53_SI-engine-fuelled-by-biogas.pdf

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Seminar Report on

EFFECTIVE METHODS TO IMPROVE PERFOMANCE AND EMISSIONS OF SI ENGINE

FUELLED BY BIOGAS


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ABSTRACT

For understanding performance and combustion parameters at various

compression ratio a single cylinder diesel engine was modified to operate as a biogas operated spark ignition engine. The engine was operated at 1500 rpm at throttle opening of 25% and 100% at various equivalence ratios. The performance, emissions and combustion characteristics with different compression ratios are compared. It has been found from the results that the higher the compression ratio, the higher the brake thermal efficiency. When the compression ratio was above a critical value of 13:1, brake power and thermal efficiency increased little. At higher compression ratios above 13:1, increased NOx, HC, and CO emissions were measured .In the study of hydrogen blending in biogas engine a 8-L spark ignition engine fueled by biogas with various methane concentrations which we called the N2 dilution test was performed in terms of its thermal efficiency, combustion characteristics and emissions. The engine was operated at a constant engine rotational speed of 1800 rpm under a 60 kW power output condition and simulated biogas was employed to realize a wide range of changes in heating value and gas composition. The N2 dilution test results show that an increase of inert gas in biogas was beneficial to thermal efficiency enhancement and NOx emission reduction. H2 fractions ranging from 5 to 30% were blended to the biogas and the effects of hydrogen addition on engine behavior were evaluated. The engine test results indicated that the addition of hydrogen improved in-cylinder combustion characteristics, extending lean operating limit as well as reducing THC

emissions while

elevating

NOx

generation.

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CONTENTS

CHAPTER NO PAGE NO

1 INTRODUCTION 6

1.1 LITERATURE SURVEY 8

2 VARIABLE COMPRESSION RATIO TEST 9

2.1 EXPERIMENTAL SETUP AND EXPERIMENTS 9

2.2 RESULTS AND DISCUSSSIONS 13

3 EFFECT OF HYDROGEN BLENDING 18

3.1EXPERIMENTAL SETUP 18

3.2RESULTS AND DISCUSSIONS 21

3.3 HYDROGEN ADDITION TEST-LEAN ENGINE OPERATION 27

4 CONCLUSIONS 31

REFERENCE

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LIST OF FIGURES

FIGURE PAGE NO

2.1 .Experimental setup of variable compression ratio test. …………… . 10

2.21 .Variation of brake power with equivalence ratio at full throttle….. … 13

2.22Variation of brake thermal efficiency

with equivalence ratio at full throttle …………………………………..… .14

2.23 Variation of brake thermal efficiency

with equivalence ratio at part throttle……………………… …………. . 15

2.24Variation of HC emissions with equivalence ratio at full throttle….. 16

2.25Variation of NOx emissions with equivalence ratio at full throttle…… 17 3.1.Experimental setup of hydrogen blending test ……………………… 19

3.21Fow rate of each component of biogas with N2 dilution ratio …… … ..21

3.22Variation of thermal efficiency, combustion stability, manifold air pressure With N2 dilution ratio……………………… 22

3.23Variation of HC, NOx emissions with N2 dilution ratio………… ...... …23

3.24.Flow rate of each component of hydrogen – biogas blend fuel With hydrogen concentration ………………………………………… ……24 3.25.Variation of thermal efficiency with H2 concentration…………… …. 25

3.26Effect of hydrogen addition on THC and NOx emissions………… …26 3.31 Variation of thermal efficiency for various H2-biogas blend fuels As a function of excess air ratio ………………………………… ……… .…27 3.32Variation of THC emissions with excessair ratio……………….… … …29 3.33Variation of NOx emissions with excess air ratio…………………… …30

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LIST OF TABLES

TABLE NO TABLE NAME PAGE NO

2.1 Base engine Specifications of Variable compression ratio test……………10

3.1 Base engine Specifications of Hydrogen blending test… ……………… .…19

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NOMENCLATURE

COV Coefficient of variation CR Compression ratio EAR Excess air ratio HC Hydrocarbon

IMEP Indicated mean effective pressure MBT Minimum advance for best torque NO X Nitrogen oxides

BTDC Before top dead centre

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CHAPTER 1

INTRODUCTION

Biogas is typically composed of 40-60% of methane and rest of inert gases

such as nitrogen(N2) and (CO2).however its main composition varies by origin.There for it is difficult to establish optimum and consistent air- fuel mixture conditions and combustion phasing in order to ensure a stable operation of I C engines. In addition a significant amount of inert gases contained in biogas affects in-cylinder combustion charecteristics,reducing flame propagation speed which increase average length of combustion duration and ignition delay.One of the efforts to overcome these drawbacks is a rise of compression ratio in a biogas engine.it enables improvements in engine performance, especially in thermal efficiency and power output. However, an increased compression ratio can also exacerbate knock tendency and produce more nitrogen oxides (NOx) and hydrocarbon (HC) emissions.Although the presence of inert gases in biogas can suppress knock vulnerability and NOx emission, modifying piston (in an SI based engine) or replacing the diesel injectors by spark plugs (in a CI based engine) to achieve a high compression ratio will require extra cost.

Another approach for better engine performance is a biogas diesel dual fuel engine.In this case, mixture of gaseous fuels and fresh air is supplied to a cylinder and then, a small amount of diesel-like fuel is injected to ignite the combustible mixture. Since it is usually converted from a diesel CI engine, the dual fuel engine has a high

compression ratio. This means that it can achieve higher efficiency than a biogas dedicated SI engine. In addition, the injected diesel fuel behaves as multi-point ignition sources so that higher burning speed and relatively complete combustion can occur inside a combustion chamber. The other way to improve biogas engine performance is to add hydrogen (H2) in gaseous fuels. Hydrogen is well known for its excellent combustion characteristics and has been considered as a combustion enhancer for a gas engine fueled by natural gas or methane.Several researchers reported that use of H2-natural gas or H2-methane blend fuels cannot Furthermore, this multi-point ignition can make in-cylinder combustion process less sensitive to composition fluctuation of biogas, leading to stable onsite engine operations. Despite these advantages, the dual fuel engine is expensive together with high maintenance cost compared with a SI engine. It also suffers the possibility of frequent injector failures due to reduced cooling effect by a decrease in injected fuel and has issues on engine durability due to high in-cylinder pressure in power generation applications The other way to improve biogas engine performance is to add hydrogen (H2) in gaseous fuels. Hydrogen is well known for its excellent combustion characteristics

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and has been considered as a combustion enhancer for a gas engine fuelled by natural gas or methane.Several researchers reported that use of H2-natural gas or H2-methane blend fuels can not only extend lean operation limit of the engine but also decrease HC and carbon monoxide (CO) emissions while increasing NOx.However, most of these studies have focused on transportation applications.In addition, since all these previous works were carried out with biogas of relatively high methane concentration (more than 60%) and target power less than 20 kW, further research on a power generation engine with lower quality biogas less than 40% methane fraction needs to be done which is actually the case requiring an assistance of combustion promoter such as hydrogen. In this study, an experimental investigation on a naturally aspirated (NA), 8-L spark ignition engine fueled by biogas with various methane concentrations e which we called the N2 dilution test e was performed in terms of its thermal efficiency, combustion characteristics and emissions. The engine was

operated at a constant engine rotational speed of 1800 rpm under a 60 kW power output condition and simulated biogas was employed to realize a wide range of changes in heating value and gas composition. Then, as a way to achieve stable combustion for the lowest quality biogas, H2 addition tests were carried out in various excess air ratios. H2 fractions ranging from 5 to 30% were blended to the biogas and the effects of hydrogen addition on engine behavior were evaluated. Moreover, as one of the main results, a set of optimum operating conditions for maximum efficiency and minimum emission was suggested in terms of excess air ratio,spark ignition timing, hydrogen addition rate and so on.

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1.1 LITERATURE SURVEY

Experiments conducted on a Ricardo engine using simulated biogas at

different speeds, compression ratios and equivalence ratios revealed that the presence of CO2 can significantly lower NOx emissions However cylinder pressures, power and brake thermal efficiency were lower as compared to other gaseous fuels and the level of unburnt hydrocarbon emission was also higher. Also it has been reported that it is possible to significantly increase the Compression Ratio (CR) as an effective means of improving biogas fuelled engine’s performance when CO2 is present. They also found that an increase in compression ratio with biogas resulted in an increase in brake thermal efficiency, HC and NOx emissions. The influence of reduction in the concentration of CO2 in biogas was studied experimentally in a constant speed spark ignition engine at different equivalence ratios with a compression ratio of 13:1.It has been found that with a reduction in the CO2 level there was a significant improvement in the performance and reduction in emissions of hydrocarbons particularly with lean mixtures. Also it has been reported that a reduction in the CO2 level by 10% seemed to be sufficient for reducing HC levels and the NO levels were also not significant. Macari et al.developed a spark-ignited engine to operate on landfill gas without any derating of power from the standard natural gas rating. Crookes examined the performance and emissions from spark and compression ignition engines running on a variety of biofuels, including simulated biogas and commercial seed oil. Tests were performed at an engine speed of 2000 rpm w h a

relative air

fuel

ratio

ranging

from

rich

to

the

lean

misfire

limit

and

compression

ratios of 11:1 and 13:1 (before onset of detonation) in the SI mode. Raising the compression ratio is known to have the effect of increasing in cylinder temperature, NOx and HC. NOx was again found to reduce by dilution with either CO2 or N2, except at the highest fraction of N2.

Porpatham conducted experiments on a biogas fuelled SI engine with different hydrogen proportions (5%, 10% and 15%) on the energy basis. They found that addition of hydrogen in small quantities to biogas can improve power and thermal efficiency and reduce HC emissions. They also found that beyond 15% hydrogen the need to retard the ignition timing to control knock and there is a reduction in cycle by

cycle variations in combustion with lean mixtures. Rakopoulos and Michos examined the availability analysis on a spark ignition engine fuelled with biogas- hydrogen blends during the closed part of the engine cycle, with volumetric fractions of hydrogen up to 15%. It has been reported that the addition of increasing amounts of hydrogen in biogas promotes the degree of reversibility of the burning process.

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CHAPTER 2

VARIABLE COMPRESSION RATIO TEST

` Studies were conducted at IIT Madras on effect of compression ratio on performance and combustion of a biogas fuelled spark ignition engine.In the present study, a stationary, single cylinder agricultural diesel engine with a rated output of 4.4 kW at 1500 rpm was converted to operate as a gas engine using biogas as the fuel. Table 1 shows the specifications of the modified engine. The tests were conducted at 1500 rpm at two throttle openings namely 25% and 100% of maximum were evaluated at various equivalence ratios and number of compression ratios ranging from 9.3:1, 11:1, 13:1 and 15:1. The best spark timing was maintained for all the loads. Instruments were provided for obtaining the performance, emission, and combustion characteristics of the engine. Performance parameters like brake thermal efficiency, exhaust gas temperature, emissions of HC, CO and NO and combustion parameters like ignition delay, peak pressure, heat release rate and cycle by cycle variations in indicated mean effective pressure were studied and compared.

2.1 EXPERIMENTAL SETUP AND EXPERIMENTS:

Fig 2.1 shows the schematic diagram of the experimental setup used in this

work. A diesel engine was modified because it can withstand high compression ratios and knock. This engine was tested with different compression ratios between 9.3:1 and 15:1.In this work combustion chamber was modified to a hemispherical shape by forming a bowl on the piston. Different compression ratios were obtained by changing the volume of the piston bowl and clearance height simultaneously, while maintaining the calculated maximum squish velocity at 4 m/s.Table 2.1 shows the specifications of the engine

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Table 2.1 [2]

fig 2.1 [2]

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measured parameter was estimated based on Gaussian distribution method with confidence limits of ±2r (95.45% of measured data lie within the limits of ±2r of mean). For the analysis, 20 sets of readings have been taken at the same engine operating condition

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2.2RESULTS AND DISCUSSION

1. Variation of brake power with equivalence ratio at full throttle.

Fig.2.21 [2] The variation of brake power with equivalence ratio is shown in Fig 2.21 for

all the compression ratios at full throttle opening. We find that as the compression ratio goes up the peak power increases due to the improvement in the thermal efficiency. The leanest point shown in the figure corresponds to the condition just before the onset of misfire. The range of equivalence ratios covered the lean misfire limit on one side to the knock limit on the rich side. It may be noted that the lean limit is defined as the lowest equivalence ratio beyond which misfire occurs. On the rich side there is a drop in power output with increase in the compression ratio due to the need to retard the spark timing to prevent knock. This particularly evident with the higher compression ratio. It is seen that the lean misfire limit gets extended as the compression ratio increases. The lean limit indicated by misfire is an equivalence ratio of 0.64 with a compression ratio of 15:1 as against 0.77 with that of a

compression ratio of 9.3:1. Thus an increase in compression ratio extends the lean limit of operation because of higher gas Temperatures and lesser dilution by the residual exhaust gases. We find that the increase in the brake power becomes least significant as we move to higher compression ratios. Even the thermal efficiency of the ideal Otto cycle exhibits this trend. The peak poweroutput with a compression

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ratio of 15:1 is 4.8 kW and it is 10% higher than that with a compression ratio of 9.3:1 is 4.4 kW.

2.Variation of brake thermal efficiency with equivalence ratio at full throttle

Fig.2.22 [2]

From fig 2.22 we can find that the brake thermal efficiency at full throttle improves significantly when the compression ratio is raised as expected. With an increase in compression ratio from 9.3:1 to 15:1, the peak brake thermal efficiency increases from 23% to 26.8%. The difference between the thermal efficiencies at the two compression ratios between 9.3:1 and 11:1 is significant. Apart from the thermodynamic advantage the increase in the compression ratio also rises the combustion rate as shown later even with lean mixtures

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3. Variation of brake thermal efficiency with equivalence ratio at part throttle

Fig.2.23 [2]

Fig. 2.23 shows the variation of brake thermal efficiency at part throttle. A rise in the compression ratio increases the brake thermal efficiency slightly. However, at high equivalence ratios it is observed that the brake thermal efficiency is lower with the highest compression ratio of 15:1. With increase in compression ratio from 11:1 to 15:1, the peak brake thermal efficiency increases from 18% to 20%. Percent increase in thermal efficiency with 100% throttle is 16.5% and at 25% throttle the value is 11%

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4. Variation of hydrocarbon emission with equivalence ratio at full throttle

Fig.2.24 [2] At any equivalence ratio an increase in the compression ratio results in an increased

level of hydrocarbon emissions as seen in Fig 2.24 At full throttle condition increase in the compression ratio results in a rise in the mass of unburned fuel in the crevices. In addition a rise in the compression ratio reduces the post oxidation of HC in the exhaust since the exhaust temperature is lowered as seen Fig. 6 At an equivalence ratio of 0.94 the compression ratio increased from 9.3:1 to 15:1, HC level increases from 1184 ppm to 2000 ppm

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5.Variation of NOX emission with equivalence ratio at full throttle

Fig.2.25 [2]

Nitric oxide emission levels are seen in Fig 2.25 at full throttle condition. An

increase in

the

compression

ratio

resulted

in

increased

levels

of

NO

emission.

This

is

due to increase in peak gas temperature. At an equivalence ratio of 0.98 the compression ratio increased from 9.3:1 to 15:1, NO level increases from 2125 ppm to 2650 ppm, which is significant. The rise in the NO level is not so dominant till a compression ratio of 13:1

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CHAPTER 3

EFFECT OF HYDROGEN BLENDING

Biogas SI engines can have high compression ratio which will make the working cycle thermodynamically more efficient.Normally,diesel engine that use a high compression ratios are modified by adopting proper combustion chamber shape to run on biogas in the S I mode. It is very essential to maintain the correct air to fuel ratio and spark timing as the fuel has poor combustion qualities due to presence of CO 2.Other fuels like hydrogen and L P G that have better combustion characteristics in terms of flammability range, flame velocity and calorific value. We can improve performance of biogas engine by adding those in small quantities.

3.1 EXPERIMENTAL SETUP

In this study, an experimental investigation on a naturally aspirated, 8-L spark ignition engine fueled by biogas with various methane concentrations which we called the N2 dilution test was performed in terms of its thermal efficiency, combustion characteristics and emissions. The engine was operated at a constant engine rotational speed of 1800 rpm under a 60 kW power output condition and simulated biogas was employed to realize a wide range of changes in heating value and gas composition. Then, as a way to achieve stable combustion for the lowest quality biogas, H2 addition tests were carried out in various excess air ratios. H2 fractions ranging from 5 to 30% were blended to the biogas and the effects of hydrogen addition on engine behavior were evaluated. Moreover, as one of the main results, a set of optimum operating conditions for maximum efficiency and minimum emissions was suggested in terms of excess air ratio, spark ignition timing, hydrogen addition rate and so on. The biogas plant site expected to install the biogas engine after its development requires constant power output of 60 kW. To achieve this target power using biogas as low heating value as 20% of methane, a large engine displacement volume is necessary. Therefore, in this study, an 8 L, 6-cylinder natural gas engine was selected as a base engine. Table 3.21 shows the specifications of the

engine

and

Fig

3.1.shows

a

schematic

of

the

experimental

setup.

Several

modifications were made to the natural gas engine for use of biogas. Since intake manifold pressures were maintained below 1 bar for all the operating conditions, the turbocharger of the engine was removed so that fresh air was naturally aspirated.

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Fig 3.1 [1]

Table 3.1 [1]

Fig 3.1 shows a schematic of the experimental setup. From this biogas engine experiment of replacing CO2 to N2,it is reported that the effects of CO2 in biogas on engine performance and emissions are not of chemical origin but mainly compatible with the thermal influence of “exhaust gas recirculation (EGR)” to elevate heat capacity of the in-cylinder charge, and overall trends of N2 dilution results are not noticeably different from those described in CO2 dilution. Therefore, in this study, simulated biogas was supplied using natural gas and N2 instead of CO2. In real biogas, there exist impurities such as hydrogen silphide and ammonia (NH3), but they

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are not taken into consideration here since their amounts are negligible. Natural gas was provided from a compressed gas supply line distributed by Korea Gas Corporation To simulate as high as 80% of N2 concentration in biogas, liquid nitrogen was vaporized through a forced evaporator with an electric heater and delivered to a small surge tank placed upstream of the ETA for mixing with natural gas. Regulating and metering devices were fitted in each gas supply line to control the composition of the simulated biogas. Though most engine experiments were performed with these flow controlling devices, the ETA was also tested with low pressure (0.3 bar) natural gas from a separate gas line in order to prepare for an actual operation of the engine in the biogas plant site at Jeong-eup, South Korea. To add hydrogen to simulated biogas, a mass flow controller was mounted downstream of high pressure hydrogen bottles and hydrogen line was merged with the other gas supply lines in the surge tank for complete mixing This study was

designed with two major parts.ie N2 dilution and H2 addition experiments. In theN2 dilution test, simulated biogas fuels with varying methane concentrations (and thus varying N2 concentration accordingly) were provided to the engine so as to determine a set of operating conditions for Table 3.1Base engine specifications. The power output requirement of the engine was set to 60 kW and engine revolution speed was fixed at 1800 rpm to synchronize 60 Hz AC electricity. Spark timing sweeps were carried out for all the test conditions in order to determine the minimum advance for best torque (MBT). Coefficient of variation (COV) for the indicated mean effective pressure (IMEP) which represents the cyclic variation was calculated from incylinder pressure data to evaluate a degree of combustion stability at a given operating condition.

Excess air ratio of the N2 dilution test was maintained at stoichiometry to ensure stable combustion even for the biogaswith 20% heating value of natural gas (80% N2 dilution) while excess air ratios for the hydrogen addition tests were varied from stoichiometry to lean operation limit to decide optimum operating conditions for both maximum efficiency and minimum NOx emissions. The engine operating conditionswere determined by adjusting a throttle valve opening to achieve the torque requirements as well as a desired excess air ratio.the thermal efficiency and exaust emissions for biogas hydrogen mixture fuel is compared with those of biogas without

hydrogen addition.

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3.2 RESULTS AND DISCUSSION

1. N2 Dilution test

Fig 3.21 indicates the measured flow rates of natural gas and N2 at each fuel mixture condition in the N2 dilution test. While the level of natural gas to achieve a target power output was almost constant, the corresponding nitrogen flow rate was drastically increased to match the concentration within simulated biogas. Especially, 80% N2 dilution required about 2.5 times higher nitrogen flow rate than 60% case which can induce sudden changes in engine behavior. Moreover, if nitrogen mostly acts a role of EGR, the 80% N2 dilution was equivalent to about 32% of EGR and such a large amount of EGR considerably affects

Combustion phenomena in a spark ignition engine.

1. Flow rate of N2 and Natural gas with N2 dilution ratio

Fig 3.21 [1]

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3.Variations of NOX,THC emissions with N2 dilution

Fig 3.23 [1]

Fig 3.23 shows the variations of THC and NOx emissions as a function of N2 dilution ratio at the MBT timing. The increased inert gas induced a decrease in flame propagation velocity, advancing the MBT spark ignition timing. In general this advanced spark timing leads to a reduction in THC and a rise in NOx when using

high energy density fuels such as natural gas and gasoline. However, in biogas, the opposite tendencies were observed. i.e., NOx formation was significantly lessened while THC emissions increased. The variations in these emission measurement values can be explained by decreased combustion temperature and subsequently lowered oxidative capacity of the post-combustion, respectively, due to reduced flame propagation speed at high inert gas conditions.

2. H2 addition test

Hydrogen is well known for its outstanding combustion characteristics including a wide range of flammability, rapid flame speed, short quenching distance and high adiabatic flame temperature. Such characteristics enable stable combustion of low energy density fuels as well as low exhaust emissions even in the low load region when hydrogen is added in the fuels.Some of these features of hydrogen, however, can also induce disadvantages in actual H2 or H2 dominated fuel SI engine applications such as low thermal efficiency and abnormal combustion (backfire, and

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premature ignition) at a high load limit. Especially, there are several studies that the significant cooling loss due to the short quenching distance and high laminar burning speed makes thermal efficiency of a H2 engine lower than that of a hydrocarbon (e.g. methane) fueled engine at high load conditions.

In the H2 addition test, the biogas with the lowest heating value was selected as a main fuel. As mentioned in Fig. 3.23, a significant amount of biogas e 80% of which was N2 , was required to achieve the target power output in this fuel condition, which can inevitably lead to relatively high hydrogen supply rates under the experimental conditions of the 5 to 30% H2 concentrations. Therefore, it is important to compare the flow rate of each species in the blended fuel in order to understand the effects of hydrogen addition on engine behavior.

4.Flow rate of H2, NG and N2 with Hydrogen concentration

Fig 3.24 [1]

Fig.3.24 illustrates such measurement results in a stoichiometric engine operation under the power output of 60 kW at the MBT timing. As shown in the figure, hydrogen flow rate was significantly raised with an increase of H2 concentration in the total fuel flow rate. Under a fixed energy density condition for every fuel blend, it became more than a half of total combustible gases at the 20% H2 fraction condition and reached about 70% at the 30% H2 addition. On the contrary,

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volume combustion by H2 addition on thermal efficiency were limited to low H2 addition

6. Variations of THC and NOX emissions with hydrogen concentration

Fig 3.26[1]

Fig 3.26 shows THC and NOx emission variations with respect to hydrogen concentration at stoichiometry and the MBTspark ignition timing. As H2 fraction was raised, not only was the natural gas flow rate,the only source of carbon atoms are reduced as in Fig.11 but combustion efficiency was also promoted. Thus, THC emissions dropped drastically less than 330 ppm. Conversely, NOx emissions were elevated from 97 ppm to 258 ppm with an increase in hydrogen due to higher peak combustion temperature caused by faster burning speed.

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3.3 H2 ADDITION TEST-LEAN ENGINE OPERATION

The stoichiometric H2 addition tests indicated that an appropriate amount of hydrogen addition to biogas can be beneficial to thermal efficiency, combustion characteristics and THC emissions. However, it is also observed that even for the lowest energy density fuel with significant inert gas, too much blending of H2 e in this study, higher than 5% e is not always helpful, causing a considerable increase in NOx emissions as well as exacerbation of thermal efficiency. Especially, the NOx emission levels obtained in the entire operations cannot satisfy the government regulations (50 ppm, 15% O2), leading to the need for the lean engine operation.

7.Variations of thermal efficiency with excess air ratio

Fig3.31 [1]

The thermal

efficiency

results

for

varying

H2

concentrations

and

excess

air

ratios

are given in Fig3.32 .Up to 5% H2 concentration, only stoichiometric engine operation was possible due to the deterioration of in-cylinder combustion stability in a lean operating range. For higher hydrogen percentages, however, not only stable combustion was realized in lean mixture conditions, but also thermal efficiency was improved as excess air ratio went lean. This tendency can be attributed to several

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factors. More intake air was induced to the cylinder in a lean EAR, resulting in the reduced pumping loss which eventually raised engine efficiency. Combustion temperature was also lowered in a lean air/fuel mixture, which subsequently lessened cooling loss. Furthermore, the increase of the theoretical efficiency for lean mixtures was contributed by the enhancement of thermal efficiency. The figure also shows that the trend of higher thermalefficiency for lower hydrogen blending was maintained in each EAR condition, indicating that cooling energy loss was still an important factor even in lean engine operations. Unlike higher H2 concentration cases which demonstrate a monotonous increase of engine efficiency, the 10% hydrogen addition has a peak thermal efficiency of 32.3% at an EAR of 1.3 which is the maximum value for the entire conditions including both lean and stoichiometric operations. Moreover, the efficiency drop at an EAR of 1.4 suggested that combustion enhancement by adding 10% H2 was not high enough at the air/fuel ratio owing to the large amount of

N2 and excessive air. The lean operation limit was expanded up to 1.4 where the wide open throttle was reached, meaning that hydrogen acted as an effective combustion enhancer in the biogas engine. Yet, EAR window was rather narrow since the volume of biogas-hydrogenblend itself was substantial and occupied a considerable portion of the cylinder.

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8.Variations of THC emission with excess air ratio

Fig 3.33 [1]

Fig.3.33 describes THC emission results with respect to excess air ratio and

hydrogen fractions. Regardless of the air/ fuel ratio, more hydrogen addition produced less THC emissions due to promoted combustion characteristics. In terms of excess air ratio, THC emission levels were increased as the mixture went further lean after passing an EAR of 1.2 due to the elevated possibility of incomplete combustion. Especially, it can be seen that the degree of this THC rise was relatively lower for higher H2 addition. For example, THC emission with the 30% hydrogen blending was increased from 119 ppm at EAR of 1.2 to 294 ppm at EAR of 1.4 while being raised from 345 ppm to 952 ppm in the 10% H2. This result is not only because more hydrogen in the fuel allowed further improvement of lean combustion, but also because less natural gas which was the only source of hydrocarbon was introduced to the engine.

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CHAPTER 4

CONCLUSIONS

Various methods can be adopted to improve the emission and combustion characteristics of S I engine fuelled by biogas. From the results of various compression ratio test it can be concluded thatthe lean misfire limit of combustion of biogas under actual engine operating conditions gets considerably extended with increase in compression ratio. The lean limit indicated by misfire is an equivalence ratio of 0.64 with a compression ratio of 15:1 as against 0.77 with that of a compression ratio of 9.3:1.There is an improvement in thermal efficiency and brake power output with increase in compression ratio. The peak power output with a compression ratio of 15:1 is 4.8 kW and it is 10%higher than that with a compression ratio of 9.3:1 is 4.4 kW. With an increase in compression ratio from 9.3:1 to 15:1, the peak brake thermal efficiency increases from 23% to 26.8%.There is an increase in HC and NO level with rise in compression ratio. This is mainly due to the improvement in combustion by way of extension of the lean limit and increase in the combustion rate. At an equivalence ratio of 0.94 the compression ratio increased from 9.3:1 to 15:1, HC level increases from 1184 ppm to 2000 ppm and NO level increases from 2125 ppm to 2650 ppm, which is significant.

From hydrogen

blending

test

it

can

be

concluded

that

the

N2

dilution

tests

were conducted to assess the effects of inert gas concentration (and therefore fuel energy density) variations within biogas on performance and emission characteristics of a spark ignition engine operated at a fixed excess air ratio of 1. As a portion of inert gas in the fuel was raised, thermal efficiency was elevated due to a decrease in both pumping loss and cooling loss. As a way to improve combustion stability for the lowest energy density fuel (80% inert gas dilution case), H2 addition tests were carried out in a stoichiometric excess air ratio. Due to fast flame propagation speed of hydrogen, an increase of hydrogen concentration in the fuel blends enhanced combustion characteristics so that combustion duration was shortened, THC

emissions were reduced, and NOx emissions were elevated. In addition, the stoichiometric engine tests indicated that a moderate blending of hydrogen was beneficial to engine efficiency. However, when H2 fraction went over about 5%, cooling energy loss became dominant and efficiency dropped.

Finally, the engine was tested in various lean mixture conditions in order to satisfy NOx regulations as well as to maximize thermal efficiency. As excess air ratio

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went lean, a competition was observed between deterioration of combustion stability and a decrease in both pumping loss and cooling loss, resulting in peak engine efficiency at EAR of 1.3 in 10% H2 concentration. However, since NOx production in thiscondition still turned out to exceed the regulation value, two alternative approaches e slight retardation of spark timing or a little leaner operation e which can meet the emission standard without sacrificing considerable thermal efficiency were suggested as an optimum and practical operating point for use in an actual biogas site in the future.

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SEMINAR REPORT 2014

REFERENCE

[1].Cheolwoong Park,Seunghyun Park ,Yonggyu Lee, Changgi Kim, Sunyoup Lee ,Yasuo Moriyoshi.Perfomance and emission characteristics of a SI engine fueled by low calorific biogas blended with hydrogen.Hydrogen energy 2011. Environmental System Research Division, Engine Research Team, Korea Institute of Machinery and Materials (KIMM), Republic of Korea,Department of Environmental System Engineering,University of Science and Technology, Republic of Korea

[2].E. Porpatham, A. Ramesh , B. Nagalingam. Effect of compression ratio on the performance and combustion of a biogas fuelled spark ignition engine.fuel 2011. School of Mechanical and Building Sciences, VIT University, Vellore 632 014, India, IC Engines Laboratory, Indian Institute of Technology Madras, Chennai 600 036,

India

[3].E. Porpatham, A. Ramesh , B. Nagalingam .Investigation on the effect of concentration of methane in biogas when used as a fuel for a spark ignition engine.Fuel 2008. School of Mechanical and Building Sciences, VIT University, Vellore 632 014, India, IC Engines Laboratory, Indian Institute of Technology Madras, Chennai 600 036, India

[4].Jingdang Huang,and R J crooks.Assessment of simulated biogas as a fuel for spark ignition engine.department of mechanical engineering,Fujian agricultural university,fuzhoi,Fujian,350002,peopl’s republic of china.

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SEMINAR REPORT 2014

QUESTIONS AND ANSWERS

1. What you mean by compression ratio?

Ans. It is the ratio of maximum volume of cylinder to minimum volume of cylinder.

ie, r =(Vc+Vs)/Vc

2. What you mean by coefficient of variation?

Ans. It is the statistical term used to represent cyclic variations

C O V = Standard deviation /mean

3.What is the flame propagation velocity of hydrogen?

Ans. It is 275 cm/sec.whereas flame propagation velocity of biogas is 25cm/sec.

4.What happens to spark timing when hydrogen is blended? And what is the reason?

Ans. Spark can be retarded from 50 degree BTDC to 16 degree BTDC when hydrogen concentration is increased from 0% to 30%.the reason is due to enhanced combustion characteristics due to higher flame propagation velocity of hydrogen.

5.what is the maximum possible efficiency of biogas fuelled engine? Ans. The maximum possible efficiency is around 32%

53_SI-engine-fuelled-by-biogas.pdf

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