performance and emission characteristics of a low heat ... · performance and emission...

Journal of Mechanical Science and Technology 26 (4) (2012) 1241~1250

www.springerlink.com/content/1738-494x DOI 10.1007/s12206-012-0206-0

Performance and emission characteristics of a low heat rejection

spark ignited engine fuelled with E20† C. Ramesh Kumar1,* and G. Nagarajan2

1CEAR, School of Mechanical and Building Sciences, VIT University, Vellore, 632014, India 2CEG, Anna University, Chennai, India

(Manuscript Received March 4, 2011; Revised August 30, 2011; Accepted December r 1, 2011)

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Abstract In internal combustion engines, the concept of low heat rejection (LHR) using thermal barrier coating on the surface of combustion

chamber is gaining attention. Thermal barrier coating reduces the heat transfer to the cooling system, protects engine components from peak heat flux and fluctuating temperature produced during combustion and improves the performance of the engine. Information in the literature is plentiful for LHR diesel engine and only few studies exist on LHR spark ignited engine. The application of thermal barrier coating in spark ignited engine is limited by pre-ignition and knocking due to elevated combustion chamber temperature. A spark ignited engine with moderate insulation on the combustion chamber and higher octane fuel can overcome this difficulty. The objective of the present experimental study is to quantify the changes in performance and emission characteristics brought by partial thermal insulation on the combustion chamber of a four stroke spark ignited engine fueled with E20 blend. Partial thermal insulation was created by coating 0.3 mm thick Alumina (Al2O3) on the cylinder head, inlet and exhaust valves. The changes are quantified with respect to unmodified engine fueled with gasoline. The combustion parameters such as flame development and rapid burn duration are also estimated and com-pared. The results indicate that partially insulated SI engine when fueled with E20 improves performance and reduces emission. A max-imum of 48% reduction in THC and 50% reduction in CO emission at part load was achieved.

Keywords: Combustion; Emission; E20; LHR; SI engine ---------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------- 1. Introduction

Among different alternate fuels available, ethanol has been recognized as a renewable alternate transportation fuel for SI engine. As it has properties similar to gasoline it can be used in SI engines with some or no modifications. Information in the literature is abundant for automotive use of ethanol. Pres-ently ethanol is primarily used as blending fuel with gasoline in many countries. Ethanol provides additional oxygen for air fuel charge and improves combustion. When ethanol is blended with gasoline (< 30% by volume) it reduces carbon monoxide, net carbon dioxide, and hydrocarbon emissions [1-4]. Ethanol reduces the cyclic variation of the MEP and im-proves the power and torque output [2]. Addition of ethanol increases the cylinder pressure and temperature and also de-creases the combustion duration [5]. Ethanol also improves the octane value of the fuel blend. A fuel with a higher octane number can endure higher compression ratio without knock-

ing and improve the thermal efficiency of the engine [6]. Etha-nol’s higher latent heat of evaporation increases the evapora-tive cooling effect of air-fuel charge which increases the volumetric efficiency and also decreases the compressed charge temperature during the compression stroke. Ethanol has higher hydrogen to carbon ratio because of which it pro-duces more volume of gases per unit of energy compared to gasoline (i.e. ethanol yields higher volume of total exhaust gas than gasoline). This leads to higher mean cylinder gas pressure which performs additional work during expansion. Unfortu-nately some of the physical properties of the ethanol present problems when it is adapted in higher volume or as a sole fuel in SI engines. The major engine operation issues with higher levels of ethanol in blended fuels are higher latent heat, cold start, cold start emissions and aldehyde emission. Ethanol’s higher latent heat of evaporation can also be disadvantageous in cold start situations. A blend containing 10% ethanol re-quires about 15.2% more manifold heating in order to attain proper degree (homogeneous mixture) of air fuel mixture. Higher levels of ethanol (greater than 30%) reduce the vapor pressure of blend and adversely affect the cold starts, engine warm-up and heating (light off) of the three way catalytic

*Corresponding author. Tel.: +91 9894189439 E-mail address: [email protected]

† Recommended by Associate Editor Kyoung Dong Min © KSME & Springer 2012

1242 C. R. Kumar and G. Nagarajan / Journal of Mechanical Science and Technology 26 (4) (2012) 1241~1250

converter. In an ethanol blend fueled engine exhaust alde-hydes are formed as an intermediate species form the post flame oxidation of the unburned fuel. During the end of com-bustion the flame front cannot reach the crevice volume as it gets extinguished by quenching effect near the cylinder wall. This makes the unburned hydrocarbons in crevice volume to undergo partial oxidation which results in aldehyde production. Higher ethanol concentration in the blend produces higher levels of aldehyde emissions. Aldehydes play an important role in formation of photochemical smog. To improve the cold start and cold emission a hotter environment in the combus-tion chamber will help in faster vaporization of the fuel and increase the temperature of the air fuel charge at the time of spark. Higher exhaust gas temperature helps the catalytic con-verter reach the lit-off temperature as quickly as possible.

Thermal efficiency of an internal combustion engine can be improved by reducing heat transfer to the coolant. Thermal barrier coating on to the combustion chamber surfaces of an internal combustion engine can suppress the heat transfer to coolant. When the combustion chamber of the internal com-bustion engine is coated with a thermal barrier material it is called as Low Heat Rejection Engine (LHR Engine). Lot of work has been done using the concept of thermal barrier coat-ing on diesel engines but few explored the effect on SI engine. Engine knocking and higher NOx emission rate limit the ap-plication of thermal barrier coating in SI engines. Some of the published work still shows potential for thermal barrier coat-ing in SI engines. Shiew Hwa Chan [8] conducted experi-ments on a SI engine coated with Zirconia & Yttria of varying composition and reported a reduction in fuel consumption by about 4.8% at lower load and improved thermal efficiency. Ramesh P Bola et al. [9] investigated ceramic coated two stroke SI engine fueled with gasoline and methanol. The au-thor was able to achieve reduction in CO emission 4% by Vol and improvement in brake thermal efficiency by about 20%. Krzystof Z Mendera [10] studied the heat transfer characteris-tics of plasma sprayed ceramic coated SI and CI engine and concluded that zirconia coatings were partially transparent to thermal radiation. The author also concluded that the deposi-tion of soot on the combustion chamber increases the radiative heat transfer as soot’s radiation is five times the radiation of gaseous combustion products in SI engines. Assanis and Mathur [11] reported that thin ceramic coating on the combus-tion chamber of an SI engine improved the brake power and emission. Parlak and Ayan [12] experienced knocking opera-tion at 8.2 compression ratio when they tested 0.15 mm zirco-nia coated piston in a four stroke SI engine. At lower com-pression ratio (below 8.2) the authors were able to achieve reduction in fuel consumption. The limitations of LHR SI engine can be improved by using fuel with higher octane value and latent heat of vaporization The shortcomings of ethanol blended fuel and LHR engine can be beneficial if ethanol blended fuel is used in a LHR engine. Ethanol blended fuels improves the octane value and also absorbs relatively higher amount of heat for its vaporization. This may bring

down the peak combustion temperature and lessen the produc-tion of oxides of nitrogen. The objective of this work is to obtain more quantitative information on performance and exhaust emission characteristics of E20 in a LHR spark ig-nited engine. Optimization of engine for ethanol blended fuel in low heat rejection operation is not focused in this study.

2. Materials and methods

2.1 Thermal barrier coating (TBC)

To create partial insulation (moderate) alumina was used as a thermal barrier coat material in the present investigation. The properties of alumina are given in Table 1. Before apply-ing the TBC, 400 μm of material is machined off from the cylinder head and the surface was grid blasted. The cylinder head is then coated with 100 μm NiAl bond coat and 300 μm Alumina using 40 kW atmospheric plasma spray. Spray speci-fications are given in Table 2. Several experimental studies [8, 9, 10, 11, 13] have proved that thin ceramic coats were effec-tive in engines. Thick ceramic coatings have reduced service life. Fig. 1 shows the picture of Alumina coated and stock (metallic) cylinder heads.

Table 1. Properties of Alumina.

Composition Al2O3

Purity Alumina 99.9%

Density 3.9 gm/cc

Melting point 2015oC

Specific heat at 100oC 930 J / kg K

Thermal conductivity 40 W/mK at 20oC

Thermal shock index 0.2

Thermal cycle index 0.8

Flexural strength 380 MPa

Hardness HV 1500 kg f / mm2

Tensile strength 262 MPa

Poison ratio 0.26

Young’s modulus 370 GPa

Co-efficient of thermal expansion 8 μm / m oC

Table 2. Plasma spray specifications.

Particle velocity 500 – 550 m/s

Oxide content 1 – 2 %

Porosity 1 – 8 %

Deposition rate 1 – 5 Kg / hr

Current 530 A

Voltage 72 V

Spray distance 100 mm

Torch nozzle diameter 6 mm

C. R. Kumar and G. Nagarajan / Journal of Mechanical Science and Technology 26 (4) (2012) 1241~1250 1243

2.2 Test fuel preparation

Commercially available unleaded gasoline was blended with ethanol for preparing fuel blend on a volume basis. The purity of ethanol is 99.6%. E20 is formulated using 20% etha-nol + 80% unleaded gasoline. Table 3 shows some of the properties of the base test fuels and Table 4 shows some of the properties of blended fuel. The property of pure ethanol was obtained from the manufacturer. In order to avoid absorption of moisture by ethanol in the blend, test fuel blends were pre-pared five minutes before experimentation.

2.3 Experimental setup

The experimental work was done on a single cylinder, car-bureted, S.I. engine. The selected engine is a good representa-tive of non-hand held utility engine found in lawn movers,

generators, agriculture pumps, compressor sets and two wheelers. Most of the two-wheelers manufactured in India are powered by small carbureted, 4-stroke SI engines. The speci-fication of the engine and complete schematic of the experi-mental setup is given in Table 5 and Fig. 2.

An eddy current dynamometer which was controlled by a digital controller was connected to the engine. The dynamic cylinder pressure was measured by a piezo-electric water cooled pressure transducer (Kistler make, model: 601A). Calibrated high precision weighing scale was used to measure the fuel consumption rate. The air flow rate was measured using a hot film anemometer type air mass flow sensor. Spark plug spikes were captured using an inductive coil pickup to monitor the ignition timing. HORIBA MEXA 554 JA gas analyzer was used to measure HC, CO, CO2 concentration in the exhaust. For NOx measurement a chemical cell NOx ana-lyzer was used. During the test engine speed was maintained constant and the torque was varied in order to achieve the required test point. The engine speed is maintained at 65% of the rated maximum speed, which is in accordance with ISO 8178 – 4 G1 cycle [14]. It is a six mode cycle test, a mode being part of test cycle with defined speed and torque. Table 6 shows the engine test matrix used in this study. In order to change the engine configuration from base line to LHR, con-ventional metallic cylinder head is replaced with the head

Table 3. Properties of fuel.

Chemical formula C2H5OH C4 to C12

Molecular weight [g/mol] 46.07 100–105

Carbon [mass%] 52.2 85–88

Hydrogen [mass%] 13.1 12–15

Oxygen [mass%] 34.7 0

Liquid density, 20 C [kg/l] 0.792 0.745

Boiling temperature, 1 atm [°C] 78.4 27–225

Reid vapor pressure, [kPa] 16 50–100

Flammability limit 20 °C [vol %] 3.3-19 1.0–8.0

Stoichiometric Air / Fuel ratio 9 14.5–14.7

Autoignition temperature [°C] 423 257

Heat of vaporization [kJ/kg] 910 330-400 Heat of combustion

(Lower heating value) [kJ/kg] 26900 43000

Research octane no. 108 91

Table 4. Properties of different ethanol blends.

Blend Density (kg/m3)

Stoichiometric Air/Fuel

Calorific value

(MJ/kg)

Vapor pressure

(kPa) Gasoline 745 14.7 43 64.63

E20 754.2 13.54 39.8 70.01

Fig. 1. Picture of Alumina (Al2O3) coated and stock cylinder heads.

Table 5. Specifications of the engine.

Engine type : Single cylinder, SI engine

Cooling : Air cooled

Cylinder bore : 79.28 mm

Stroke : 61.67 mm

Max. power : 7.46 kW @ 4000 rpm

Max. torque : 18.7 Nm @ 2600 rpm

Compression ratio : 8:1

Valve arrangement : Two vertical over head valves

Fig. 2. Schematic of the experimental setup.


coated with Al2O3. Cylinder head gaskets and lubricating oil were changed with new identical one during head changeover. For performance tests and raw emissions measurements, data was collected under steady state conditions. Measurements were made according to ISO 8178- 4; G1 cycle. The measured power was corrected to the standard atmospheric conditions.

2.4 Heat release

One-zone heat release analysis based on first law of ther-modynamics was carried out using dynamic in-cylinder pres-sure and volume data. Heat transfer effect is also included in order to estimate close approximate values of rate of heat re-lease. Heat transfer by conduction through piston rings and skirt is not focused in this study. Instantaneous convective heat transfer coefficient is estimated through Woschni’s corre-lation (15).

ch htdQ γ dV 1 dP dQp V

dθ γ 1 dθ γ 1 dθ dθ= + +

− − (1)

5 8 2g g1.338 6.0* 10 * T 1.0* 10 * Tγ − −= − + (2)

cyl cylg

char

P VT

m R= (3)

( )htwall g w

dQ h A T Tdθ

= − (4)

( ) ( ) ( ) ( )b 0.75 1.62b bb 1gh θ a P θ T θ D W θ− −= (5)

( ) ( ) ( )( )d r1 p 2 m

r r

v TW C V C P θ P θPV

θ⎡ ⎤

= + −⎢ ⎥⎣ ⎦

(6)

The S – shaped mass fraction burn curve represents the per-

centage of fuel- air charge consumed versus crank angle dur-ing combustion. Mass fraction burned is obtained by dividing cumulative gross heat release with maximum value of cumu-lative gross heat release which occurs at the end of combus-tion to get unity at its maximum value. The flame develop-ment angle (ST angle to 10% MFB angle) and rapid burning angle (10% MFB to 90% MFB) are estimated using mass fraction burned [16].

3. Results and discussion

Four experiments, two for each configuration (baseline and ceramic coated) using gasoline and E20 were conducted. The impacts of partial insulation on combustion chamber with fuel variation was assessed by comparing and analyzing the char-acteristics of engine performance with respect to fuel conver-sion efficiency, exhaust gas temperature, mass fraction burned and exhaust emissions under various steady state operations. Ignition timing of the engine was maintained at MBT. Fig. 3 shows the spark timing for best torque at various modes.

3.1 Brake thermal efficiency

The variation of brake thermal efficiency with brake power is shown in the Fig. 4. In order to study the impact of thermal barrier coating on combustion chamber with fuel variation, two fuels: base gasoline and higher octane fuel blend E20 were used in this study. From the test results, it was observed that there was a considerable increase in thermal efficiency in LHR engine at all modes using gasoline and E20 fuel when compared to base engine fueled with gasoline. Significant improvements in brake thermal efficiency is obtained by LHR engine with E20 fuel due to the combined effect of thermal insulation, lower flame temperature & mean cylinder tempera-

Table 6. Engine test matrix.

Mode 1

Mode 2

Mode 3

Mode 4

Mode 5

Mode 6

Speed (65% of rated)

2600 2600 2600 2600 2600 -

Power (kW) 6.34 4.77 3.17 1.58 0.634 Idle

% Rated power 85 75 50 25 10 -

Torque 100% 75% 50% 25% 10% 0 Weighting

factor 0.09 0.2 0.29 0.3 0.07 0.05

Fig. 3. Variation of brake thermal efficiency with brake power.

Fig. 4. Variation of brake thermal efficiency with brake power.


ture, combustion characteristics, increased oxygen and lower heating value of the fuel (which is 9.25% lower than gasoline). At lower loads the difference in the brake thermal efficiency is marginal for both baseline and LHR engine fueled with gaso-line and E20. The effect of thermal insulation was realized at higher loads as the energy is more effectively utilized. As the load was increased the amount of heat released increases due to increased air-fuel induction and the heat loss to cooling system may increase when the engine is maintained at con-stant speed. Thermal insulation on the cylinder head reduces the loss of heat during the start of combustion and as a result additional heat which could have lost through cylinder head is converted into indicated power. Since friction is pro-portional to speed, the engine output increases when the en-gine is running at constant speed (within expansion efficiency limit). For example brake thermal efficiency of the engine increases from 22.9% to 26% with LHR and E20 fuel, and from 22.9% to 24.5% with LHR and gasoline fuel at 4.77 kW when compared to baseline engine. At the same power output, the efficiency of the base line engine improves from 22.9% to 24% when fueled with E20 fuel.

3.2 Exhaust gas temperature

If the heat loss to the cylinder walls could be suppressed, such heat could be converted into indicated power to the effi-ciency of the expansion alone, which is 30 – 40% and the remaining of the heat so recovered would be rejected to the exhaust after expansion [17]. The exhaust gas temperature from the engine in both the configurations followed expected trends. Fig. 5 shows the comparison of exhaust gas tempera-ture between unmodified engine and LHR engine. Gasoline and E20 operation in LHR engine registers a higher tempera-ture than baseline engine at all loads. Exhaust gas temperature of gasoline in ceramic coated engine is higher by 24°C at no load, 13°C at part load and 23°C at maximum load than gaso-line in baseline engine. For E20 the difference was 28°C at no load, 8°C at part load and 16°C at maximum load. At 4.77 kW exhaust temperature reaches to a maximum of 631°C with

gasoline and 615°C with E20 in LHR engine.

3.3 Combustion characteristics

Increasing load in SI engines is a result of increase in throt-tle angle and increased mass flow rate of air-fuel mixture. As the mass flow rate of air-fuel mixture increases, varying load results in significant change in heat release. By understanding the influence of load and fueling differences (change in physi-cal and chemical properties of the fuel) on heat release, per-formance changes due to partial insulation (change in combus-tion chamber thermodynamic conditions) will be able to be isolated and analyzed. Fig. 6 presents the variation of heat release rate (HRR) and Mass Fraction Burned (MFB) with crank angle at different engine loads. For both baseline and LHR engines at lower loads the heat release rate profiles are wider. With increase in load, heat release rate increases and peak heat release takes place closer to TDC when fueled with gasoline and E20 fuels. It is also noted that with increase in load, post combustion duration is shortened as heat release rate drops quickly after reaching peak heat release. Lowest peak heat release rate occurred with E20 fuel at lower loads. It is very clear from the heat release profiles that instantaneous heat release rates are higher in LHR engine due to accumula-tion of additional heat which would have lost in cooling sys-tem. With gasoline, and E20 around 16 to 20% increase in peak heat release rate at various engine load condition was observed in LHR engine when compared to heat release rate of gasoline in baseline engine. The normalized heat release given by mass fraction burned (MFB) curve explains the rate at which the air-fuel mixture is consumed. The slope of the curve increases after the spark timing and reaches maximum at 50% of MFB which is halfway through burning process and decreases to zero as the combustion process gets completed.

Fig. 7 shows the variation of mean gas temperature with re-spect to crank angle at 0.63 and 6.3 kW. Fig. 8 shows the variation of flame development duration (calculated from 0 to 10% of MFB) with brake power. When baseline engine is fueled with gasoline, flame development duration decreases with increase in the load due to higher mean temperature of the burning gas, and increased swirl and turbulence. An aver-age reduction in flame development duration by 12% was realized with increase in load by a factor of 1.7 till 75% of the maximum load. At wide open throttle (maximum load) due to higher swirl ratio, flame development period is susceptible to flame stretch and excessive flame stretch may lead to unsteady burning and flame quench. Higher swirl ratio also leads to loss in volumetric efficiency as the heat transfer from the combus-tion chamber surface to the in coming air-fuel mixture in-creases during suction. This leads to longer flame develop-ment period at maximum load. Ethanol though has properties similar to gasoline, when mixed with gasoline, its higher latent heat and reduced lower heating value has significant effect on flame development duration. At lower loads with E20, a rela-tive increase in flame development duration by about 5% with

Fig. 5. Variation of exhaust gas temperature with brake power.


respect to pure gasoline was noticed due to reduced energy release. Twenty percent of ethanol when added to gasoline reduces the heating value of the blend by 9.25% when com-pared to pure gasoline. At 3.2 kW and 6.3 kW loads, relative reduction in flame development duration by about 11% and 18% when compared to gasoline was noticed. It is clear that amount of heat released has strong influence on ethanol flame development duration. At 4.8 kW a relative increase by 7.2% was noticed in flame development period with ethanol – gaso-line blended fuel. The reason was not very clear and it was suspected that variation in charge induction and increased

residual gases during each cycle could be the reason. Except at 4.8 kW, an average reduction in flame development duration by 14% when compared to lowest load was realized with in-crease in load by a factor of 1.7 when the baseline engine is fueled with E20. When combustion is taking place, if the cen-ter of combustion (50% MFB) is positioned earlier to TDC, combustion counteracts compression process and work done by the piston on the charge increases.

If center is positioned after TDC the thermodynamic expan-sion ratio becomes lesser than compression ratio leading to reduced work output as greater part of the positive work is

Fig. 6. Variation of Heat release and MFB with respect to crank angle at different engine loads.

(a) (b) Fig. 7. Variation of rate of mean gas temperature with respect to crank angle at different engine loads.


done by the gases on piston only during the process of expan-sion. This also reduces the magnitude of the peak pressure occurring in expansion stroke. Fig. 9 shows the comparison of 50% mass fraction burned angle with respect to brake power. For baseline engine when fueled with gasoline the 50% MFB angle appears 10° to 18° after TDC and the same engine when fueled with E20 the angle appears 9° to 23° after TDC. Fig. 10 shows the variation of rapid burn duration (calculated from 10 – 90% of MFB) with brake power. An average reduction in

rapid burn duration by 29% when compared to lowest load was realized with increase in load by a factor of 1.7 when the baseline engine is fueled with gasoline. Though flame devel-opment duration of gasoline in baseline engine is better than E20, the following 80% MFB ends up burning slower than E20. This clearly shows that with E20, once the flame is es-tablished, combustion is faster than gasoline. The reason is attributed ethanol’s oxygen which augments the combustion rate and because of this large amount of fuel burn takes place closer to TDC. Except at 0.634 kW, rapid burn duration of E20 is reduced by 1.6 to 5.7% at various engine loads com-pared to gasoline. With E20 an average reduction in rapid burn duration by 40% when compared to lowest load was realized with increase in load by a factor of 1.7.

It is also noted that E20 in baseline engine has the longest rapid burn duration at lowest load (0.634 kW). During the early stage of combustion, burned gas behind the flame front is hotter than unburned gas and the burned gas is in the state of violent commotion so that heat is conveyed very readily by convection (17). During the same period as the area of the cylinder wall is smaller than the area of cylinder head and cylinder head is comparatively at low temperature, heat may be lost from the burned gas through the cylinder head. Ther-mal barrier coating on the surface of the cylinder head reduces the heat loss and increases the energy available in the burned gas. This increases the temperature and decreases the density of burned gases. As a result burned gas expands to occupy more volume and compress the unburned gases. This results in compressive heating and causes the chemical reaction time to decrease and flame front speed to increase. The flame devel-opment angle is reduced consistently by about 1 to 6° crank angle with gasoline and 5 to 6° crank angle with E20 in LHR engine when compared to baseline engine fueled with gasoline. In general an average reduction in flame development dura-tion by 18% when compared to lowest load was realized with gasoline with increase in load by a factor of 1.7. With E20 an average reduction in flame development duration by 21% when compared to lowest load was realized with increase in load by a factor of 1.7.

The use of thermal barrier coating therefore improves the time lag associated with the flame development for substantial heat release. Once this has occurred, the time taken to com-plete combustion is relatively fast. For LHR engine when fueled with gasoline, the 50% MFB appears about 3 to 14° after TDC and when fueled with E20 it appears 3 to 12° after TDC at various engine loads. Rapid burn duration of the LHR engine is reduced by 8 to 18% with gasoline and 5 to 20% with E20 at various engine loads when compared to baseline engine fueled with gasoline. With gasoline an average reduc-tion in rapid burn duration by 30% when compared to lowest load was realized with increase in load by a factor of 1.7. With E20 an average reduction in rapid burn duration by 20% when compared to lowest load was realized with increase in load by a factor of 1.7.

Longer flame development and rapid burn durations are ob-

Fig. 8. Variation of flame development angle with brake power.

Fig. 9. Variation of 50% MFB angle with brake power.

Fig. 10. Variation of rapid burn angle with brake power.


served in baseline engine fueled with gasoline and E20. The reason is attributed to the heat loss during the early stage of combustion which lowers the energy available during the pre-ignition phase and as a result the reaction rate and flame ker-nel growth slows down. In general it was observed that due to the thermal barrier coating, the flame development duration is reduced by an average of 14 to 28% and rapid burn duration reduced by an average of 9 to 10% when LHR engine is fu-eled with gasoline, and E20.

3.4 HC emission

Unburned hydrocarbons emission in internal combustion engines result from multiple processes. Possible sources of HC emission in SI engines are Quench layers, Crevice volume, lubricating oil layers and combustion chamber deposits. Quench layers are regions very close to cooled engine walls through which a flame during premixed combustion cannot propagate. The unburned charge closer to cooled engine walls (which cannot be approached by propagating flame) exits combustion chamber as partially oxidized hydrocarbon or as unburned fuel. During compression and combustion process crevice volumes are filled with unburned charge. As the pre-mixed flame cannot reach this volume due to quenching effect, unburned charge leaves crevice during expansion. The maxi-mum contribution of the quench layer and crevice to HC emissions during normal engine operation will take place un-der idle and low load conditions due to lower cylinder pres-sure and peak temperature. If the temperature of the burned gas is high enough, the exiting unburned charge can be oxi-dized.

Fig. 11 shows the comparison of THC emission with re-spect to brake power. It is observed that THC emissions are reduced in LHR engine with both gasoline and E20. THC is reduced by about 2.3 to 17.6% with gasoline and 15 to 48% with E20 at various loads when compared to base engine fu-eled with gasoline. These substantial reductions of THC with E20 are obtained due to additional oxygen, improved combus-tion and increased combustion chamber wall temperature. At

lower loads for both the configurations with both gasoline and E20, high levels of hydrocarbon emission was registered due to rich mixture, slow initial flame propagation rates, lower heat release rate and lower combustion chamber temperature. Even then LHR engine has marginal hydrocarbon emission reduction at lower loads when compared to baseline engine. Again at the maximum load HC emission level exhibits an increasing trend caused by increased fuel consumption and increased bulk quenching at combustion chamber wall at the end of combustion.

3.5 CO emission

Carbon monoxide is an intermediate product of combustion of any hydrocarbon fuel. It is produced by rich combustion at both full-load and idle operation in an SI engine. CO is also produced by partial oxidation of unburned hydro-carbons during the exhaust stroke as well as dissociation of CO2 pro-duced during combustion. Reduction of CO can be achieved by improved fuel-air management. A comparison of CO emis-sion with brake power for LHR and unmodified engine is shown in Fig. 12. The reduction in CO level is consistent at various operating conditions with LHR engine over base en-gine when fueled with gasoline and E20. The trend of the curve also tallies with THC emission and fuel conversion efficiency. The reduction in CO is about 3 to 16% with gaso-line and 20 to 50% with E20 when compared to base engine fueled with gasoline. The additional oxygen brought by etha-nol and higher combustion chamber temperature (due to insu-lation) are contributing towards improved combustion. Reduc-tion in CO about 3 to 40% is also recorded with base engine when fueled with E20. As stated earlier, oxygen enrichment resulting from ethanol in E20 contributes to the CO reduction in base engine also. At higher loads CO increases marginally in both the configurations with both the fuels due to rich mix-ture combustion. The CO emission from this engine with and without thermal barrier coating is below the regulatory thresh-old. The current regulatory threshold for CO is 610 g / kW-hr.

Fig. 11. Variation of THC with brake power.

Fig. 12. Variation of CO with brake power.


3.6 NOx emission

Fig. 13 shows the variation of brake specific NOx levels in engine exhaust. Nitric oxide (NO) which is also referred to as NOx, is formed during combustion as the high flame tempera-ture break down molecular O2 and N2 from the inducted air, which then recombine into NO (the level of NO2 formed in spark-ignited engines is negligible). The production of NO depends primarily on the peak temperature achieved during combustion (18). An increase in NOx emissions was observed at all modes in LHR engine. Thermal barrier coating reduces the heat transfer to the cylinder head and elevates the peak temperature of the gas inside the combustion chamber. Be-cause of this combustion chamber wall temperature and resid-ual gas temperature are higher. During end of induction stroke, the temperature of the air-fuel charge increases as it is heated by residual gas and combustion chamber wall. Since the tem-perature of air-fuel charge at the end of induction and peak gas temperature inside the combustion chamber are higher in LHR engine NOx production gets accelerated. Further additional oxygen brought by ethanol in E20 fuel also favors the NOx formation when E20 is fueled in ceramic coated engine. NOx

emission increases by about 6 to 50% with gasoline and 5 to 46% with E20 in ceramic coated engine when compared to baseline engine fueled with gasoline. Baseline engine with E20 showed lowest NOx emission by about 6 to 22% lesser when compared to baseline engine fueled with gasoline. This is attributed to lower peak temperature and reduced charge temperature brought by higher latent heat of vaporization of ethanol in E20 fuel.

4. Conclusions

A thin ceramic coating of 0.3mm thickness using alumina was applied on the cylinder head surface of the combustion chamber and experiments were conducted. On the basis of results of the experiments conducted the following conclu-sions are drawn:

• Using thermal barrier coating a maximum increase in thermal efficiency by about 7% using gasoline 13.5% using

E20 was achieved at part load when compared to baseline engine fueled with gasoline and E20.

• Heat release rates are better due to LHR and MFB curves are steeper than baseline engine, which resulted in improved brake thermal efficiency.

• The Flame development and rapid burn duration of LHR engine fueled with gasoline and E20 are shorter than baseline engine fueled with gasoline and E20 fuel.

• The maximum reduction in THC emissions is in the range of 2.3 to 17.6% using gasoline and 15 to 48% using E20 de-pending upon the mode with thermal barrier coating when compared to baseline engine fueled with gasoline and E20.

• The maximum reduction in CO emissions is in the range of 3 to 16% using gasoline and 20 to 50% using E20 depend-ing upon the mode with LHR when compared to baseline engine fueled with gasoline and E20.

• Through significant improvements were attained using thermal barrier coating, it produces higher NOx levels when compared to baseline engine. A maximum of 40% increase in NOx is registered at 75% of the maximum torque.

In summary, the study revealed that the engine with partial

thermal insulation was found to be superior when operated with E20 from the aspect of brake thermal efficiency, combus-tion and emission characteristics. This experimental study covered a narrow range of engine operating conditions and more research is required to explore the effect of TBC on SI engines with NOx reduction technique.

Abbreviation

BDC : Bottom dead center BTDC : Before top dead center CA : Crank angle CO : Carbon monoxide HC : Hydrocarbon IC : Internal combustion IMEP : Indicated mean effective pressure NOx : Nitrogen oxides SI : Spark-ignition ST : Spark timing TBC : Thermal barrier coating TDC : Top dead center WOT : Wide open throttle

Nomenclature------------------------------------------------------------------------

D : Cylinder bore dia h : Convective heat transfer coeff. (W/m2 K) N : Engine speed (rpm) Pr : Pressure at reference state Sp : Mean piston speed, m/s Tw : Effective wall temperature (K) Tg : Mean gas temperature at point (K) Ti : Bulk gas temperature at point i (K)

Fig. 13. Variation of NOx with brake power.


Tr : Temperature at reference state Vd : Engine displacement volume Vr : Volume at reference state Wc : Work per cycle W : Effective gas velocity

htdqdθ

: Heat transfer, J / ° CA

(P(θ) - Pm(θ)) - Pressure raise due to combustion C1 = 6.18, C2 = 0 - for gas exchange C1 = 2.28, C2 = 0 - for compression period C1 = 2.28, C2 = 3.24×10-3- for combustion Abbreviations used in graphs

Gas. base Baseline unmodified engine with gasoline fuel

Gas. LHR Partially insulated engine with gasoline fuel

E20. base Baseline unmodified engine with E20 fuel

E20. LHR Partially insulated engine with E20 fuel

References

[1] N. Jeuland, X. Montagne and X. Gautrot, Potentiality of ethanol as a fuel for dedicated engine, Oil & Gas Science and Technology – Rev. IFP, 59 (6) (2004) 559-570.

[2] M. A. Ceviz and F. Yüksel, Effects of ethanol – unleaded gasoline blends on cyclic Variability and emissions in an SI engine, Elsevier - Applied Thermal Engineering, 25 (2005) 917-925.

[3] F. Yüksel and B. Yüksel, The use of ethanol – gasoline blend as a fuel in an SI engine, Elsevier - Renewable Energy, 29 (2004) 1181-1191.

[4] M. Al-Hasan, Effect of ethanol – unleaded gasoline blends on engine performance and exhaust emission, Elsevier - En-ergy Conversion and Management, 44 (2003) 1547-1561.

[5] H. Bayraktar, Theoretical investigation of flame propagation process in an SI engine running on gasoline–ethanol blends, Elsevier - Renewable Energy, 32 (2007) 758-771.

[6] Hakan Bayraktar, Experimental and theoretical investigation of using gasoline–ethanol blends in spark-ignition engines, Elsevier- Renewable Energy, 30 (2005) 1733-1747.

[7] J. A. A. Yamin, H. N. Gupta and B. B. Bansal, The effect of combustion duration on the performance and emission char-acteristics of propane-fueled 4-stroke s. i. engines, Emirates Journal for Engineering Research, 8 (1) (2003) 1-14.

[8] S. H. Chan, Performance and emission characteristics of a partially insulated gasoline engine, International Journal of Thermal Sciences, C40 (2001) 255-261.

[9] R. B. Poola, B. Nagalingam and K. V. Gopalakrishnan, Per-formance of thin ceramic coated combustion chamber with

gasoline and methanol as fuels in a two stroke S. I. engine, SAE Paper No. 9410173 (1994).

[10] K. Z. Mendera, Effects of plasma sprayed Zirconia coatings on Diesel engine heat release, Journal of cones, Internal combustion engines, 7 (1-2) (2000).

[11] D. N. Assanis and T. Mathur, The effect of thin ceramic coatings on spark-ignition engine performance” SAE Techni-cal Paper Series, 900903 (1990).

[12] A. Parlak and V. Ayhan, Effect of using a piston with ther-mal barrier layer in a spark ignited engine, Journal of Energy Institute, 80 (4) (2007) 223-228.

[13] I. Taymaz, The effect of thermal barrier coatings on diesel engine performance, Surface and Coatings Technology, 201 (2007) 5249-5252.

[14] ISO 8178, International standard, Reciprocating Internal combustion – Exhaust Emission Measurement – Part-4; Test cycles for different engine applications.

[15] Woschni, A universally applicable equation for the instan-taneous heat transfer coefficient in the internal combustion engine, SAE Paper No. 670931 (1967).

[16] J. B. Heywood, Internal combustion engine fundamentals, Singapore: McGraw-Hill International Editions (1988).

[17] H. R. Ricardo, The high-speed internal-combustion engines, Blackie & Sons Limited, London (1931).

[18] E. Sher, Hand book of air pollution from internal combus-tion engines – Pollution formation and control, Academic Press, London (2010).

C. Ramesh Kumar received his BE (Mechanical Engineering) and ME (Thermal Engineering) degrees in 1999 and 2002. At present he is working as Senior Assistant Professor in School of Mechanical and Building Sciences, VIT University, Vellore, India. He has ten years of industrial and teaching experi-

ence. His areas of interest are low heat rejection internal com-bustion engines, alternate fuels and engine simulation.

Dr. G. Nagarajan received his BE (Mechanical), ME (I.C Engines) and Ph.D degrees in 1986, 1988 and 2000 form Anna University, Chennai. He currently works as a Professor in Me-chanical, Anna University. He has more than 27 years of industrial, research and teaching experience. His areas of interest

includes combustion engines, alternate fuel and engine elec-tronics.

performance and emission characteristics of a low heat ... · performance and emission...

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