicating the range of crank angle over which combustion

occurs.

Figure 6.2 Detailed view of pressure versus crank angle diagram showing the sequence of combustion processes relative to crank position.

In a typical spark ignition engine the fuel and air are mixed during induction into the cylinder with the inlet valve open. The inlet valve is closed and the mixture is compressed, during which time the mixture heats enabling complete vaporisation of the fuel; consequently a homogenous evenly distributed mixture is formed within the cylinder. Through inlet port design and piston face design,

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motion is created during induction within the fuel mixture, creating swirl and tumble turbulence on a macro scale this aids the combustion process.

The spark plug is fired at between 10-30° (30° in Figure 6.1 and 6.2) before TDC (Heisler H., 1995), this causes a high temperature plasma flame nucleus to develop at the spark. There is an ignition delay period as much of this heat energy is absorbed by heating the surrounding air fuel mixture and spark plug electrodes. The flame nucleus develops into a flame embryo as the surrounding air fuel mixture heats and ignites.

Once formed, the flame embryo increases in size and the flame front increases in area. The pressure of the heated combustion products behind the flame front increases very rapidly as their temperature increases. This causes rapid expansion of the flame front outward from the ignition point towards the edge of the cylinder, creating turbulence on a micro eddy scale within the flame front. These micro eddies further accelerate the process as the effective surface area of the flame front is increased (Heisler H., 1995).

As the flame front approaches the edge of the cylinder, the piston reaches the top of its stroke, squishing the air fuel mixture from the edges into the centre of the piston. This creates yet more turbulence, spreading the flame front chaotically within the mixture again accelerating combustion.

Maximum cylinder pressure is reached 5-15° after TDC as 70-80% (Pulkrabek W., 2004) of the air fuel mixture completes combustion. Beyond 80% combustion the piston is accelerating in its downward stroke and the remaining uncombusted air fuel mixture at the edges of the piston is drawn away and diluted with combustion products by slowing turbulence eddies. This along with the increasing surface area of the relatively cool cylinder walls slows combustion ensuring the edges of the piston and the cylinder walls are protected from excessively high temperatures. It also ensures a smooth delivery and fall-off of torque as the expansion phase progresses, this reduces fatigue of the engine parts.

Poor engine design, tuning or poor quality fuel can all cause spontaneous combustion during the expansion phase called detonation or knocking, where shockwaves and high temperatures are experienced by the engine parts having detrimental effects on engine longevity.

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7.INTAKE VALVE FLOW INVESTIGATIONFigures 7.1 and 7.2 show a model developed in SolidWorks by superposition of diagrams and using the manufacturer’s dimension data from the Tecumseh engine manual. This model illustrates the arrangement of the L-shaped head, and the position of the valves in relation to the cylinder.

The main drawback of the L-shaped head design is the longer, contorted pathway that the gases must travel during the 4-stroke exchange process (see Figure 7.8). This creates greater molecular frictional pumping losses, sapping the finite pressure differential available on induction, and needed for good volumetric efficiency. This can be seen in the results of the dynamometer testing data in Table 3.3, where the engine only achieves a volumetric efficiency of approximately 25%, which rapidly drops off above 3300 psi. This design is therefore only suitable for lower engine speeds, which in its intended application this would not be a concern as it would most likely run at a constant low speed.

Given the test results, it would be best to make some assumptions for estimating the inlet valve diameter that correspond to this engine’s characteristics and application.

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Figure7.1 Model of Tecumseh TD110 Cylinder Block at Bottom Dead Centre showing inlet valve open

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Figure7.2 Model of Tecumseh TD110 at Top Dead Centre, showing inlet valve closed

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MACH NUMBERVolumetric efficiency is heavily dependent on the flow pathways the gases must travel to and from the cylinder. This includes the valve ports, which form one of the main choke points of the system. Engines rotate at high cyclic speeds drawing and expelling gases at high velocities into the cylinder. There is a physical speed limit to these speeds governed by the Mach number. The Mach number is a ratio related to the speed of sound, a physical constant inherent to system and its local conditions; whereby a Mach number of 1 is equal to the local speed of sound under the local conditions.

Sound propagates as a longitudinal compression wave, creating cyclical wave fronts as it travels formed by the molecules of the medium in which it is travelling. Each medium has an inherent, condition dependent speed at which sound waves travel naturally, similar to the way a coil spring will transmit pulses along its length. The local speed of sound in air is determined by Equation 7.1

Local Speed of Sound=√( γ ×GasConstant of Air × AbsoluteTemperature )

uair [m /s ]=√( γ ×R [ J /kgK ]×T [K ])

Equation 7.1

This natural speed also forms a speed limit in certain circumstances. One example is the movement of air in a duct, like an engine air intake. The moving air creates a pressure wave in front of it; this pressure wave is a high concentration zone of air molecules. These concentrated air molecules are closer together and therefore experience more collisions between themselves and the walls of the duct, creating increased dynamic friction. As the velocity of the air approaches the local speed of sound in the duct the molecules are forced so close together and they collide so frequently it is analogous to them rubbing; the friction of which causes a very rapid increase in heating. The heating dissipates the kinetic energy of the moving air as conducted heat and electromagnetic radiation, resulting in a loss of velocity.

When an air flow passes through a choke point or nozzle, such as a valve port, the Mach number relationship is compounded by pressure changes. In engines this sets the practical limit for an inlet valve port to a Mach number of 0.6 (Taylor C. 1985), however for this analysis of an engine that has already been established as volumetrically challenged Taylor’s preferred value of Mach number Z=0.4 shall be assumed as a practical limit.

FLOW COEFFICIENTThe valve has two critical aspects; its seat diameter (d) and the amount it lifts (l) from its seat to its fully open position. As the valve lift approaches ¼ of the valve

seat diameter, the valve curtain area (πdl )equals the valve seat diameter( π d2

4 ), 21


when adjusted for the effects of the valve stem and the shape of the ports separation effects, this leads to a practical limit for l/d of 0.4. This is demonstrated by Figure 7.3. which shows the relationship of C f vs l /d for a typical valve port. Beyond 0.4 the limiting factor becomes the valve seat diameter.

Figure 7.3 Characteristic flow coefficient curves for inlet and exhaust valves. (Ferguson C., Kirkpatrick A., 2013)

The valve flow coefficient (C f )is a ratio of the valves effective area (A¿¿ e)¿ in a given open position relative to an open unobstructed port equal to the valve seat area. (A¿¿ v)¿ in terms of their capacity to convey air through flow.

FlowCoeff icient=A e

A v

C f=Ae [mm ]Av [mm]

(Ferguson C., Kirkpatrick A., 2013) Equation 7.3

The inlet valve opening, the induction of air into the cylinder and the closing of the inlet valve are 3 synchronous intervals of the process of inducting the air into the cylinder. They have a finite duration dictated by the speed of the cam shaft which is proportional to the speed of the engine when operating within its practical working range. A mean flow coefficient is therefore formed by these three processes as, which could be calculated by integrating the area under an C f vs l /d plot for the process. For this investigation a value of C f=0.35 has been assumed to agree with Ferguson and Kirkpatrick .

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CALCULATION OF SUITABLE INTAKE VALVE PORT DIAMETERUsing Equation 7.1 the local speed of sound can be calculated:

Local Speed of Sound∈Air=√ (Gamma×SpecificGasConstant of Air ×Temperature )

uair [m /s ]=√ ( γ ×R [J /kgK ]×T [K ])

uair [m /s ]=√(1.40×287.1 [ J /kgK ]×293.15 [ K ] )=343.3 [m /s ]

It is assumed that the presence of the fuel droplets in the air has a negligible effect on the local speed of sound.

The maximum speed achieved in the testing before the volumetric efficiency, torque and power drop off was 3300rpm. Assuming this is the maximum engine speed by design the volumetric rate of change within the cylinder can be calculated.

V fuel /air=Sppiston×Piston Area

V fuel /air [m3/s ]=Sp piston [m /s ]× A p [m2 ]=Sppiston [m / s ]× π b2

4[m2]

V fuel /air [m3/s ]=5.42 [m /s ]×( π 0.0666924 )[m2]=18.93×10−3 [m3/s ]

Calculation 7.3

The Tecumseh engine is Carburettor fed prior to the inlet valve port, therefore the volumetric rate of change through the valve port can be assumed to be equal to the volumetric rate of change within the cylinder. A calculation can be performed to design a minimum effective inlet valve port area based on the assumed values for Mach number and mean flow coefficient discussed.

V fuel /air=Effectvive Inlet Valve Port Area×MachNumber ×uair

V fuel /air [m3/s ]=A e [m2 ]×Z×uair [m /s ]=C f A v [m2 ]×Z×uair [m / s ]

A v [m2 ]= V fuel/air [m3/s ]C f×Z×uair [m /s ]

A v [m2 ]=18.93×10−3 [m3 /s ]

0.4×343.3 [m / s ] =3.93×10−4 [m2 ]

Calculation 7.4

From this the inlet valve port diameter can be calculated:

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π ×( Inlet ValvePort Diameter )2

4=Av

π (d¿¿ i [m ])2

4=A v [m2 ]¿

d i[m ]=√( 4 A v [m2 ]π )

d i [m ]=√( 4×3.93×10−4 [m2 ]π )=0.02239 [m ]=22.39 [mm ]

Calculation 7.5

This gives a piston bore to inlet valve port diameter ratio:

d i [mm ]b [mm ] =

22.39 [mm ]66.69 [mm ] =0.336 :1

Calculation 7.6

According to Taylor’s table (Taylor C., 1985) this is a low ratio, and should be closer to 0.5:1 for a small 4-stroke engine. From Figure7.8, the detailed view of the model developed in SolidWorks, a measurement of 29.34mm for valve port diameter can be inferred, this would give a piston bore to inlet valve port diameter ratio:

d i [mm ]b [mm ] =

29.34 [mm ]66.69 [mm ] =0.440 :1

Calculation 7.7

This is more in agreement with Taylor’s data, and would account for losses that have not been included in the above calculations; these would include intermolecular friction losses and choked flow pressure drops discussed previously.

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Figure 7.8 Detailed Views from Figure 7.1 illustrating the contorted flow path the inducted air must follow (Section A-A), and the measurements used in

Calculation 7.7 (Section C-C)

8. HARMFUL POLLUTANTSBelow is a summary of five harmful pollutants found in the exhaust gas of internal combustion engines, their origin and their effects on humans.

C8H18+12.5O2+47N2→8CO2+9H 2O+47N2

Formula 8.1 Complete combustion of Octane in Air

During complete combustion of Octane in Air, only one pollutant is produced, Carbon Dioxide, CO2. Carbon dioxide is a naturally occurring component gas of the atmosphere; however it is only present at approximately 0.03% by volume. Concentrations above 4% are toxic to humans, causing hypercapnia and acidemia, which can lead to involuntary muscle twitching, delirium, hallucinations, nerve damage, heart damage, and eventually death. Carbon dioxide is also a Greenhouse gas which contributes to global warming.

In a real engine complete combustion is never truly achieved due to fuel impurities and tuning aspects and compromises to prolong longevity of the engine.

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bd


Gasoline+ Impurities+Oxygen+Nitrogen=Unburned∨Partially Burned Hydrocarbons (VOCs )+Oxides of Nitrogen (NOx )+CarbonMonoxide (CO )+Oxides of Impurities+Carbon Dioxide(CO2)+Water (H2O)

Formula 8.2 Incomplete combustion of Gasoline in Air

Most pollutants occur during incomplete combustion of fuels, these can include:

Unburned or Partially Burned Hydro Carbons, termed Volatile Organic Compounds, these include Octane, Benzene, which is added to fuel to prolong valve seat life, and other shorter chain hydrocarbons that result from incomplete combustion. Many of these chemicals are harmful or toxic to humans and the environment, causing mutations that lead to cancers, respiratory diseases and deformity in foetuses.

Carbon Monoxide, CO – A more toxic oxide of carbon, formed as an intermediary product during combustion, which remains when combustion is incomplete. Carbon monoxide has a higher affinity for the Iron reaction sites in blood haemoglobin that ordinarily attach to and transport oxygen around the body. Once attached to the reaction sites, the carbon monoxide cannot be removed by the normal chemical processes in the body, and the haemoglobin molecules become redundant.

Oxides of Nitrogen, N Ox include NO and N O2.NO Nitric Oxide is formed during combustion at high temperature. NO readily combines with more oxygen to formN O2 Nitrogen Dioxide. Both gases are acidic and cause irritation of the respiratory tract in humans, leading to long term respiratory diseases. They are soluble in water and contribute to the creation of Acid Rain which can cause leaf damage to plants and poison lakes and waterways.

Oxides of Sulphur, SOx include SO2 Sulphur Dioxide and SO3 Sulphur Trioxide. They are formed by the combustion of sulphurous impurities in fuels. Both gases are acidic and cause irritation of the respiratory tract and eyes in humans, leading to long term respiratory diseases and contributing to eye diseases. They are also soluble in water and contribute to the creation of Acid Rain which can cause leaf damage to plants and poison lakes and waterways.

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REFERENCESBoretti, A.A. and Nebuloni .P. (1993), Turbulence effects in reciprocating engines - Engineering Turbulence Modeling andExperiments. Elsevier Science

Crolla, D. (2009), Automotive Engineering - Powertrain, Chassis System and Vehicle Body, Oxford, Elsevier

Ferguson, C. Kirkpatrick, A. (2013), Internal Combustion Engines – Applied Thermosciences, Third Edition, Chichester, Wiley

Ganesan, V. (2012), IC Engines, 4th Edition, New Delhi, Tata Mcgraw-Hill

Heisler H. (1995), Advanced Engine Technology, London, SAE International

Heywood, J. (1988), Internal Combustion Engine Fundamentals, London, Mcgraw-Hill

Majer, V. and Svoboda, V. (1985), Enthalpies of Vaporization of Organic Compounds, Oxford, Blackwell Scientific Publications

Paul, S. (1998), An optimized alternative motor fuel formulation: natural gas liquids, ethanol, and a biomass-derived ether, New Jersey, Princeton University

Pulkrabek, W. (2004), Operating Characteristics - Engineering Fundamentals of the Internal Combustion Engine , 2nd Edition, .Prentice Hall

Rathakrishnan E. (2010), Applied gas dynamics, Singapore, Wiley

Rodgers, G. and Mayhew, Y. (2013), Thermodynamic and Transport Properties of Fluids, 5th Edition, Oxford, Blackwell Scientific Publications

SAE International, (2004), J1349AUG2008 Revised 2004-08, SURFACE VEHICLE STANDARD - Engine Power Test Code—Spark Ignition andCompression Ignition—Net Power Rating, SAE International.

Taylor, C. (1985), The internal-combustion engine in theory and practice, Second Edition, Cambridge, Mass. : M.I.T. Press,

Tecumseh, (1995), Tecumseh Technicians Handbook, 3 to 11HP 4-Cycle L-Head Engines, Tecumseh

TQ Education and Training Ltd., (2004), TD110-TD115 Test Bed and Instrumentation for Small Engines, TQ Education and Training Ltd.

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