Energy Conversion and Management 50 (2009) 3129–3139
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Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman
A comparison between EGR and lean-burn strategies employed in a natural gas SI engine using a two-zone combustion model Amr Ibrahim *, Saiful Bari Sustainable Energy Centre, School of Advanced Manufacturing and Mechanical Engineering, University of South Australia, Mawson Lakes SA 5095, Australia
a r t i c l e
i n f o
Article history: Received 13 June 2008 Received in revised form 21 January 2009 Accepted 21 August 2009 Available online 18 September 2009 Keywords: EGR SI engine Lean burn Natural gas NO
a b s t r a c t Exhaust gas recirculation (EGR) strategy has been recently employed in natural gas SI engines as an alternative to lean burn technique in order to satisfy the increasingly stringent emission standards. However, the effect of EGR on some of engine performance parameters compared to lean burn is not yet quite certain. In the current study, the effect of both EGR and lean burn on natural gas SI engine performance was compared at similar operating conditions. This was achieved numerically by developing a computer simulation of the four-stroke spark-ignition natural gas engine. A two-zone combustion model was developed to simulate the in-cylinder conditions during combustion. A kinetic model based on the extended Zeldovich mechanism was also developed in order to predict NO emission. The combustion model was validated using experimental data and a good agreement between the results was found. It was demonstrated that adding EGR to the stoichiometric inlet charge at constant inlet pressure of 130 kPa decreased power more rapidly than excess air; however, the power loss was recovered by increasing the inlet pressure from 130 kPa at zero dilution to 150 kPa at 20% EGR dilution. The engine fuel consumption increased by 10% when 20% EGR dilution was added at inlet pressure of 150 kPa compared to using 20% air dilution at 130 kPa. However, it was found that EGR dilution strategy is capable of producing extremely lower NO emission than lean burn technique. NO emission was reduced by about 70% when the inlet charge was diluted at a rate of 20% using EGR instead of excess air. Ó 2009 Elsevier Ltd. All rights reserved.
1. Introduction Natural gas is one of the cleanest economically available fuels for internal combustion engines. Studies around the world have shown that engines running on natural gas emit significantly lower emissions compared to engines running on conventional fuels. For instance, Baldassari and coworkers [1] compared natural gas and diesel engine emissions, they showed that SI natural gas engine emissions of THC, NOx, and PM were significantly lower than that of the diesel fueled engine with a reduction of 67%, 98%, and 96% respectively. Compared to gasoline engine emissions, another study showed that natural gas SI engines have the potential to achieve a reduction in CO, CO2, NOx, and non methane hydrocarbon emissions of 90–97%, 25%, 35–60%, and 50–75% respectively [2]. Catania and coworkers [3] showed that natural gas engine emissions have less impact on the global warming than gasoline emissions, taking the global warming potential of the methane into account, the authors concluded that the natural gas fueled engine showed a carbon dioxide equivalent reduction of 15–24% with respect to gasoline. In addition to its lower pollution impact, natural * Corresponding author. Tel.: +61 8 8302 5123; fax: +61 8 8302 3380. E-mail address:
[email protected] (A. Ibrahim). 0196-8904/$ - see front matter Ó 2009 Elsevier Ltd. All rights reserved. doi:10.1016/j.enconman.2009.08.012
gas is available in many parts of the world that have poor oil reserves. Using natural gas as an alternative clean fuel will decrease the dependence on imported oil in these countries. Furthermore, the world reserves of natural gas are larger than the petroleum oil, thus the research in utilizing natural gas in engines represents an investment for the future. Recently, environmental and economical concerns have motivated many governments to expand in natural gas infrastructure in order to be feasible to passenger vehicles as well as stationary engines. One of the natural gas engine combustion technologies, which begun in the early 1980s, is the ‘‘lean burn” combustion technique. This technology became dominant in gas engine industry as it led to high engine efficiency accompanied with longer durability and lower cost. Today after almost a quarter century of continuous lean burn engine development and investment, most of the conventional gas engines operate with lean burn mode. According to the Engine Manufacturers Association, USA 2004, over 80% of all heavy duty stationary natural gas engines sold in the USA employ lean burn combustion technology [4]. Most of the research conducted in the lean-burn strategy basically focused on extending the maximum burning lean limit in order to reduce NOx emissions to satisfy the increasing emission restrictions. That usually was achieved by designing fast-burning combustion chambers and/or
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Nomenclature B k L rc Sl Sp T Xb Z Dh Dhb Dhd ho
m
cylinder bore, m rate constant, m3/kmol s distance between cylinder head and piston, m compression ratio laminar flame speed, m/s mean piston speed, m/s temperature, K burned gas fraction mole fraction combustion angle, rad rapid burning angle, rad flame development angle, rad crank angle at start of combustion, rad kinematic viscosity, m2/s
Subscripts b burned u unburned Abbreviations bsfc brake specific fuel consumption EGR exhaust gas recirculation MBT maximum brake torque PM particulate matter SCR selective catalytic reduction SI spark ignition THC total hydrocarbon TWC three way catalyst WOT wide open throttle
employing the stratified charge concept, usually by using either a combustion pre-chamber or direct fuel injection. Recently, laser ignition systems have been developed in order to ignite extremely lean fuel air mixtures, which require high ignition energy. Currently, increasingly stringent ambient air quality standards demand engine emissions to be extremely low; see Table 1 [5]. In order for the engine under the lean burn mode to produce lower NOx emissions, it has to operate with a leaner mixture. In other words, the engine has to operate near the misfire limit to produce relatively lower NOx emissions. As the engine operates near the misfire limit, the engine stability deteriorates, the hydrocarbon (HC) and CO emissions increase, and the engine efficiency decreases. 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. Therefore, it seems that any efforts towards a future decrease in NOx emissions would lead to an increase in HC emission and a decrease in engine thermal efficiency. At the end, a compromise must be made between the increase in NOx emissions and the decrease in engine efficiency. 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 especially for NOx emissions without using exhaust gas after-treatment. The current emission reduction technologies used for the 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, 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 [6]. In addition, an oxidation catalyst would also be necessary to reduce both the HC and CO emissions. It could be concluded that in order for the engines to meet the future emission standards, some alternative techniques must be investigated and developed. One of these alternative techniques is the use of a three way catalyst (TWC) to reduce NOx, HC, and CO emissions. The three way catalyst technology was developed
Table 1 Emission standards, g/kW h [5]. Year
Standard
CO
HC
NOx
PM
1996 2000 2005 2008
Euro2 Euro3 Euro4 Euro5
4 2.1 1.5 1.5
1.1 0.66 0.46 0.46
7 5 3.5 2
0.15 0.1 0.02 0.02
in the 1970s for the automobile industry to reduce the gasoline engine emissions. 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 at near stoichiometric fuel– air ratio (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. This would lead to some restrictions on the use of turbocharging, high compression ratio, and maximum brake torque (MBT) spark advance timing. As a result, the engine would operate less efficiently than a similar lean burn engine. 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 into the cylinder intake with the inlet mixture. This method is called exhaust gas recirculation (EGR). Using EGR with the stoichiometric inlet mixture will lead to a decrease in the in-cylinder temperature and a decrease in knocking tendency and could permit the engine to use turbocharging, relatively higher compression ratio, and MBT spark advance timing to achieve a relatively higher thermal efficiency compared to non diluted stoichiometric mixture operation. In addition, 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. Furthermore, as EGR will be added to a stoichiometric mixture, the use of a TWC for necessary emission reductions is also possible. Although the use of EGR with a TWC technique is expected to economically produce lower emissions than lean-burn strategy, the effect of using EGR compared to lean burn on some of engine performance parameters such as engine fuel consumption is still not quite certain. Some conflicting results were found in the literature review regarding to this issue. For instance, Corbo and coworkers [7] converted a heavy duty turbocharged diesel engine to work on natural gas fuel. They employed both lean burn and stoichiometric mixture with EGR and a TWC approaches after conversion and compared the engine performance and emissions for both cases. The authors concluded that the use of both EGR and lean burn techniques led to a similar maximum thermal efficiency of 34%. Nellen and Boulouchos [8] used a stoichiometric mixture with cooled EGR and a TWC in a turbocharged natural gas SI engine used for cogeneration applications. The authors aimed to achieve low emissions and high efficiency by using this concept. The authors optimized the same engine for lean burn operation with an oxidation catalyst. They concluded that the EGR concept resulted in a more superior engine performance and emissions compared to the lean burn technique. The engine achieved a thermal
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efficiency of 40% at bmep of 12 bar by using EGR technique compared to 38% at the same bmep for the lean burn operation. Saanum and Bysveen [9] compared the use of lean burn and stoichiometric mixture with EGR strategies in a natural gas SI engine. They found that the maximum brake thermal efficiency was higher for lean burn operation than for EGR operation. The authors concluded that a penalty in engine efficiency must be accepted when EGR is used as an alternative to excess air. Perhaps these conflicting results are due to the differences in the optimized operating conditions and engine design parameters used for each technique. For example the lean burn technique would give higher efficiency if the spark timing was optimized for maximum brake torque (MBT) condition rather than low NOx emission condition. In the current study, both EGR and lean-burn strategies are compared and assessed regarding to their effect on a natural gas engine performance and NO emission at similar operating conditions. This would help to assess the use of EGR as an alternative to excess air and identify some of the operating conditions and engine design parameters that can be optimized for better engine performance when EGR technique is employed. For this purpose, a computer simulation of the four-stroke spark-ignition natural gas engine was developed. A two-zone combustion model was constructed to simulate the in-cylinder conditions during combustion. The simulation has been validated by experimental results and a good agreement between the results was found. 2. Model description The following assumptions and approximations are considered for simplification: 1. The contents of the cylinder are fully mixed and spatially homogeneous in terms of composition and properties during intake, compression, expansion, and exhaust processes. Thus, the thermodynamic properties vary only with time (or crank angle). 2. For the combustion process, two zones (each is spatially homogeneous) are used. The two zones are the unburned and the burned zones. The two zones are separated from each other by the flame front (see Fig. 1). 3. The intake and exhaust manifolds are assumed to be infinite plenums containing gases at constant temperature and pressure. The exhaust pressure was set at a value of 102 kPa, which is slightly higher than the atmospheric pressure.
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4. All gases are considered to be ideal gases during the engine thermodynamic cycle. 5. All crevice effects are ignored, and the blow-by is assumed to be zero. 6. The cylinder wall temperature is assumed to be constant (400 K) and the heat transfer is determined using Woschni correlation [10]. 7. The engine is in steady state such that the thermodynamic state at the beginning of each thermodynamic cycle (two crankshaft revolutions) is the same as the end state of the cycle. The flow rates in both the intake and exhaust processes were determined from quasi-steady one-dimensional compressible flow rate equations [10]:
vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi " #ffi 1=c u c1 u 2c c C d AR po pt p t t _ ¼ pffiffiffiffiffiffiffiffi m 1 c1 po RT o po
ð1Þ
_ is the mass flow rate through intake and exhaust valves, where m Cd is the discharge coefficient (assumed to be 0.7), AR is the reference area which was selected to be equivalent to the curtain area as suggested by Heywood [10] (AR = p dvlv(t), where dv is valve diameter, lv(t) is valve lift as a function of time (or crank angle)), To and po are stagnation temperature and pressure upstream of the valve respectively, pt is static pressure down stream of the valve, and finally c and R are specific heat ratio and gas constant of the mixture flowing through the valve respectively. For flow into the cylinder through an intake valve, po is the intake manifold pressure, and pt is the cylinder pressure. For flow out of the cylinder through an exhaust valve, po is the cylinder pressure, and pt is the exhaust pressure. c c1 2 When the flow through the valve is choked, i.e. ppot 6 cþ1 , the mass flow rate is calculated from the following equation [10]:
vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi u ccþ1 1 C d AR p o u 2 _ ¼ pffiffiffiffiffiffiffiffi tc m cþ1 RT o
ð2Þ
In the present model, the thermodynamic cycle simulation starts with assumed guesses of the values of pressure and temperature of the contents within the cylinder at the instant the intake valve opens. After two crankshaft revolutions (720 crank angle degrees), the calculated values of pressure and temperature are compared to the initial guesses. If the calculated values are not within an acceptable tolerance to the initial guesses, the simulation is repeated using the final calculated values as initial guesses. 2.1. The combustion process The following assumptions are assumed during combustion:
Fig. 1. Schematic of the two-zone combustion modeling.
1. The flame front thickness is assumed to be negligible. 2. The cylinder pressure is assumed to be the same in the burned and unburned zones. 3. Only the convective heat transfer mode, between the cylinder contents and the cylinder wall, is considered. 4. The heat transfer between the two zones is neglected. 5. For the burned zone, ten species (CO2, H2O, CO, N2, O2, OH, NO, H, O, and H2) are considered in chemical equilibrium during combustion and expansion. 6. The combustion chamber wall area in contact with the burned gases is assumed to be proportional to the square root of the burned mass fraction to account for the greater volume filled by burned gases against the unburned volume as suggested by Ferguson [11].
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2.1.1. The thermodynamic formulations Fig. 1 is a schematic of the engine cylinder during combustion, which shows the cylinder heat transfer from both the unburned (u) and burned (b) zones, and the piston work. The basic relations used in the development of the present simulation are the first law of thermodynamics, the conservation of mass law, and the ideal gas law. These three principles were applied to both the unburned and burned control volumes in order to derive expressions for the time (or crank angle) derivative of the unburned and burned gas temperatures and volumes in addition to the cylinder pressure during combustion. These expressions are expressed in terms of engine design parameters and operating conditions. The Euler numerical solution technique as described by Caton [12] was used to solve the differential equations to determine the in-cylinder pressure and temperature. 2.1.2. The burning rate The S-shaped mass fraction burned profile, the Wiebe function, was used to determine the burning rate:
" mþ1 # h ho X b ¼ 1 exp a Dh
ð3Þ
where h is the crank angle, ho is the crank angle at the start of combustion, Dh is the total combustion duration (from Xb = 0 to Xb = 1), and a and m are adjustable parameters which fix the shape of the curve. Actual mass fraction burned curves have been fitted witha 5, and m 2 as suggested by Heywood [10]. The empirical rule for relating the mass burning profile to crank angle at maximum brake torque (MBT) spark timing is used in this model. With optimum spark timing, half the charge is burned at about 10 crank angle degrees after top dead centre [10]. Thus, referring to Eq. (3), putting Xb = 0.5 at h = 370 degrees enables ho to be determined at a specified combustion duration. In the current study, all results were obtained at the MBT spark timing condition. 2.1.3. The combustion angle A turbulent flame propagation model developed by Tabaczynski and coworkers [13] was used by Hires and coworkers [14] to obtain explicit relations for the flame development angle, Dhd, and the rapid burning angle, Dhb, as function of engine design and operating variables:
2=3 L Sl 10=9 2=3 B qi Sp m Þ1=3 Li Dhb ¼ C 0 ð L qu Sl
Dhd ¼ CðSp mÞ1=3
ð4Þ ð5Þ
where m is the kinematic viscosity, L is the distance between cylinder head and piston, Sl is the laminar burning speed, Sp is the mean piston speed, q is the density, and B is the cylinder bore. The subscript i denotes the value at ignition and the subscript u refers to the unburned mixture, whereas the superscript () denotes the va0 lue at cylinder conditions where Xb = 0.5. C and C are constants, which depend on engine geometry. The empirical correlation of the laminar burning speed of the natural gas-air-EGR mixture was determined from Ref. [15]:
Sl ¼ Sl0
a Tu p b ð3:4259 x2EGR 3:6993 xEGR þ 1:002Þ 100 300
Sl0 ¼ 177:43 U3 þ 340:77 U2 123:66 U 0:2297 2
a ¼ 5:75 U 12:15 U þ 7:98 2
b ¼ 0:925 U þ 2 U 1:473
ð6Þ ð7Þ ð8Þ ð9Þ
where Sl0 is the reference burning velocity, cm/s, Tu is the unburned mixture temperature, K, p is the cylinder pressure, kPa, a and b are fitting coefficients, xEGR is the volumetric fraction of EGR in the unburned mixture, and U is the equivalence ratio. The empirical laminar flame speed correlation was validated for equivalence ratio range of 0.49–1.43, pressure range from 50 to 1000 kPa, EGR ratio range from 0 to 0.43, and the tested temperature ranged from 300 to 400 K [15]. Since the dynamic viscosity of hydrocarbon-air combustion products differs little from that of air as demonstrated by Heywood [10], the cylinder content dynamic viscosity could be expressed using the air dynamic viscosity correlation which has the following form [10]:
l ¼ 3:3 107 T 0:7
ð10Þ
where l is the dynamic viscosity in kg/ms, and T is the temperature in K. Hence, the kinematic viscosity can be determined using the well known correlation: m ¼ lq . Both Eqs. (4) and (5) are used in the present model in order to calculate the combustion duration (Dh = Dhd + Dhb) at different operating conditions. The combustion duration is then used to determine the burned mass fraction using the Wiebe function.
2.2. NO formation kinetic model The extended Zeldovich mechanism [10] is used to determine the rate of change of NO mole fraction during combustion and expansion processes as follows:
2r 1 1 ðZ NO =Z NO;e Þ2 dZ NO ¼ 1 þ ðZ NO =Z NO;e Þr 1 =ðr2 þ r3 Þ dt
ð11Þ
where: þ p r1 ¼ k1 Z O;e Z N2;e RT b p r2 ¼ k2 Z NO;e Z O;e RT b p r3 ¼ k3 Z NO;e Z H;e RT b
ð12Þ ð13Þ ð14Þ
is the uniwhere Z is the mole fraction, p is the cylinder pressure, R versal gas constant, and Tb is the burned gas temperature. The subscript e refers to equilibrium value. The rate constants (k), in units of m3/kmol s, were calculated from Ref. [10]. The prescribed combustion model with the integration of both NO formation and knocking sub-models was previously used to optimize a natural gas engine performance under EGR operation in two different studies; more details can be found in Ref. [16,17].
Table 2 Engine specifications. Number of cylinders Bore, mm Stroke, mm Capacity, cc Maximum speed, rpm Max. cylinder pressure, bar Inlet valve opens, deg BTDC Inlet valve closes, deg ABDC Exhaust valve opens, deg BBDC Exhaust valve closes, deg ATDC
1 76.2 111.125 507 3000 150 9 34 43 8
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3. Experimental set-up and model validation 3.1. Experimental set-up In order to validate the prescribed model, experimental data was collected from a 507 cc single cylinder variable compression ratio Ricardo E6 SI engine. Table 2 shows the Ricardo engine specifications. The engine used to run on petrol before it was converted to run on natural gas. Fig. 2 shows a schematic of the experimental set-up. The recycled exhaust gas was taken from a hole located on the exhaust pipe with the help of a small vacuum pump. The hot exhaust gas was cooled by passing it through a water-cooled heat exchanger. A regulating valve was used to control the amount of the recycled exhaust gas while a 6 mm square orifice flow meter was used to measure the flow rate of the exhaust gases recycled back to engine intake as follows:
pffiffiffiffiffiffiffiffiffiffiffiffi _ ¼ C d A 2qDp m
ð15Þ
_ is the mass flow rate across the orifice meter, Cd is the diswhere m charge coefficient which is equivalent to 0.6, A is the orifice hole area, q is the density of the gas downstream of the orifice meter, and Dp is the pressure difference across the orifice meter measured by a u-tube manometer with an accuracy of ±0.01 kPa. A supercharger was installed in order to provide the engine with inlet charge at high pressure. The supercharger was driven by an electrical motor via a belt. Air, natural gas, and cooled exhaust gas were mixed in the supercharger intake before they were delivered at high pressure to an intercooler. The intercooler cooled down the air–fuel–exhaust gas mixture before it was delivered to
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the engine intake. A pressure gauge was used to measure the inlet pressure with an accuracy of ±2 kPa. The natural gas was supplied from a pipe line which supplied a continuous flow of natural gas at a pressure of slightly higher than the atmospheric pressure. The natural gas flow rate, and hence the air–fuel ratio, was controlled by a regulating valve. The natural gas flow rate was measured using Dwyer RMC flow meter with an accuracy of ±0.045 m3/h while the air flow rate was measured using Alcock viscous flow meter with an accuracy of ±0.09 m3/h. The pressure drop across the air flow meter, which was measured using an inclined water manometer, was used with meter calibration constant to calculate the air flow as follows:
V_ a ¼ ch
ð16Þ
where V_ a is the air flow down stream the meter cell in m3/hr, c is the calibration constant (c = 4.57), and h is the pressure difference across the meter in cm water. The engine load was applied by an electrical direct current dynamometer, which is capable to run as a variable speed motor in order to determine the engine friction power. The engine load was measured with an accuracy of ±0.4 Nm by balancing the torque exerting on the dynamometer casing with an equivalent opposite torque by hanging weights on the torque arm which is attached to the dynamometer casing. A spring balance of maximum force of 50 N was connected to the arm torque to provide an upward force for fine balancing adjustment. The reading was taken when the torque arm was horizontal and the dynamometer casing was floated. Toque was calculated as follows:
T ¼FR
Fig. 2. Schematic of the experimental set-up.
ð17Þ
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where F is the net force on the torque arm, and R is the radius of the torque arm which is equivalent to 45.7 cm. The engine speed was measured by a mechanical tachometer with an accuracy of ±10 rpm. The in-cylinder pressure was measured using water-cooled piezoelectric Kistler type 7061B pressure sensor, TDC position sensor, and shaft encoder. A charge amplifier type Kister 5007 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 numbers which are suitable to be handled by a laptop via a data acquisition card as indicated in Fig. 2. Engine NO emissions were measured using the chemiluminescence technique with an accuracy of ±30 ppm.
3.2. Model validity Fig. 3 shows the p–V diagram as predicted by the model at an inlet pressure of 98 kPa (wide open throttle, WOT, condition), stoichiometric mixture, engine speed of 1000 rpm, engine compression ratio of 8, MBT spark timing, and no EGR inlet condition. The p–V relationship was integrated in order to calculate the indicated work and consequently the indicated power. The engine friction power was determined experimentally using the motoring test and then it was used to calculate the engine brake power. Both brake power and brake specific fuel consumption were calculated at different engine speeds and compared with experimental results as shown in Figs. 4 and 5 respectively. In addition, both measured and calculated in-cylinder pressure data were compared at different operating conditions. Fig. 6 compares the measured and calculated in-cylinder pressure at percentage EGR dilution in the inlet mixture of 10%, inlet pressure of 113 kPa, compression ratio of 8, and 1500 rpm operating conditions. Fig. 7 compares measured and calculated in-cylinder pressure at excess air factor (k) of 1.2, atmospheric inlet conditions, compression ratio of 8, and 1500 rpm operating conditions. Also, both measured and calculated NO emissions were compared at different operating conditions. Fig. 8 compares measured and calculated NO emissions at an inlet pressure of 113 kPa, compression ratio of 8, 1500 rpm, and different %EGR dilution conditions. Fig. 9 compares measured and calculated NO emissions at atmospheric inlet condition, compression ratio of 12, 1500 rpm, and different %EGR dilution conditions.
Fig. 4. A comparison between measured and calculated brake power at WOT, stoichiometric fuel–air mixture, rc = 8, MBT spark timing, and different speed conditions.
Fig. 5. A comparison between measured and calculated brake specific fuel consumption at WOT, stoichiometric fuel–air mixture, rc = 8, MBT spark timing, and different speed conditions.
Figs. 4–9 show that there is a good agreement between the experimental and the computer model results at various operating conditions, which gives the confidence that the computer model has been well constructed. 4. Results and discussion 4.1. The effect of EGR vs. excess air on power Fig. 10 shows the effect of the increase of excess air factor on brake power at an inlet pressure of 130 kPa (which simulates turbocharged inlet pressure conditions), 1500 rpm, and MBT spark timing condition. The excess air factor, k, is defined as the ratio between the actual air–fuel-ratio to the stoichiometric air–fuel–ratio:
k¼ Fig. 3. p–V diagram at inlet pressure of 98 kPa, stoichiometric mixture, rc = 8, MBT spark timing, and 1000 rpm.
ðA=FÞact ðA=FÞst
ð18Þ
Similarly, Fig. 11 shows the effect of the increase of the percentage of the mass of recycled exhaust gases, %EGR, on brake power at
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Fig. 6. A comparison between measured and calculated in-cylinder pressure at.%EGR of 10%, inlet pressure of 113 kPa, rc = 8, and 1500 rpm.
Fig. 7. A comparison between measured and calculated in-cylinder pressure at k = 1.2, atmospheric inlet condition, rc = 8, and 1500 rpm.
Fig. 8. A comparison between measured and calculated NO emissions at inlet pressure of 113 kPa, rc = 8, 1500 rpm, and different %EGR dilution conditions.
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Fig. 9. A comparison between measured and calculated NO emissions at atmospheric inlet pressure, rc = 12, 1500 rpm, and different %EGR dilution conditions.
Fig. 10. Variations of brake power with excess air factor at inlet pressure of 130 kPa, 1500 rpm, rc = 10, and MBT spark timing condition.
Fig. 11. Variations of brake power with EGR dilution at inlet pressure of 130 kPa, 1500 rpm, rc = 10, and MBT spark timing condition.
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an inlet pressure of 130 kPa, 1500 rpm, and MBT spark timing condition. %EGR is defined as the ratio between the mass of the recycled exhaust gases to the total inlet mass (i.e. the mass of air, fuel, and recycled exhaust gases):
%EGR ¼
mEGR 100 ma þ mf þ mEGR
ð19Þ
The engine brake power decreases with the increase of both excess air and EGR as they both replace some of the inlet fuel mass because they were added at a constant inlet pressure. In order to compare the effect of both excess air and EGR on engine performance using the same basis of charge dilution, the air dilution is defined in a similar way like EGR dilution. Air dilution can be defined as the ratio between the mass of the excess air to the total inlet mass. The mass of the excess air can be calculated as the difference between the actual mass of air used for combustion and the stoichiometric mass of air used for stoichiometric combustion:
%Air dilution ¼ ¼
Fig. 12. A comparison between the effect of both air and EGR dilution on brake power at inlet pressure of 130 kPa, 1500 rpm, rc = 10, and MBT spark timing condition.
ma;act ma;st 100 ma;act þ mf ðA=FÞact ðA=FÞst k1 100 100 ¼ k þ ðF=AÞst ðA=FÞact þ 1
ð20Þ
Knowing that the stoichiometric fuel–air ratio for natural gas is about 1/15.8, then Eq. (20) expresses the percentage of air dilution as a function of excess air factor, k only. Air dilution is equivalent to zero for stoichiometric mixture and increases with the increase of k as shown in Table 3. Fig. 12 compares the effect of both air and EGR dilution on engine power at an inlet pressure of 130 kPa, 1500 rpm, and MBT spark timing condition. Adding EGR to the stoichiometric mixture decreased power more rapidly than excess air. That was mainly because of the more significant effect of EGR dilution on combustion duration compared to air dilution. Adding EGR to the inlet charge decreased the in-cylinder oxygen concentration and slowed down the combustion rate significantly. The extended combustion duration resulted in burning most of the fuel away from the top dead centre which led to a loss in engine power. In addition, the higher inlet charge temperature in the presence of EGR (assumed to be 333 K) compared to the inlet charge temperature in the presence of air (assumed to be 310 K) reduced the inlet density and hence the power was reduced. The use of EGR dilution at a rate of 20% and inlet pressure of 130 kPa reduced engine power by about 20% compared to using air at the same dilution rate and inlet pressure conditions. However, the loss of engine power which occurred due to the use of EGR instead of excess air was recovered by increasing the inlet pressure. It was found that when inlet pressure increased from 130 kPa at no EGR condition to about 150 kPa at 20% EGR dilution, brake power increased and became equivalent to the corresponding air dilution power results. Fig. 13 shows the increase of combustion duration with both the addition of air at constant inlet pressure of 130 kPa and the addition of EGR at constant and variable inlet pressure. The inlet pressure was varied from 130 to 150 kPa to obtain power equiva-
Table 3 Variations of % air dilution with k. k
Air dilution (%)
1 1.1 1.2 1.3 1.4
0 8.6 15.8 22 27.3
Fig. 13. A comparison between the effect of both air and EGR dilution on combustion duration at 1500 rpm, rc = 10, and MBT spark timing condition.
lent to air dilution power results as it was mentioned earlier. The increase in inlet pressure increased the inlet density and slightly decreased the combustion duration as shown in Fig. 13. Fig. 13 indicates that the increase of EGR dilution has a stronger effect on combustion duration than the increase of excess air dilution. These results agree well with different experimental results such as the results obtained by Einewall and coworkers [18]. Einewall and coworkers [18] compared the effect of lean burn and EGR on the heat release rate using a natural gas engine having a fast-burning combustion chamber with high turbulence. The authors found that EGR has a stronger influence on the flame development angle than excess air since the laminar flame speed was more reduced compared to lean burn operation. The authors concluded that excess air leads to much shorter duration during early combustion compared to EGR for high dilution levels. 4.2. The effect of EGR vs. excess air on bsfc Fig. 14 shows the effect of both air and EGR dilution on brake specific fuel consumption at 1500 rpm and MBT spark timing
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sipated as heat loss. Fig. 15 shows the effect of air and EGR dilution on engine heat loss as a percentage from fuel power. The percentage heat loss was calculated as follows:
%heat loss ¼
Q_ 100 _ f LHV m
ð21Þ
_ f is the fuel mass flow rate, and LHV where Q_ is the heat transfer, m is the lower heating value. The use of EGR dilution at a rate of 20% and inlet pressure of 130 kPa increased engine fuel consumption by about 18% compared to using air at the same dilution rate and inlet pressure conditions. On the other hand, when 20% EGR dilution was added at an inlet pressure of 150 kPa in order to obtain the same amount of power as when 20% air dilution was employed at 130 kPa, the fuel consumption increased by only 10%. 4.3. The effect of EGR vs. air on NO emission Fig. 14. A comparison between the effect of both air and EGR dilution on brake specific fuel consumption at 1500 rpm, rc = 10, and MBT spark timing condition.
condition. The addition of excess air was studied at constant inlet pressure of 130 kPa whereas the addition of EGR was studied for two different inlet pressure conditions, which were the constant 130 kPa inlet pressure condition and the variable inlet pressure condition that led to brake power results similar to air dilution brake power results at an inlet pressure of 130 kPa. The fuel consumption was slightly improved with the increase of air dilution up to about 15%. The increase of air dilution decreased the maximum cylinder burned gas temperature which led to a decrease in the dissociation losses near top dead centre which resulted into an increase in the fraction of sensible energy that was converted to work. However, the fuel consumption started to increase at higher air dilution conditions due to the significant increase of the combustion duration at higher air dilution conditions as shown in Fig. 13. On the other hand, engine fuel consumption increased with EGR dilution which had a more significant effect on combustion duration than excess air. The extended combustion duration led to most of fuel to be burned earlier during compression stroke and later during expansion stroke which resulted in a decrease in power and an increase in the fraction of fuel energy that was dis-
Fig. 15. A comparison between the effect of both air and EGR dilution on engine heat loss as a percentage from fuel power at 1500 rpm, rc = 10, and MBT spark timing condition.
Fig. 16 shows the effect of using both EGR and air dilution on engine NO emission at 1500 rpm and MBT spark timing condition. NO emission increases with the increase of air dilution until it reaches to its peak value at about 9% air dilution which is equivalent to an excess air factor of about 1.1. This is due to NO formation mechanism depends on both temperature and oxygen concentration. Although burned gas temperature is higher for stoichiometric mixture (zero dilution) as shown in Fig. 17, the availability of oxygen at leaner mixtures results in the maximum NO emission to be occurred at mixtures slightly leaner than stoichiometric. The burned gas temperature and consequently NO emission decrease when air dilution is increased above 9%. On the other hand, the use of EGR dilution decreased NO emission significantly as shown in Fig. 16. The use of EGR dilution decreases both temperature and oxygen concentration and leads to lower NO emission compared to the use of air dilution. The use of 20% EGR dilution decreased NO emission by about 70% compared to using the same percentage of air dilution. Furthermore, employing EGR dilution technique would allow using a three way catalyst, which usually has a conversion efficiency of about 90%, for exhaust gas after-treatment. That makes the EGR dilution technique capable of producing extremely lower NO emission compared to using air dilution technique. These results agree well with several experimental results done by different researchers. For example, Bhargava and coworkers
Fig. 16. A comparison between the effect of both air and EGR dilution on NO emission at 1500 rpm, rc = 10, and MBT spark timing condition.
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Fig. 17. A comparison between the effect of both air and EGR dilution on maximum cylinder burned gas temperature at 1500 rpm, rc = 10, and MBT spark timing condition.
[19] studied the effect of replacing air with EGR on natural gas engine emissions at constant load and speed. The authors found that NOx emissions were reduced from 4 to about 1 g/bhp-h with a percentage reduction of about 75% when EGR replaced excess air at the same amount of charge dilution of about 22%. Corbo and coworkers [7] compared engine emissions using both EGR and lean burn techniques for a converted natural gas engine. The authors reported that the use of lean burn technique led to NOx emissions of 4.7 g/kW h whereas the use of EGR technique with a three way catalyst resulted in only 0.1 g/kW h NOx emissions. Although the authors did not report the value of NOx emission before the catalyst, this could be calculated assuming the catalyst conversion efficiency to be comparable to 90%. This would lead NOx emission before the catalyst to be about 1 g/kW h with a percentage reduction of about 78% compared to using lean burn technique. Saanum and Bysveen [9] compared lean burn and EGR operation of an engine fueled with natural gas and hydrogen enriched natural gas. The authors studied the effect of both lean burn and EGR on NOx emissions at various engine loads. It was demonstrated that NOx emission increases with the increase of excess air until it reaches its maximum value at an excess air factor ranged from 1.1 to 1.2 depending on engine load and then decreases with the increase of excess air. On the other hand, the use of EGR reduced NOx emission efficiently with a near linear decrease with the increase of EGR dilution. The authors concluded that EGR technique resulted in much lower NOx emissions compared to using excess air at the same level of dilution.
Fig. 18. A comparison between the effect of both air and EGR dilution on start of combustion angle optimized for MBT condition at 1500 rpm, rc = 10, and MBT spark timing condition.
of dilution increased the combustion duration significantly as shown in Fig. 13. That led to an excessive advance for the start of combustion timing which was optimized for MBT condition as shown in Fig. 18. The increased advance of the start of combustion timing led to initiating the combustion earlier in the compression stroke which increased the maximum temperature when EGR dilution was used at rates higher than 15% compared to air dilution. Fig. 19 shows the effect of both EGR and air dilution on maximum cylinder pressure. When both air and EGR dilution were studied at constant inlet pressure of 130 kPa, the use of EGR slowed down the combustion rate more significantly and consequently the maximum cylinder pressure was lower than the corresponding maximum pressure for air dilution strategy. On the other hand, when EGR was added at higher inlet pressure conditions which varied from 130 kPa at zero dilution to about 150 kPa at 20% EGR dilution in order to obtain similar power as for air dilution strategy, the maximum cylinder pressure increased compared to using air dilution at an inlet pressure of 130 kPa. That indicates the significant dependence of the maximum cylinder pressure on inlet pressure.
4.4. The effect of EGR vs. excess air on in-cylinder maximum temperature and pressure Both in-cylinder pressure and temperature are important variables which affect the in-cylinder mechanical and thermal stresses in addition to the formation of NO emission. Fig. 17 shows the variations of maximum burned gas temperature with both EGR and air dilution at 1500 rpm and MBT spark timing condition. The use of air dilution up to 15% resulted in a slight increase in the maximum burned gas temperature compared to using EGR dilution. This is because EGR contains water vapor which has higher specific heat capacity than air. However, when air dilution increased above 15%, the maximum burned gas temperature decreased compared to using EGR dilution. This is because the use of EGR at higher rates
Fig. 19. A comparison between the effect of both air and EGR dilution on maximum cylinder pressure at 1500 rpm, rc = 10, and MBT spark timing condition.
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The addition of EGR to the inlet charge decreased engine power more rapidly compared to using excess air at the same inlet pressure of 130 kPa; however, the loss of power was recovered by increasing inlet pressure from 130 kPa at zero dilution to 150 kPa at 20% EGR dilution. The fuel consumption increased by 18% when the inlet charge was diluted at a rate of 20% using EGR instead of excess air at the same inlet pressure of 130 kPa; however, the fuel consumption increased by only 10% when EGR was added at higher inlet pressure of 150 kPa. The use of EGR is capable of achieving extremely low NO emission compared to excess air. NO emission decreased by about 70% when the inlet charge was diluted at a rate of 20% using EGR instead of air. References
Fig. 20. A comparison between the effect of both air and EGR dilution on average exhaust gas temperature at 1500 rpm, rc = 10, and MBT spark timing condition.
4.5. The effect of EGR vs. air on average exhaust temperature The average exhaust gas temperature can indicate the amount of available thermal energy in the exhaust gases which can be used for a turbocharger in order to increase the inlet pressure. Also, the average exhaust temperature is an important quantity for determining the performance of catalytic converters which are used to achieve further emission reduction. The exhaust temperature was calculated as an enthalpy-averaged temperature using the following integration as suggested by Heywood [10]:
Tex ¼
Z
_ p Tdt mc
Z
_ p dt mc
ð22Þ
_ is the mass flow rate exiting through the exhaust valve, cp where m is the exhaust gas specific heat, and T is the instantaneous exhaust gas temperature. The integration is performed from the time the exhaust valve opens to the closing time. Fig. 20 shows the variations of the average exhaust temperature with both air and EGR dilution at 1500 rpm and MBT spark timing condition. The more significant effect of EGR dilution on combustion duration compared to air dilution led to terminating the combustion process later in the expansion stroke and increased the exhaust gas temperature compared to using air dilution strategy. 5. Conclusions The effect of both EGR and lean burn techniques on engine performance and NO emission was compared numerically using a two-zone combustion model. The following conclusions were obtained: The use of EGR slowed down the combustion rate and increased the combustion duration significantly compared to using air at the same amount of dilution. The longer combustion duration which was achieved when EGR was used resulted in more advance in start of combustion timing which was optimized for MBT condition.
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