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An experimental investigation on the use of EGR in a supercharged natural gas SI engine
Fuel 89 (2010) 1721–1730
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An experimental investigation on the use of EGR in a supercharged natural gas SI engine Amr Ibrahim *, Saiful Bari Sustainable Energy Centre, School of Advanced Manufacturing and Mechanical Engineering, University of South Australia, SA 5095, Australia
a r t i c l e
i n f o
Article history: Received 4 May 2009 Received in revised form 3 July 2009 Accepted 7 July 2009 Available online 22 July 2009 Keywords: EGR SI engine Natural gas Inlet pressure NO
a b s t r a c t The use of lean burn technology in spark-ignition engines has been dominant; however, lean burn technique can not economically satisfy the increasingly restricted future emission standards. Consequently, alternative combustion techniques need to be investigated and developed. In this paper, the use of the stoichiometric air–fuel mixture with exhaust gas recirculation (EGR) technique in a spark-ignition natural gas engine was experimentally investigated. Engine performance and NO emissions were studied for both atmospheric and supercharged inlet conditions. It was found that the use of EGR has a significant effect on NO emissions. NO emissions decreased by about 50% when EGR dilution increased from zero with an inlet pressure of 101 kPa to close to the misfire limit with an inlet pressure of 113 kPa. In addition, the use of EGR effectively suppressed abnormal combustion which occurred at higher inlet pressure. The use of higher inlet pressure in the presence of EGR improved engine performance significantly. Engine brake power increased by about 20% and engine fuel consumption decreased by about 7% while NO emissions decreased by about 12% when 5% of EGR dilution was employed with an inlet pressure of 113 kPa compared to using undiluted stoichiometric inlet mixture with an inlet pressure of 101 kPa. Ó 2009 Elsevier Ltd. All rights reserved.
1. Introduction The World Health Organization has stated recently that urban air pollution is the most significant environment risk factor and that it is among the most important causes of death in Europe. According to the World Health Organization in 2002, exposure to particulate matter (PM) emissions was estimated to be responsible for about 100,000 premature deaths per year [1]. Automotive emissions greatly contribute to urban atmospheric pollution. Diesel engine emissions currently represent a main source of the urban PM emissions. Diesel exhaust is currently classified by the International Agency for Research on Cancer (IARC) as probably carcinogenic to humans. According to the US Environmental Protection Agency, EPA, motor vehicle exhaust was responsible for about 56% of carbon monoxide and 49% of nitrogen oxides emissions emitted into the atmosphere in USA [2]. The conventional fuels which have dominated the internal combustion engine chemical energy supply are basically petrol and diesel. These fuels were preferred to be used because of their availability, competitive price, and high energy density [3]. However, the use of both petrol and diesel has resulted in a serious harmful
* Corresponding author. Address: Sustainable Energy Centre, School of Advanced Manufacturing and Mechanical Engineering, University of South Australia, Mawson Lakes, SA 5095, Australia. Tel.: +618 830 25123; fax: +618 830 23380. E-mail address: [email protected] (A. Ibrahim). 0016-2361/$ - see front matter Ó 2009 Elsevier Ltd. All rights reserved. doi:10.1016/j.fuel.2009.07.005
impact on both environment and human health [4]. In addition, these fuels are derived from petroleum oil which is non renewable and it is expected to be totally consumed within the next few decades [4,5]. As the petroleum oil reserve declines, its cost is expected to increase dramatically in the future especially in the countries that have poor oil reserves and depend on imported oil for their energy supply. Recently, these environmental and economical concerns have motivated governments and research organizations to investigate different types of engine fuels that can be widely used as an alternative to the dominant conventional fuels of both petrol and diesel. Alternative fuels are expected to be friendlier to the environment and more sustainable than the conventional fuels. Natural gas might be considered as the cleanest fossil fuel. As natural gas consists basically of methane, the hydrogen to carbon mole ratio (H/C) in most natural gas compositions is close to 3.8, which is the highest hydrogen to carbon ratio compared to any other hydrocarbon fuel. The carbon mass percentage in natural gas is close to 75% compared to 86–88% for both petrol and diesel, which makes natural gas produces less carbon dioxide per unit of energy released [6]. Although natural gas is not renewable, natural gas is more sustainable than petroleum oil as the world reserves of natural gas is bigger than the oil. In addition, some of natural gas reserves exist in parts of the world that have poor oil reserves including Australia. Using natural gas as an affordable energy source in these countries can reduce their dependence on imported oil.
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Until a clean renewable energy source becomes widely available to the public with an affordable price, natural gas would remain as one of the main clean alternative fuel options that can be economically available in many parts of the world at least for some time in the future [7]. One of the natural gas spark-ignition engine combustion technologies, which might use up to almost 100% excess air [8], begun in the early 1980s and became known as ‘‘lean burn” combustion. Lean burn combustion resulted in relatively lower in-cylinder temperature, and consequently lower thermal stresses and lower knocking tendency. This permitted the use of turbocharging, relatively higher compression ratio, and optimum spark advance timing to achieve high engine efficiency with an acceptable durability. In addition, this technique satisfied the previous governmental emission standards without the need of exhaust gas after-treatment which resulted in relatively lower engine costs. As a result, most of the research and investment were directed towards the development of lean burn natural gas engines. Currently, increasingly stringent ambient air quality standards are forcing the production of engines that emit lower emissions; see Table 1 [9]. In order for the engine under the lean burn mode to produce lower nitrogen oxides (NOx) emissions, it has to operate with a leaner mixture. However, the use of excessive air dilution can deteriorate engine stability, increase hydrocarbon (HC) and CO emissions, and decrease engine efficiency. Another way to control NOx emissions is to retard the spark timing, which also leads to a decrease in engine efficiency and an increase in HC emissions [10]. Therefore, it seems that any efforts towards further decrease in NOx emission would lead to an increase in HC emission and a decrease in engine thermal efficiency [11]. It has become obvious that it would be difficult for the conventional gas engine operating on lean burn mode to meet the stringent future emission standards (see Table 1) especially for NOx emissions without using exhaust gas after-treatment [12]. The current technologies used for NOx emission after-treatment in lean burn engines such as the selective catalytic reduction (SCR) devices are expensive and add some complexity to the engine use. For example, the SCR technique consists of ammonia storage, feed, and injection system and a catalyst. In this system, the ammonia is injected in the exhaust gases upstream of the catalyst. In order for this system to operate properly, a certain exhaust gas temperature range must be maintained [13]. In addition, an oxidation catalyst would be necessary to reduce both the HC and CO emissions. On the other hand, the three-way catalyst (TWC) can be used to reduce NOx, HC and CO emissions in spark-ignition engines. The TWC is capable of reducing the three emissions at the same time and it is much less expensive than the SCR devices used in lean burn engines. However, in order for the TWC to operate efficiently, the engine must operate with almost a stoichiometric air–fuel mixture (i.e. without excess air). When the engine operates near the stoichiometric mixture, the in-cylinder temperature increases, and consequently, the thermal stresses and the knocking tendency increase. In order to reduce the in-cylinder temperature, an inlet charge dilution must be employed. One of the methods used to dilute the inlet charge is to recycle some of the exhaust gases back
Table 1 The European emission standards, g/kW h [9]. Year
Standard
CO
HC
NOx
PM
1996 2000 2005 2008 2013
Euro2 Euro3 Euro4 Euro5 Euro6a
4 2.1 1.5 1.5 1.5
1.1 0.66 0.46 0.46 0.13
7 5 3.5 2 0.4
0.15 0.1 0.02 0.02 0.01
a
Proposal only.
into the cylinder intake with the inlet mixture. This method is called exhaust gas recirculation (EGR). Adding EGR to the inlet mixture will reduce the oxygen partial pressure in the inlet mixture, and consequently the in-cylinder NOx production will decrease [14,15]. Also, the use of EGR with a stoichiometric air–fuel mixture can economically reduce engine emissions by allowing the use of a three-way catalyst. Several investigations which compared the effects of the use of lean combustion and EGR strategies on natural gas engine fuel consumption indicated that the use of the stoichiometric air–fuel mixture with EGR strategy resulted in higher fuel consumption compared to the lean burn strategy [11,16,17]. For instance, in 2007, Saanum and co-workers [11] concluded that a penalty in engine thermal efficiency must be accepted when EGR is used as an alternative to lean burn. However, these results can indicate that the use of a stoichiometric air–fuel mixture with EGR in natural gas spark-ignition engines has not been fully optimised yet. Further research is still needed in the current time and in the future to optimise the natural gas engine operation for the EGR strategy in order to achieve extremely low emissions accompanied with high engine power and efficiency. The aim of the research described in this paper is to experimentally investigate the effect of using EGR at both atmospheric and supercharged inlet conditions on natural gas engine performance and NO emissions. Although nitric oxide (NO) and nitrogen dioxide (NO2) are usually grouped together as NOx, nitric oxide is the predominant oxide of nitrogen produced inside an engine cylinder [18]. Consequently, only NO emission was experimentally measured and analyzed in this paper. 2. Experimental setup The experimental research was carried out using a single cylinder spark-ignition Ricardo engine whose specifications are shown in Table 2. The Ricardo engine was coupled to an electrical dynamometer in order to provide the load for the engine. The engine used to operate using petrol fuel before it was converted to operate on natural gas. Fig. 1 shows a schematic of the experimental setup. The recycled exhaust gas was taken from a hole located on the exhaust pipe with the help of a small suction pump. The hot exhaust gas was cooled by passing it through a water-cooled heat exchanger. Both a regulating valve and an orifice flow meter were installed down stream from the heat exchanger in order to regulate and measure the exhaust gas flow respectively. The increase of the percentage of EGR in the inlet mixture was done by increasing the amount of the exhaust gas flowing back to the engine intake. A supercharger was installed in order to provide the engine with inlet charge at high pressure. The supercharger was driven by an electric motor via a belt. Air, natural gas (from a pipe line supply), and cooled exhaust gas were mixed in the supercharger intake before they were delivered at high pressure to an intercooler. Table 2 Ricardo engine specifications. Item
Value
Serial No. No. of cylinders Bore (mm) Stroke (mm) Capacity (cc) Maximum speed (rpm) Inlet valve opens (° BTDC) Inlet valve closes (° ABDC) Exhaust valve opens (° BBDC) Exhaust valve closes (° ATDC)
120/73 1 76.2 111.125 507 3000 9 34 43 8
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Fig. 1. A schematic of experimental set-up.
The intercooler cooled down the air–fuel-exhaust gas mixture before it was delivered to the engine intake. A by pass valve was installed in order to control the inlet mixture pressure as shown in Fig. 1. A pressure gauge was used to measure the inlet pressure with an accuracy of ±2 kPa while K-type thermocouples were used to measure the EGR cooler inlet and outlet temperatures in addition to the intercooler inlet and outlet temperatures with an accuracy of ±1 °C. Engine speed was measured with an accuracy of ±10 rpm using a mechanical tachometer. All the research described in the paper was carried out at the wide open throttle condition while the inlet mixture pressure increased from 101 to 113 kPa by a supercharger. An Alcock Model 450 V viscous flow air meter was used to measure the air flow rate flowing to the engine with an accuracy of ± 0.09 m3/h while a Dwyer Model RMC gas flow meter was used to measure the natural gas flow rate flowing into the engine with an accuracy of ±0.045 m3/h. The fuel flow rate was controlled by a regulating valve so that the air–fuel ratio was fixed at the stoichiometric condition with an accuracy of ±4%. Also, a 6 mm square orifice meter was used to measure the flow rate of the exhaust gases recycled back to engine intake. The pressure difference across the orifice meter was measured with an accuracy of ±0.01 kPa using a U tube manometer. The percentage of exhaust gases recycled back to the engine intake (%EGR) was calculated as a percentage of the total inlet mass flow rate as follows:
%EGR ¼
_ EGR m 100 _ f þm _ EGR _ aþm m
ð1Þ
_ EGR ; m _ a; m _ f are the flow rates of EGR, air, and fuel respecwhere m tively in kg/s. NO emissions were measured using Beckman Industrial Model 951A NO emission analyzer. This emission analyzer uses the chemiluminescence technique, which depends on the emission of light, for measuring NO emissions. The setup for collecting the in-cylinder pressure crank-angle data is also shown in Fig. 1. The in-cylinder pressure crank-angle data was determined using water-cooled piezoelectric Kistler 7061B pressure sensor, top dead centre (TDC) position optical sensor, and magnetic pickup shaft encoder. A Kistler 5007 charge amplifier was used to convert the output electrical charge from the piezoelectric pressure sensor into DC voltage. The output signals from the charge amplifier, TDC position sensor, and shaft encoder were received by an analogue digital converter in order to convert the continuous signals into digital which were suitable to be handled by a laptop via a data acquisition card as indicated in Fig. 1. Lab View 7.1 computer software was used to monitor and save the outflow data from the charge amplifier, shaft encoder, and TDC position sensor.
3. Results and discussion In this section, the effect of varying the percentage of EGR in the inlet mixture at both atmospheric and supercharged inlet conditions on engine performance and NO emissions is discussed. Table 3 shows the engine operating conditions which were used during this study. The average inlet temperature of the air–gas–EGR
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Table 3 Engine operating conditions. Item
Value
Speed (rpm) Spark timing (° BTDC) Compression ratio Cooled EGR temperature (°C) Atmospheric inlet pressure (kPa) Supercharged inlet pressure (kPa) Air–fuel ratio
1500 47 8 65 101 113 Stoichiometric
mixture was about 27 and 35 °C for the atmospheric and supercharged inlet conditions, respectively.
3.1. Engine power Fig. 2 shows the effect of varying the percentage of EGR in the inlet mixture on engine brake power at both atmospheric and supercharged inlet pressures. When the inlet mixture was diluted with EGR at constant inlet pressure, the EGR replaced some of the inlet air, and consequently, the air flow rate decreased as shown in Fig. 3. Subsequently, the fuel flow rate was decreased in order to maintain a stoichiometric air fuel ratio as shown in Fig. 4. As a result, engine power generally decreased with the increase of percentage of EGR dilution for both atmospheric and supercharged inlet conditions. The increase of the inlet pressure from 101 to 113 kPa increased the inlet air and fuel flow rates as shown in Figs. 3 and 4, respectively, and consequently, engine power significantly increased. The use of undiluted air–fuel mixture (no EGR) at an inlet pressure of 113 kPa increased the in-cylinder temperature substantially and led to abnormal combustion occurrence as described in the next section. This led to some loss in engine power compared to the use of 5% of EGR at the same inlet pressure of 113 kPa. Fig. 2 indicates that increasing the inlet pressure from 101 to 113 kPa without EGR dilution increased engine power by about 16% as it increased from about 3.4 to 3.95 kW. However, the use of an EGR dilution of 5% with an inlet pressure of 113 kPa increased engine power by about 20% compared to the use of undiluted inlet mixture with an inlet pressure of 101 kPa. On the other hand, employing the maximum EGR dilution limit of about 12% with an inlet pressure of 113 kPa increased engine power by only 5% compared to the use of an inlet pressure of 101 kPa without EGR.
Fig. 3. Air flow rate variations with EGR dilution for atmospheric and supercharged inlet conditions.
Fig. 4. Fuel and EGR flow rate variations with EGR dilution for atmospheric and supercharged inlet conditions.
3.2. Misfire and abnormal combustion
Fig. 2. Brake power change with varying the percentage of EGR dilution for atmospheric and supercharged inlet conditions.
The higher increase in EGR dilution resulted in engine misfire which was identified by the measured in-cylinder pressure data as shown in Fig. 5. The misfire pressure cycles were similar to the pressure cycles measured during engine motoring process at the same conditions. Misfire started to occur with an EGR dilution above 10% for atmospheric inlet conditions and 12% for supercharged inlet conditions. The increase of EGR dilution misfire limit from about 10% to 12% with the increase of inlet pressure could be explained by the effect of inlet pressure on residual exhaust gas fraction. The residual exhaust gas fraction decreases with the increase of inlet pressure, which makes the engine more tolerant to EGR dilution. Also, the increase in the in-cylinder unburned mixture density with the increase of inlet pressure slightly increased the combustion rate as described in Section 3.6. The increase of the combustion rate with inlet pressure could also contribute in the increase of the engine tolerance to EGR dilution from 10% to 12%. Misfire occurrence resulted in high cycle to cycle variations and engine instability.
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Fig. 5. In-cylinder pressure data for supercharged inlet conditions and an EGR dilution of 13%.
On the other hand, when the engine was operated with a stoichiometric air–fuel mixture without EGR dilution at a supercharged inlet pressure of 113 kPa, some abnormal combustion pressure cycles were recorded. Abnormal combustion occurrence was identified by the excessively high maximum pressure of some combustion cycles compared to the maximum pressure of the normal combustion cycles for the same operating conditions as shown in Fig. 6. In order to further investigate these abnormal combustion cycles, one abnormal combustion cycle was isolated and used for combustion analysis. Fig. 7 shows the isolated abnormal combustion pressure cycle data, which shows the maximum pressure occurring almost at TDC. This could suggest that either the combustion was too fast or the combustion started too early. Both the heat release rate and the heat release were calculated using this pressure data and shown in Figs. 8 and 9, respectively. Both Figs. 8 and 9 illustrate that the combustion started at about 80° BTDC while the normal spark timing was set at 47° BTDC. This can explain that the air–fuel mixture was pre-ignited before the occurrence of a normal spark. When the engine was operated with a stoichiometric air–fuel mixture without EGR dilution at a supercharged inlet pressure of 113 kPa, the combustion pressure and
Fig. 6. In-cylinder pressure data for supercharged inlet conditions without EGR.
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Fig. 7. Pressure crank-angle data for an abnormal combustion cycle at supercharged inlet conditions without EGR.
Fig. 8. Heat release rate calculated for an abnormal combustion cycle at supercharged inlet conditions without EGR.
Fig. 9. Heat release calculated for an abnormal combustion cycle at supercharged inlet conditions without EGR.
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temperature excessively increased and caused some overheated spots inside the cylinder which ignited the air–fuel stoichiometric mixture earlier during some of the combustion cycles. This phenomenon is called pre-ignition or surface ignition. Surface ignition is the ignition of the air–fuel mixture by any means other than the spark discharge such as an overheated valve or spark plug. Following surface ignition, a turbulent flame develops at each surface ignition location and starts to propagate across the combustion chamber in an analogues manner to what occurs with normal spark-ignition [18]. It is worth mentioning that surface ignition differs from engine knock. Engine knock occurs when essentially a portion of the air– fuel mixture ahead of the propagating flame (the end gas) is auto-ignited due to high in-cylinder temperature and pressure. When knock occurs, much of the chemical energy of the end gas is released rapidly causing high local pressures and the propagation of pressure waves of substantial amplitude across the combustion chamber. Both surface ignition and knock can lead to engine instability and may cause engine damage. Engine knock can be identified using the recorded cylinder pressure data. Knock causes in-cylinder pressure to fluctuate with different amplitudes and frequencies according to knock severity during a combustion cycle. Light knock leads to pressure fluctuations with smaller amplitudes and frequencies compared to heavy knock. The surface ignition combustion pressure signal shown in Fig. 7 has no fluctuations, which indicates that the surface ignition was not accompanied with knock occurrence. Fig. 8 shows that there is a noticeable large portion of negative net heat release rate near the end of combustion which indicates that large amount of cylinder heat transfer occurred due to the excessive in-cylinder temperature caused by pre-ignition. In addition, the too advanced start of combustion for the pre-ignition combustion cycle resulted in a substantial increase in compression work, and consequently, the indicated power calculated for the pre-ignition combustion cycle was essentially lower compared to the indicated power calculated for normal combustion cycles. This increased cycle to cycle variations and led to engine instability. In order to show if the setting of the advanced spark timing of 47° BTDC was the reason behind the overheated hot spots inside the cylinder which caused surface ignition, the spark timing was retarded from 47° to 42°, 37°, and 35° BTDC, respectively for the undiluted stoichiometric air–fuel mixture and the in-cylinder pressure was measured for each spark timing condition. It was found that pre-ignition occurred at all the retarded spark timing conditions. In addition, both pre-ignition and misfire took place with a spark timing of 35° BTDC as shown in Fig. 10. On the other hand, when the stoichiometric air–fuel mixture was diluted with cooled EGR at a high inlet pressure of 113 kPa and a spark timing of 47° BTDC, the pre-ignition combustion cycles entirely disappeared with a percentage of EGR dilution in the inlet mixture of about 5%. This indicates the positive effect of EGR dilution on reducing in-cylinder combustion temperature and thermal stresses, which can eliminate abnormal combustion occurrence such as surface ignition. The presence of high specific heat gases such as water vapor in the EGR in addition to the increase of the total inlet mass with the increase of the percentage of EGR dilution in the inlet mixture help reduce the combustion temperature. Also, the increase of percentage of EGR dilution in the inlet mixture decreases oxygen concentration and slows down the combustion rate which leads to a decrease in both maximum cylinder pressure and temperature. 3.3. Engine stability One important measure of cyclic variability, which can be calculated from the measured cylinder pressure data, is the coefficient
Fig. 10. Pressure crank-angle data at supercharged inlet conditions without EGR and with a retarded spark timing of 35° BTDC.
of variation in indicated mean effective pressure, COV (%), which can be calculated as follows [18]:
COV ¼
rimep imep
100
ð2Þ
where imep (kPa) is the average indicated mean effective pressure calculated for a number of cycles, n, while rimep (kPa) is the standard deviation in indicated mean effective pressure. Both parameters can be calculated as follows [19]:
imep ¼
i¼n X
, imepðiÞ
n
ð3Þ
i¼1
rimep
vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi u i¼n u 1 X ¼t ðimepðiÞ imepÞ2 n 1 i¼1
ð4Þ
The COV defines the cycle to cycle variations in indicated work per cycle and it can be considered as a good indicator for engine stability determination. Heywood [18] demonstrated that vehicle drivability problems usually result when the coefficient of variation in indicated mean effective pressure, COV, exceeds about 10%. However, other studies showed that engine stability starts to deteriorate when COV increases above 5% [20]. Fig. 11 shows the effect of varying the percentage of EGR dilution in the inlet mixture on the coefficient of variation in indicated mean effective pressure, COV, for both atmospheric and supercharged inlet conditions. The dilution of the inlet mixture with EGR up to 10% at atmospheric inlet conditions resulted in a modest increase in COV. However, when the percentage of EGR dilution increased above 10%, COV started to increase rapidly due to misfire occurrence, which was identified for some of the combustion pressure cycles. On the other hand, when the inlet mixture was diluted with EGR at higher inlet pressure, the COV was improved compared to the undiluted inlet mixture condition as EGR dilution reduced thermal stresses and eliminated the pre-ignition occurrence. The COV decreased from 7% with no EGR dilution to about 1% with 10% EGR dilution. However, when the percentage of EGR increased above about 12%, the COV started to increase rapidly due to misfire occurrence. Fig. 11 shows that the increase in inlet pressure increased engine tolerance to EGR dilution from about 10% to 12%. The maximum tolerable EGR dilution limit, which is about 10% for the atmospheric inlet condition and about 12% for the
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Fig. 11. The variation of coefficient of variation in indicated mean effective pressure with the change of percentage of EGR dilution at both atmospheric and supercharged inlet conditions.
supercharged inlet condition and for the operating conditions shown in Table 3, might be considered low compared to the maximum tolerable limit for some modern engines which might tolerate EGR dilution limits of more than 15%. This is because the Ricardo engine had an essentially flat type combustion chamber with the spark plug located on its side. This type of combustion chamber does not support the creation of a significant flow motion such as swirl during combustion and consequently the combustion rate is not rapid enough to tolerate substantially high percentages of EGR dilution conditions. 3.4. In-cylinder pressure Both Figs. 12 and 13 show the effects of varying the percentage of EGR dilution in the inlet mixture on in-cylinder pressure for both atmospheric and supercharged inlet pressures, respectively, with a spark timing of 47° BTDC. The maximum cylinder pressure decreases while the crank angle at which the maximum cylinder pressure occurs is further shifted away from TDC with the increase of percentage of EGR. This
Fig. 12. The effect of varying the percentage of EGR dilution in the inlet mixture on in-cylinder pressure at atmospheric inlet conditions and spark timing of 47° BTDC.
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Fig. 13. The effect of varying the percentage of EGR dilution in the inlet mixture on in-cylinder pressure at inlet pressure of 113 kPa and spark timing of 47° BTDC.
can confirm that EGR dilution is capable of reducing in-cylinder stresses, and consequently, it can be used to suppress abnormal combustion occurrence such as surface ignition and knock. Fig. 14 shows the effect of the increase in inlet pressure on maximum cylinder pressure at different percentages of EGR dilution and spark timing of 47° BTDC. The increase in inlet pressure increased the maximum cylinder pressure while the dilution of the inlet mixture with EGR reduced it to lower levels. The increase of the inlet pressure of the undiluted air–fuel mixture from 101 to 113 kPa increased the maximum cylinder pressure by about 16.6% as it increased from 44.6 to 52 bar. However, when the inlet mixture was diluted with EGR with a percentage dilution of 12%, the maximum cylinder pressure decreased by about 31.5% as it decreased from 52 to 35.6 bar as shown in Fig. 14. 3.5. Heat release rate and heat release In order to show the effect of the increase of percentage of EGR dilution in the inlet mixture on both the net heat release rate and net heat release at both atmospheric and supercharged inlet conditions, the pressure signals shown in Figs. 12 and 13 were analyzed
Fig. 14. The effect of the increase of inlet pressure on maximum cylinder pressure at different percentages of EGR dilution and spark timing of 47° BTDC.
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for combustion analysis. Both Figs. 15 and 16 show the effect of varying the percentage of EGR dilution in the inlet mixture on the net heat release rate for both atmospheric and supercharged inlet conditions, respectively. It can be noted from both Figs. 15 and 16 that the maximum heat release rate decreases with the increase of percentage of EGR dilution. Also, the crank angle at which the maximum heat release rate occurs is shifted away further from the TDC with the increase of percentage of EGR dilution. The increase of EGR dilution in the inlet mixture decreases the in-cylinder oxygen concentration, and consequently, it reduces the heat release rate. Both the heat release rate in Figs. 15 and 16 were integrated to calculate the net heat release. Both Figs. 17 and 18 show the effect of the change of percentage of EGR dilution in the inlet mixture on the heat release for both atmospheric and supercharged inlet conditions, respectively. The heat release crank-angle relationship has the characteristic S-shape. The heat release starts with almost zero at the time of spark which was fixed at 313° (or 47° BTDC), and then it starts to increase 10–20° after the spark timing. This angle is usually called flame development angle or sometimes ignition delay. Fol-
Fig. 15. The effect of varying EGR dilution in the inlet mixture on the net heat release rate at atmospheric inlet condition and spark timing of 47° BTDC.
Fig. 17. The effect of varying the percentage of EGR dilution in the inlet mixture on the net heat release at atmospheric inlet condition and spark timing of 47° BTDC.
Fig. 18. The effect of varying the percentage of EGR dilution in the inlet mixture on the net heat release at an inlet pressure of 113 kPa and spark timing of 47° BTDC.
lowing the flame development duration, the heat release noticeably increases with crank angle until it reaches its maximum value where essentially almost all of the fuel chemical energy has been released at basically the end of combustion. Both Figs. 17 and 18 indicate that the maximum heat release, which essentially identifies the end of combustion, occurs later as the percentage of EGR dilution increases. 3.6. Combustion duration
Fig. 16. The effect of varying EGR dilution in the inlet mixture on the net heat release rate at an inlet pressure of 113 kPa and spark timing of 47° BTDC.
Fig. 19 shows the effect of varying the percentage of EGR dilution in the inlet mixture on the total combustion duration for both atmospheric and supercharged inlet conditions. The total combustion duration was calculated as the crank angle interval from the spark timing to essentially the end of combustion where the heat release reaches its maximum value. Fig. 19 indicates that the increase of percentage of EGR dilution in the inlet mixture increases the total combustion duration for both atmospheric and supercharged inlet conditions. The increase of EGR dilution decreased the oxygen concentration during combustion which slowed down the combustion rate and increased the total combustion duration.
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Fig. 19. The variation of the total combustion duration with the percentage of EGR dilution at atmospheric and supercharged inlet conditions.
Fig. 21. The effect of the increase of percentage of EGR dilution in the inlet mixture on NO emissions at two different inlet pressures.
However, the combustion duration slightly decreased with the increase of inlet pressure. The decrease of combustion duration with the increase of inlet pressure could be due to the increase of the unburned mixture density during combustion which slightly increased the burning rate.
compared to the use of diluted inlet mixture at the same inlet pressure of 113 kPa. The dilution of the inlet mixture with EGR dilution of 5% reduced in-cylinder thermal and mechanical stresses which prevented surface ignition occurrence and decreased cylinder heat transfer, and consequently, this resulted in a significant decrease in engine fuel consumption. Fig. 20 shows that the increase of EGR dilution from 5% to 10% has less significant effect on engine fuel consumption, however, the engine fuel consumption started to increase rapidly as the EGR misfire limit approached. Increasing the inlet pressure increased engine power as demonstrated in Fig. 2 and significantly decreased fuel consumption as shown in Fig. 20. Engine fuel consumption decreased by about 7% when 5% of EGR dilution was employed at an inlet pressure of 113 kPa compared to the use of undiluted inlet mixture with an inlet pressure of 101 kPa. However, the use of the maximum EGR dilution limit of 12% decreased engine fuel consumption by only 1% compared to using an inlet pressure of 101 kPa without EGR. The supercharger work was neglected during brake power and bsfc calculations as the mechanical supercharger used with the Ricardo single cylinder engine was assumed to simulate the use of a turbocharger in a multi-cylinder engine.
3.7. Engine fuel consumption Fig. 20 shows the effect of varying the percentage of EGR dilution in the inlet mixture on brake specific fuel consumption (bsfc) for both atmospheric and supercharged inlet conditions. EGR dilution had little effect on fuel consumption when it was added to the inlet mixture at atmospheric inlet pressure with lower percentage of dilution of up to about 8%. However, when the EGR dilution was increased above 8%, engine fuel consumption started to increase rapidly as the EGR misfire limit approached. On the other hand, the use of an undiluted stoichiometric air– fuel mixture at an inlet pressure of 113 kPa increased the in-cylinder pressure and temperature significantly which led to surface ignition occurrence in addition to a substantial increase in heat transfer, which resulted in an increase in engine fuel consumption
3.8. NO emissions Fig. 21 shows the effect of the increase of percentage of EGR in the inlet mixture on NO emissions at both atmospheric and supercharged inlet conditions. Fig. 21 indicates that the use of EGR dilution has a significant capability on reducing NO emissions. The increase of EGR dilution in the inlet mixture decreased both the combustion temperature and oxygen concentration which led to a significant decrease in NO emissions. NO emissions decreased by about 50% when the EGR dilution was employed at a rate close to the misfire limit with an inlet pressure of 113 kPa compared to the use of undiluted inlet mixture with an inlet pressure of 101 kPa. On the other hand, using 5% of EGR dilution with an inlet pressure of 113 kPa reduced NO emissions by about 12% compared to using undiluted inlet mixture with an inlet pressure of 101 kPa. 4. Conclusions Fig. 20. The variation of brake specific fuel consumption with the percentage of EGR dilution at atmospheric and supercharged inlet conditions.
In this paper, the use of EGR strategy was experimentally investigated in a natural gas spark-ignition engine. Engine performance
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and NO emissions were studied for both atmospheric and supercharged inlet conditions and the following conclusions have been obtained: 1. The increase of the percentage of EGR dilution in the inlet mixture decreases the oxygen concentration, and consequently, it decreases the combustion rate significantly. For instance, the increase of EGR dilution from 0% to 10% at atmospheric inlet conditions increased the total combustion duration from about 74° to 95°. 2. The decrease in combustion rate with the increase of EGR dilution in the inlet mixture helps reduce both the maximum cylinder pressure and temperature, and consequently, prevents abnormal combustion occurrence such as surface ignition, which occurred at higher inlet pressure. 3. The increase of EGR dilution in the inlet mixture decreases both the maximum cylinder temperature and oxygen concentration which leads to a significant reduction in NO emissions. For instance, NO emissions decreased by about 50% when EGR dilution increased from zero with an inlet pressure of 101 kPa to close to the misfire limit with an inlet pressure of 113 kPa. 4. The engine tolerance to EGR increases with the increase of inlet pressure. The maximum tolerable EGR dilution limit increased from about 10% to 12% when the inlet pressure increased from 101 to 113 kPa. 5. The use of an undiluted stoichiometric air–fuel mixture at higher inlet pressure deteriorates engine stability. However, engine stability was improved when the inlet mixture was sufficiently diluted with EGR. 6. Engine performance can be significantly improved when the EGR dilution strategy is employed at higher inlet pressure. Engine brake power increased by about 20% and engine fuel consumption decreased by about 7% while NO emissions decreased by about 12% when 5% of EGR dilution was employed with an inlet pressure of 113 kPa compared to using undiluted stoichiometric inlet mixture with an inlet pressure of 101 kPa. References [1] Baldassarri LT, Battistelli CL, Conti L, Crebelli R, De Berardis B, Iamiceli AL, et al. Compressed natural gas and comparison with liquid fuels. Sci Total Environ 2006;355(1–3):64–77.
[2] US Environmental Protection Agency, EPA, web information:; 2008 [accessed in August 2008]. [3] Demirbas A. Biodiesel production from vegetable oils via catalytic and noncatalytic supercritical methanol transesterification methods. Progr Energy Combust Sci 2005;31:466–87. [4] Agarwal AK. Biofuels (alcohols and biodiesel) applications as fuels for internal combustion engines. Progr Energy Combust Sci 2007;33:233–71. [5] Akansu SO, Dulger Z, Kahraman N, Veziroglu TN. Internal combustion engines fuelled by natural gas–hydrogen mixtures. Int J Hydrogen Energy 2004;29:1527–39. [6] Maclean HL, Lave LB. Evaluating automobile fuel/propulsion system technologies. Progr Energy Combust Sci 2003;29:1–69. [7] International Association for Natural Gas Vehicles, ; 2004 [accessed in August 2008]. [9] Rabl A. Environmental benefits of natural gas for buses. J Transport Res, Part D 2002;7:391–405. [10] Bhargava S, Clark NN, Hildebrand MW. Exhaust Gas Recirculation in a LeanBurn Natural Gas Engine. SAE paper no. 981395, 1998. [11] Saanum I, Bysveen M, Tunestal P, Johansson B.Lean Burn versus Stoichiometric Operation with EGR and 3-Way Catalyst of an Engine Fueled with Natural Gas and Hydrogen Enriched Natural Gas. SAE paper no. 2007-01-0015, 2007. [12] Nellen C, Boulouchos K. Natural Gas Engines for Cogeneration: Highest Efficiency and Near-Zero-Emissions through Turbocharging, EGR and 3-Way Catalytic Converter. SAE paper no. 2000-01-2825, 2000. [13] Control Technologies for Hazardous Air Pollutants. Handbook, EPA 625/6-91014, 1991. [14] Ibrahim A, Bari S. Optimization of a natural gas SI engine employing EGR strategy using a two zone combustion model. J Fuel 2008;87:1824–34. [15] Ibrahim A, Bari S, Bruno F. A Study on EGR Utilization in Natural Gas SI Engines Using a Two-zone Combustion Model. SAE paper no. 2007-01-2041, 2007. [16] Einewall P, Tunestal P, Johansson B. Lean Burn Natural Gas Operation vs. Stoichiometric Operation with EGR and a Three Way Catalyst. SAE paper no. 2005-01-0250, 2005. [17] Reppert T, Chiu J. Heavy Duty Waste Hauler with Chemically Correct Natural Gas Engine Diluted with EGR and Using a Three-Way Catalyst, National Renewable Energy Laboratory Report, report no. NREL/SR-540-38222, USA, 2005. [18] Heywood JB. Internal combustion engine fundamentals. USA: McGraw-Hill; 1988. ISBN:0-07-028637-X. [19] Czarnigowski J. A Simple Method of Analysis and Modeling of Cycle-to-Cycle Variation of Engine Work – The Example of Indicated Pressure. SAE paper no. 2007-01-2079, 2007. [20] Allenby S, Chang WC, Megaritis A, Wyszynski ML. Hydrogen enrichment: a way to maintain combustion stability in a natural gas fuelled engine with exhaust gas recirculation, the potential of fuel forming. Automob Eng 2001;215:405–18.
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An experimental investigation on the use of EGR in a supercharged natural gas SI engine Amr Ibrahim *, Saiful Bari Sustainable Energy Centre, School of Advanced Manufacturing and Mechanical Engineering, University of South Australia, SA 5095, Australia
a r t i c l e
i n f o
Article history: Received 4 May 2009 Received in revised form 3 July 2009 Accepted 7 July 2009 Available online 22 July 2009 Keywords: EGR SI engine Natural gas Inlet pressure NO
a b s t r a c t The use of lean burn technology in spark-ignition engines has been dominant; however, lean burn technique can not economically satisfy the increasingly restricted future emission standards. Consequently, alternative combustion techniques need to be investigated and developed. In this paper, the use of the stoichiometric air–fuel mixture with exhaust gas recirculation (EGR) technique in a spark-ignition natural gas engine was experimentally investigated. Engine performance and NO emissions were studied for both atmospheric and supercharged inlet conditions. It was found that the use of EGR has a significant effect on NO emissions. NO emissions decreased by about 50% when EGR dilution increased from zero with an inlet pressure of 101 kPa to close to the misfire limit with an inlet pressure of 113 kPa. In addition, the use of EGR effectively suppressed abnormal combustion which occurred at higher inlet pressure. The use of higher inlet pressure in the presence of EGR improved engine performance significantly. Engine brake power increased by about 20% and engine fuel consumption decreased by about 7% while NO emissions decreased by about 12% when 5% of EGR dilution was employed with an inlet pressure of 113 kPa compared to using undiluted stoichiometric inlet mixture with an inlet pressure of 101 kPa. Ó 2009 Elsevier Ltd. All rights reserved.
1. Introduction The World Health Organization has stated recently that urban air pollution is the most significant environment risk factor and that it is among the most important causes of death in Europe. According to the World Health Organization in 2002, exposure to particulate matter (PM) emissions was estimated to be responsible for about 100,000 premature deaths per year [1]. Automotive emissions greatly contribute to urban atmospheric pollution. Diesel engine emissions currently represent a main source of the urban PM emissions. Diesel exhaust is currently classified by the International Agency for Research on Cancer (IARC) as probably carcinogenic to humans. According to the US Environmental Protection Agency, EPA, motor vehicle exhaust was responsible for about 56% of carbon monoxide and 49% of nitrogen oxides emissions emitted into the atmosphere in USA [2]. The conventional fuels which have dominated the internal combustion engine chemical energy supply are basically petrol and diesel. These fuels were preferred to be used because of their availability, competitive price, and high energy density [3]. However, the use of both petrol and diesel has resulted in a serious harmful
* Corresponding author. Address: Sustainable Energy Centre, School of Advanced Manufacturing and Mechanical Engineering, University of South Australia, Mawson Lakes, SA 5095, Australia. Tel.: +618 830 25123; fax: +618 830 23380. E-mail address: [email protected] (A. Ibrahim). 0016-2361/$ - see front matter Ó 2009 Elsevier Ltd. All rights reserved. doi:10.1016/j.fuel.2009.07.005
impact on both environment and human health [4]. In addition, these fuels are derived from petroleum oil which is non renewable and it is expected to be totally consumed within the next few decades [4,5]. As the petroleum oil reserve declines, its cost is expected to increase dramatically in the future especially in the countries that have poor oil reserves and depend on imported oil for their energy supply. Recently, these environmental and economical concerns have motivated governments and research organizations to investigate different types of engine fuels that can be widely used as an alternative to the dominant conventional fuels of both petrol and diesel. Alternative fuels are expected to be friendlier to the environment and more sustainable than the conventional fuels. Natural gas might be considered as the cleanest fossil fuel. As natural gas consists basically of methane, the hydrogen to carbon mole ratio (H/C) in most natural gas compositions is close to 3.8, which is the highest hydrogen to carbon ratio compared to any other hydrocarbon fuel. The carbon mass percentage in natural gas is close to 75% compared to 86–88% for both petrol and diesel, which makes natural gas produces less carbon dioxide per unit of energy released [6]. Although natural gas is not renewable, natural gas is more sustainable than petroleum oil as the world reserves of natural gas is bigger than the oil. In addition, some of natural gas reserves exist in parts of the world that have poor oil reserves including Australia. Using natural gas as an affordable energy source in these countries can reduce their dependence on imported oil.
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Until a clean renewable energy source becomes widely available to the public with an affordable price, natural gas would remain as one of the main clean alternative fuel options that can be economically available in many parts of the world at least for some time in the future [7]. One of the natural gas spark-ignition engine combustion technologies, which might use up to almost 100% excess air [8], begun in the early 1980s and became known as ‘‘lean burn” combustion. Lean burn combustion resulted in relatively lower in-cylinder temperature, and consequently lower thermal stresses and lower knocking tendency. This permitted the use of turbocharging, relatively higher compression ratio, and optimum spark advance timing to achieve high engine efficiency with an acceptable durability. In addition, this technique satisfied the previous governmental emission standards without the need of exhaust gas after-treatment which resulted in relatively lower engine costs. As a result, most of the research and investment were directed towards the development of lean burn natural gas engines. Currently, increasingly stringent ambient air quality standards are forcing the production of engines that emit lower emissions; see Table 1 [9]. In order for the engine under the lean burn mode to produce lower nitrogen oxides (NOx) emissions, it has to operate with a leaner mixture. However, the use of excessive air dilution can deteriorate engine stability, increase hydrocarbon (HC) and CO emissions, and decrease engine efficiency. Another way to control NOx emissions is to retard the spark timing, which also leads to a decrease in engine efficiency and an increase in HC emissions [10]. Therefore, it seems that any efforts towards further decrease in NOx emission would lead to an increase in HC emission and a decrease in engine thermal efficiency [11]. It has become obvious that it would be difficult for the conventional gas engine operating on lean burn mode to meet the stringent future emission standards (see Table 1) especially for NOx emissions without using exhaust gas after-treatment [12]. The current technologies used for NOx emission after-treatment in lean burn engines such as the selective catalytic reduction (SCR) devices are expensive and add some complexity to the engine use. For example, the SCR technique consists of ammonia storage, feed, and injection system and a catalyst. In this system, the ammonia is injected in the exhaust gases upstream of the catalyst. In order for this system to operate properly, a certain exhaust gas temperature range must be maintained [13]. In addition, an oxidation catalyst would be necessary to reduce both the HC and CO emissions. On the other hand, the three-way catalyst (TWC) can be used to reduce NOx, HC and CO emissions in spark-ignition engines. The TWC is capable of reducing the three emissions at the same time and it is much less expensive than the SCR devices used in lean burn engines. However, in order for the TWC to operate efficiently, the engine must operate with almost a stoichiometric air–fuel mixture (i.e. without excess air). When the engine operates near the stoichiometric mixture, the in-cylinder temperature increases, and consequently, the thermal stresses and the knocking tendency increase. In order to reduce the in-cylinder temperature, an inlet charge dilution must be employed. One of the methods used to dilute the inlet charge is to recycle some of the exhaust gases back
Table 1 The European emission standards, g/kW h [9]. Year
Standard
CO
HC
NOx
PM
1996 2000 2005 2008 2013
Euro2 Euro3 Euro4 Euro5 Euro6a
4 2.1 1.5 1.5 1.5
1.1 0.66 0.46 0.46 0.13
7 5 3.5 2 0.4
0.15 0.1 0.02 0.02 0.01
a
Proposal only.
into the cylinder intake with the inlet mixture. This method is called exhaust gas recirculation (EGR). Adding EGR to the inlet mixture will reduce the oxygen partial pressure in the inlet mixture, and consequently the in-cylinder NOx production will decrease [14,15]. Also, the use of EGR with a stoichiometric air–fuel mixture can economically reduce engine emissions by allowing the use of a three-way catalyst. Several investigations which compared the effects of the use of lean combustion and EGR strategies on natural gas engine fuel consumption indicated that the use of the stoichiometric air–fuel mixture with EGR strategy resulted in higher fuel consumption compared to the lean burn strategy [11,16,17]. For instance, in 2007, Saanum and co-workers [11] concluded that a penalty in engine thermal efficiency must be accepted when EGR is used as an alternative to lean burn. However, these results can indicate that the use of a stoichiometric air–fuel mixture with EGR in natural gas spark-ignition engines has not been fully optimised yet. Further research is still needed in the current time and in the future to optimise the natural gas engine operation for the EGR strategy in order to achieve extremely low emissions accompanied with high engine power and efficiency. The aim of the research described in this paper is to experimentally investigate the effect of using EGR at both atmospheric and supercharged inlet conditions on natural gas engine performance and NO emissions. Although nitric oxide (NO) and nitrogen dioxide (NO2) are usually grouped together as NOx, nitric oxide is the predominant oxide of nitrogen produced inside an engine cylinder [18]. Consequently, only NO emission was experimentally measured and analyzed in this paper. 2. Experimental setup The experimental research was carried out using a single cylinder spark-ignition Ricardo engine whose specifications are shown in Table 2. The Ricardo engine was coupled to an electrical dynamometer in order to provide the load for the engine. The engine used to operate using petrol fuel before it was converted to operate on natural gas. Fig. 1 shows a schematic of the experimental setup. The recycled exhaust gas was taken from a hole located on the exhaust pipe with the help of a small suction pump. The hot exhaust gas was cooled by passing it through a water-cooled heat exchanger. Both a regulating valve and an orifice flow meter were installed down stream from the heat exchanger in order to regulate and measure the exhaust gas flow respectively. The increase of the percentage of EGR in the inlet mixture was done by increasing the amount of the exhaust gas flowing back to the engine intake. A supercharger was installed in order to provide the engine with inlet charge at high pressure. The supercharger was driven by an electric motor via a belt. Air, natural gas (from a pipe line supply), and cooled exhaust gas were mixed in the supercharger intake before they were delivered at high pressure to an intercooler. Table 2 Ricardo engine specifications. Item
Value
Serial No. No. of cylinders Bore (mm) Stroke (mm) Capacity (cc) Maximum speed (rpm) Inlet valve opens (° BTDC) Inlet valve closes (° ABDC) Exhaust valve opens (° BBDC) Exhaust valve closes (° ATDC)
120/73 1 76.2 111.125 507 3000 9 34 43 8
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Fig. 1. A schematic of experimental set-up.
The intercooler cooled down the air–fuel-exhaust gas mixture before it was delivered to the engine intake. A by pass valve was installed in order to control the inlet mixture pressure as shown in Fig. 1. A pressure gauge was used to measure the inlet pressure with an accuracy of ±2 kPa while K-type thermocouples were used to measure the EGR cooler inlet and outlet temperatures in addition to the intercooler inlet and outlet temperatures with an accuracy of ±1 °C. Engine speed was measured with an accuracy of ±10 rpm using a mechanical tachometer. All the research described in the paper was carried out at the wide open throttle condition while the inlet mixture pressure increased from 101 to 113 kPa by a supercharger. An Alcock Model 450 V viscous flow air meter was used to measure the air flow rate flowing to the engine with an accuracy of ± 0.09 m3/h while a Dwyer Model RMC gas flow meter was used to measure the natural gas flow rate flowing into the engine with an accuracy of ±0.045 m3/h. The fuel flow rate was controlled by a regulating valve so that the air–fuel ratio was fixed at the stoichiometric condition with an accuracy of ±4%. Also, a 6 mm square orifice meter was used to measure the flow rate of the exhaust gases recycled back to engine intake. The pressure difference across the orifice meter was measured with an accuracy of ±0.01 kPa using a U tube manometer. The percentage of exhaust gases recycled back to the engine intake (%EGR) was calculated as a percentage of the total inlet mass flow rate as follows:
%EGR ¼
_ EGR m 100 _ f þm _ EGR _ aþm m
ð1Þ
_ EGR ; m _ a; m _ f are the flow rates of EGR, air, and fuel respecwhere m tively in kg/s. NO emissions were measured using Beckman Industrial Model 951A NO emission analyzer. This emission analyzer uses the chemiluminescence technique, which depends on the emission of light, for measuring NO emissions. The setup for collecting the in-cylinder pressure crank-angle data is also shown in Fig. 1. The in-cylinder pressure crank-angle data was determined using water-cooled piezoelectric Kistler 7061B pressure sensor, top dead centre (TDC) position optical sensor, and magnetic pickup shaft encoder. A Kistler 5007 charge amplifier was used to convert the output electrical charge from the piezoelectric pressure sensor into DC voltage. The output signals from the charge amplifier, TDC position sensor, and shaft encoder were received by an analogue digital converter in order to convert the continuous signals into digital which were suitable to be handled by a laptop via a data acquisition card as indicated in Fig. 1. Lab View 7.1 computer software was used to monitor and save the outflow data from the charge amplifier, shaft encoder, and TDC position sensor.
3. Results and discussion In this section, the effect of varying the percentage of EGR in the inlet mixture at both atmospheric and supercharged inlet conditions on engine performance and NO emissions is discussed. Table 3 shows the engine operating conditions which were used during this study. The average inlet temperature of the air–gas–EGR
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Table 3 Engine operating conditions. Item
Value
Speed (rpm) Spark timing (° BTDC) Compression ratio Cooled EGR temperature (°C) Atmospheric inlet pressure (kPa) Supercharged inlet pressure (kPa) Air–fuel ratio
1500 47 8 65 101 113 Stoichiometric
mixture was about 27 and 35 °C for the atmospheric and supercharged inlet conditions, respectively.
3.1. Engine power Fig. 2 shows the effect of varying the percentage of EGR in the inlet mixture on engine brake power at both atmospheric and supercharged inlet pressures. When the inlet mixture was diluted with EGR at constant inlet pressure, the EGR replaced some of the inlet air, and consequently, the air flow rate decreased as shown in Fig. 3. Subsequently, the fuel flow rate was decreased in order to maintain a stoichiometric air fuel ratio as shown in Fig. 4. As a result, engine power generally decreased with the increase of percentage of EGR dilution for both atmospheric and supercharged inlet conditions. The increase of the inlet pressure from 101 to 113 kPa increased the inlet air and fuel flow rates as shown in Figs. 3 and 4, respectively, and consequently, engine power significantly increased. The use of undiluted air–fuel mixture (no EGR) at an inlet pressure of 113 kPa increased the in-cylinder temperature substantially and led to abnormal combustion occurrence as described in the next section. This led to some loss in engine power compared to the use of 5% of EGR at the same inlet pressure of 113 kPa. Fig. 2 indicates that increasing the inlet pressure from 101 to 113 kPa without EGR dilution increased engine power by about 16% as it increased from about 3.4 to 3.95 kW. However, the use of an EGR dilution of 5% with an inlet pressure of 113 kPa increased engine power by about 20% compared to the use of undiluted inlet mixture with an inlet pressure of 101 kPa. On the other hand, employing the maximum EGR dilution limit of about 12% with an inlet pressure of 113 kPa increased engine power by only 5% compared to the use of an inlet pressure of 101 kPa without EGR.
Fig. 3. Air flow rate variations with EGR dilution for atmospheric and supercharged inlet conditions.
Fig. 4. Fuel and EGR flow rate variations with EGR dilution for atmospheric and supercharged inlet conditions.
3.2. Misfire and abnormal combustion
Fig. 2. Brake power change with varying the percentage of EGR dilution for atmospheric and supercharged inlet conditions.
The higher increase in EGR dilution resulted in engine misfire which was identified by the measured in-cylinder pressure data as shown in Fig. 5. The misfire pressure cycles were similar to the pressure cycles measured during engine motoring process at the same conditions. Misfire started to occur with an EGR dilution above 10% for atmospheric inlet conditions and 12% for supercharged inlet conditions. The increase of EGR dilution misfire limit from about 10% to 12% with the increase of inlet pressure could be explained by the effect of inlet pressure on residual exhaust gas fraction. The residual exhaust gas fraction decreases with the increase of inlet pressure, which makes the engine more tolerant to EGR dilution. Also, the increase in the in-cylinder unburned mixture density with the increase of inlet pressure slightly increased the combustion rate as described in Section 3.6. The increase of the combustion rate with inlet pressure could also contribute in the increase of the engine tolerance to EGR dilution from 10% to 12%. Misfire occurrence resulted in high cycle to cycle variations and engine instability.
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Fig. 5. In-cylinder pressure data for supercharged inlet conditions and an EGR dilution of 13%.
On the other hand, when the engine was operated with a stoichiometric air–fuel mixture without EGR dilution at a supercharged inlet pressure of 113 kPa, some abnormal combustion pressure cycles were recorded. Abnormal combustion occurrence was identified by the excessively high maximum pressure of some combustion cycles compared to the maximum pressure of the normal combustion cycles for the same operating conditions as shown in Fig. 6. In order to further investigate these abnormal combustion cycles, one abnormal combustion cycle was isolated and used for combustion analysis. Fig. 7 shows the isolated abnormal combustion pressure cycle data, which shows the maximum pressure occurring almost at TDC. This could suggest that either the combustion was too fast or the combustion started too early. Both the heat release rate and the heat release were calculated using this pressure data and shown in Figs. 8 and 9, respectively. Both Figs. 8 and 9 illustrate that the combustion started at about 80° BTDC while the normal spark timing was set at 47° BTDC. This can explain that the air–fuel mixture was pre-ignited before the occurrence of a normal spark. When the engine was operated with a stoichiometric air–fuel mixture without EGR dilution at a supercharged inlet pressure of 113 kPa, the combustion pressure and
Fig. 6. In-cylinder pressure data for supercharged inlet conditions without EGR.
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Fig. 7. Pressure crank-angle data for an abnormal combustion cycle at supercharged inlet conditions without EGR.
Fig. 8. Heat release rate calculated for an abnormal combustion cycle at supercharged inlet conditions without EGR.
Fig. 9. Heat release calculated for an abnormal combustion cycle at supercharged inlet conditions without EGR.
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temperature excessively increased and caused some overheated spots inside the cylinder which ignited the air–fuel stoichiometric mixture earlier during some of the combustion cycles. This phenomenon is called pre-ignition or surface ignition. Surface ignition is the ignition of the air–fuel mixture by any means other than the spark discharge such as an overheated valve or spark plug. Following surface ignition, a turbulent flame develops at each surface ignition location and starts to propagate across the combustion chamber in an analogues manner to what occurs with normal spark-ignition [18]. It is worth mentioning that surface ignition differs from engine knock. Engine knock occurs when essentially a portion of the air– fuel mixture ahead of the propagating flame (the end gas) is auto-ignited due to high in-cylinder temperature and pressure. When knock occurs, much of the chemical energy of the end gas is released rapidly causing high local pressures and the propagation of pressure waves of substantial amplitude across the combustion chamber. Both surface ignition and knock can lead to engine instability and may cause engine damage. Engine knock can be identified using the recorded cylinder pressure data. Knock causes in-cylinder pressure to fluctuate with different amplitudes and frequencies according to knock severity during a combustion cycle. Light knock leads to pressure fluctuations with smaller amplitudes and frequencies compared to heavy knock. The surface ignition combustion pressure signal shown in Fig. 7 has no fluctuations, which indicates that the surface ignition was not accompanied with knock occurrence. Fig. 8 shows that there is a noticeable large portion of negative net heat release rate near the end of combustion which indicates that large amount of cylinder heat transfer occurred due to the excessive in-cylinder temperature caused by pre-ignition. In addition, the too advanced start of combustion for the pre-ignition combustion cycle resulted in a substantial increase in compression work, and consequently, the indicated power calculated for the pre-ignition combustion cycle was essentially lower compared to the indicated power calculated for normal combustion cycles. This increased cycle to cycle variations and led to engine instability. In order to show if the setting of the advanced spark timing of 47° BTDC was the reason behind the overheated hot spots inside the cylinder which caused surface ignition, the spark timing was retarded from 47° to 42°, 37°, and 35° BTDC, respectively for the undiluted stoichiometric air–fuel mixture and the in-cylinder pressure was measured for each spark timing condition. It was found that pre-ignition occurred at all the retarded spark timing conditions. In addition, both pre-ignition and misfire took place with a spark timing of 35° BTDC as shown in Fig. 10. On the other hand, when the stoichiometric air–fuel mixture was diluted with cooled EGR at a high inlet pressure of 113 kPa and a spark timing of 47° BTDC, the pre-ignition combustion cycles entirely disappeared with a percentage of EGR dilution in the inlet mixture of about 5%. This indicates the positive effect of EGR dilution on reducing in-cylinder combustion temperature and thermal stresses, which can eliminate abnormal combustion occurrence such as surface ignition. The presence of high specific heat gases such as water vapor in the EGR in addition to the increase of the total inlet mass with the increase of the percentage of EGR dilution in the inlet mixture help reduce the combustion temperature. Also, the increase of percentage of EGR dilution in the inlet mixture decreases oxygen concentration and slows down the combustion rate which leads to a decrease in both maximum cylinder pressure and temperature. 3.3. Engine stability One important measure of cyclic variability, which can be calculated from the measured cylinder pressure data, is the coefficient
Fig. 10. Pressure crank-angle data at supercharged inlet conditions without EGR and with a retarded spark timing of 35° BTDC.
of variation in indicated mean effective pressure, COV (%), which can be calculated as follows [18]:
COV ¼
rimep imep
100
ð2Þ
where imep (kPa) is the average indicated mean effective pressure calculated for a number of cycles, n, while rimep (kPa) is the standard deviation in indicated mean effective pressure. Both parameters can be calculated as follows [19]:
imep ¼
i¼n X
, imepðiÞ
n
ð3Þ
i¼1
rimep
vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi u i¼n u 1 X ¼t ðimepðiÞ imepÞ2 n 1 i¼1
ð4Þ
The COV defines the cycle to cycle variations in indicated work per cycle and it can be considered as a good indicator for engine stability determination. Heywood [18] demonstrated that vehicle drivability problems usually result when the coefficient of variation in indicated mean effective pressure, COV, exceeds about 10%. However, other studies showed that engine stability starts to deteriorate when COV increases above 5% [20]. Fig. 11 shows the effect of varying the percentage of EGR dilution in the inlet mixture on the coefficient of variation in indicated mean effective pressure, COV, for both atmospheric and supercharged inlet conditions. The dilution of the inlet mixture with EGR up to 10% at atmospheric inlet conditions resulted in a modest increase in COV. However, when the percentage of EGR dilution increased above 10%, COV started to increase rapidly due to misfire occurrence, which was identified for some of the combustion pressure cycles. On the other hand, when the inlet mixture was diluted with EGR at higher inlet pressure, the COV was improved compared to the undiluted inlet mixture condition as EGR dilution reduced thermal stresses and eliminated the pre-ignition occurrence. The COV decreased from 7% with no EGR dilution to about 1% with 10% EGR dilution. However, when the percentage of EGR increased above about 12%, the COV started to increase rapidly due to misfire occurrence. Fig. 11 shows that the increase in inlet pressure increased engine tolerance to EGR dilution from about 10% to 12%. The maximum tolerable EGR dilution limit, which is about 10% for the atmospheric inlet condition and about 12% for the
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Fig. 11. The variation of coefficient of variation in indicated mean effective pressure with the change of percentage of EGR dilution at both atmospheric and supercharged inlet conditions.
supercharged inlet condition and for the operating conditions shown in Table 3, might be considered low compared to the maximum tolerable limit for some modern engines which might tolerate EGR dilution limits of more than 15%. This is because the Ricardo engine had an essentially flat type combustion chamber with the spark plug located on its side. This type of combustion chamber does not support the creation of a significant flow motion such as swirl during combustion and consequently the combustion rate is not rapid enough to tolerate substantially high percentages of EGR dilution conditions. 3.4. In-cylinder pressure Both Figs. 12 and 13 show the effects of varying the percentage of EGR dilution in the inlet mixture on in-cylinder pressure for both atmospheric and supercharged inlet pressures, respectively, with a spark timing of 47° BTDC. The maximum cylinder pressure decreases while the crank angle at which the maximum cylinder pressure occurs is further shifted away from TDC with the increase of percentage of EGR. This
Fig. 12. The effect of varying the percentage of EGR dilution in the inlet mixture on in-cylinder pressure at atmospheric inlet conditions and spark timing of 47° BTDC.
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Fig. 13. The effect of varying the percentage of EGR dilution in the inlet mixture on in-cylinder pressure at inlet pressure of 113 kPa and spark timing of 47° BTDC.
can confirm that EGR dilution is capable of reducing in-cylinder stresses, and consequently, it can be used to suppress abnormal combustion occurrence such as surface ignition and knock. Fig. 14 shows the effect of the increase in inlet pressure on maximum cylinder pressure at different percentages of EGR dilution and spark timing of 47° BTDC. The increase in inlet pressure increased the maximum cylinder pressure while the dilution of the inlet mixture with EGR reduced it to lower levels. The increase of the inlet pressure of the undiluted air–fuel mixture from 101 to 113 kPa increased the maximum cylinder pressure by about 16.6% as it increased from 44.6 to 52 bar. However, when the inlet mixture was diluted with EGR with a percentage dilution of 12%, the maximum cylinder pressure decreased by about 31.5% as it decreased from 52 to 35.6 bar as shown in Fig. 14. 3.5. Heat release rate and heat release In order to show the effect of the increase of percentage of EGR dilution in the inlet mixture on both the net heat release rate and net heat release at both atmospheric and supercharged inlet conditions, the pressure signals shown in Figs. 12 and 13 were analyzed
Fig. 14. The effect of the increase of inlet pressure on maximum cylinder pressure at different percentages of EGR dilution and spark timing of 47° BTDC.
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for combustion analysis. Both Figs. 15 and 16 show the effect of varying the percentage of EGR dilution in the inlet mixture on the net heat release rate for both atmospheric and supercharged inlet conditions, respectively. It can be noted from both Figs. 15 and 16 that the maximum heat release rate decreases with the increase of percentage of EGR dilution. Also, the crank angle at which the maximum heat release rate occurs is shifted away further from the TDC with the increase of percentage of EGR dilution. The increase of EGR dilution in the inlet mixture decreases the in-cylinder oxygen concentration, and consequently, it reduces the heat release rate. Both the heat release rate in Figs. 15 and 16 were integrated to calculate the net heat release. Both Figs. 17 and 18 show the effect of the change of percentage of EGR dilution in the inlet mixture on the heat release for both atmospheric and supercharged inlet conditions, respectively. The heat release crank-angle relationship has the characteristic S-shape. The heat release starts with almost zero at the time of spark which was fixed at 313° (or 47° BTDC), and then it starts to increase 10–20° after the spark timing. This angle is usually called flame development angle or sometimes ignition delay. Fol-
Fig. 15. The effect of varying EGR dilution in the inlet mixture on the net heat release rate at atmospheric inlet condition and spark timing of 47° BTDC.
Fig. 17. The effect of varying the percentage of EGR dilution in the inlet mixture on the net heat release at atmospheric inlet condition and spark timing of 47° BTDC.
Fig. 18. The effect of varying the percentage of EGR dilution in the inlet mixture on the net heat release at an inlet pressure of 113 kPa and spark timing of 47° BTDC.
lowing the flame development duration, the heat release noticeably increases with crank angle until it reaches its maximum value where essentially almost all of the fuel chemical energy has been released at basically the end of combustion. Both Figs. 17 and 18 indicate that the maximum heat release, which essentially identifies the end of combustion, occurs later as the percentage of EGR dilution increases. 3.6. Combustion duration
Fig. 16. The effect of varying EGR dilution in the inlet mixture on the net heat release rate at an inlet pressure of 113 kPa and spark timing of 47° BTDC.
Fig. 19 shows the effect of varying the percentage of EGR dilution in the inlet mixture on the total combustion duration for both atmospheric and supercharged inlet conditions. The total combustion duration was calculated as the crank angle interval from the spark timing to essentially the end of combustion where the heat release reaches its maximum value. Fig. 19 indicates that the increase of percentage of EGR dilution in the inlet mixture increases the total combustion duration for both atmospheric and supercharged inlet conditions. The increase of EGR dilution decreased the oxygen concentration during combustion which slowed down the combustion rate and increased the total combustion duration.
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Fig. 19. The variation of the total combustion duration with the percentage of EGR dilution at atmospheric and supercharged inlet conditions.
Fig. 21. The effect of the increase of percentage of EGR dilution in the inlet mixture on NO emissions at two different inlet pressures.
However, the combustion duration slightly decreased with the increase of inlet pressure. The decrease of combustion duration with the increase of inlet pressure could be due to the increase of the unburned mixture density during combustion which slightly increased the burning rate.
compared to the use of diluted inlet mixture at the same inlet pressure of 113 kPa. The dilution of the inlet mixture with EGR dilution of 5% reduced in-cylinder thermal and mechanical stresses which prevented surface ignition occurrence and decreased cylinder heat transfer, and consequently, this resulted in a significant decrease in engine fuel consumption. Fig. 20 shows that the increase of EGR dilution from 5% to 10% has less significant effect on engine fuel consumption, however, the engine fuel consumption started to increase rapidly as the EGR misfire limit approached. Increasing the inlet pressure increased engine power as demonstrated in Fig. 2 and significantly decreased fuel consumption as shown in Fig. 20. Engine fuel consumption decreased by about 7% when 5% of EGR dilution was employed at an inlet pressure of 113 kPa compared to the use of undiluted inlet mixture with an inlet pressure of 101 kPa. However, the use of the maximum EGR dilution limit of 12% decreased engine fuel consumption by only 1% compared to using an inlet pressure of 101 kPa without EGR. The supercharger work was neglected during brake power and bsfc calculations as the mechanical supercharger used with the Ricardo single cylinder engine was assumed to simulate the use of a turbocharger in a multi-cylinder engine.
3.7. Engine fuel consumption Fig. 20 shows the effect of varying the percentage of EGR dilution in the inlet mixture on brake specific fuel consumption (bsfc) for both atmospheric and supercharged inlet conditions. EGR dilution had little effect on fuel consumption when it was added to the inlet mixture at atmospheric inlet pressure with lower percentage of dilution of up to about 8%. However, when the EGR dilution was increased above 8%, engine fuel consumption started to increase rapidly as the EGR misfire limit approached. On the other hand, the use of an undiluted stoichiometric air– fuel mixture at an inlet pressure of 113 kPa increased the in-cylinder pressure and temperature significantly which led to surface ignition occurrence in addition to a substantial increase in heat transfer, which resulted in an increase in engine fuel consumption
3.8. NO emissions Fig. 21 shows the effect of the increase of percentage of EGR in the inlet mixture on NO emissions at both atmospheric and supercharged inlet conditions. Fig. 21 indicates that the use of EGR dilution has a significant capability on reducing NO emissions. The increase of EGR dilution in the inlet mixture decreased both the combustion temperature and oxygen concentration which led to a significant decrease in NO emissions. NO emissions decreased by about 50% when the EGR dilution was employed at a rate close to the misfire limit with an inlet pressure of 113 kPa compared to the use of undiluted inlet mixture with an inlet pressure of 101 kPa. On the other hand, using 5% of EGR dilution with an inlet pressure of 113 kPa reduced NO emissions by about 12% compared to using undiluted inlet mixture with an inlet pressure of 101 kPa. 4. Conclusions Fig. 20. The variation of brake specific fuel consumption with the percentage of EGR dilution at atmospheric and supercharged inlet conditions.
In this paper, the use of EGR strategy was experimentally investigated in a natural gas spark-ignition engine. Engine performance
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and NO emissions were studied for both atmospheric and supercharged inlet conditions and the following conclusions have been obtained: 1. The increase of the percentage of EGR dilution in the inlet mixture decreases the oxygen concentration, and consequently, it decreases the combustion rate significantly. For instance, the increase of EGR dilution from 0% to 10% at atmospheric inlet conditions increased the total combustion duration from about 74° to 95°. 2. The decrease in combustion rate with the increase of EGR dilution in the inlet mixture helps reduce both the maximum cylinder pressure and temperature, and consequently, prevents abnormal combustion occurrence such as surface ignition, which occurred at higher inlet pressure. 3. The increase of EGR dilution in the inlet mixture decreases both the maximum cylinder temperature and oxygen concentration which leads to a significant reduction in NO emissions. For instance, NO emissions decreased by about 50% when EGR dilution increased from zero with an inlet pressure of 101 kPa to close to the misfire limit with an inlet pressure of 113 kPa. 4. The engine tolerance to EGR increases with the increase of inlet pressure. The maximum tolerable EGR dilution limit increased from about 10% to 12% when the inlet pressure increased from 101 to 113 kPa. 5. The use of an undiluted stoichiometric air–fuel mixture at higher inlet pressure deteriorates engine stability. However, engine stability was improved when the inlet mixture was sufficiently diluted with EGR. 6. Engine performance can be significantly improved when the EGR dilution strategy is employed at higher inlet pressure. Engine brake power increased by about 20% and engine fuel consumption decreased by about 7% while NO emissions decreased by about 12% when 5% of EGR dilution was employed with an inlet pressure of 113 kPa compared to using undiluted stoichiometric inlet mixture with an inlet pressure of 101 kPa. References [1] Baldassarri LT, Battistelli CL, Conti L, Crebelli R, De Berardis B, Iamiceli AL, et al. Compressed natural gas and comparison with liquid fuels. Sci Total Environ 2006;355(1–3):64–77.
[2] US Environmental Protection Agency, EPA, web information: