Optimization of a natural gas SI engine employing EGR strategy using a two-zone combustion model

Transcription

1 Available online at Fuel 87 (2008) Optimization of a natural gas SI engine employing EGR strategy using a two-zone combustion model Amr Ibrahim *, Saiful Bari Sustainable Energy Centre, School of Advanced Manufacturing and Mechanical Engineering, University of South Australia, Mowson Lakes, SA 5095, Australia Received 3 August 2007; received in revised form 3 October 2007; accepted 4 October 2007 Available online 29 October 2007 Abstract Natural gas has been recently used as an alternative to conventional fuels in order to satisfy some environmental and economical concerns. In this study, a natural gas spark-ignition engine employing cooled exhaust gas recirculation (EGR) strategy in a high pressure inlet condition was optimized. Both engine compression ratio and start of combustion timing were optimized in order to obtain the lowest fuel consumption accompanied with high power and low emissions. That 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. In addition, a knocking model was incorporated with the two-zone combustion model in order to predict any auto-ignition that might occur. It was found that the value of the compression ratio at which the minimum fuel consumption occurs varies with the engine speed. A minimum fuel consumption of about 200 g/kw h was achieved at an engine speed of 1500 rpm, inlet conditions of 200 kpa and 333 K, and a compression ratio of about 12. Also, it was found that cooled EGR can significantly reduce NO emission at high compression ratio conditions. NO emission decreased by about 28% when EGR was increased from 20% at compression ratio of 10 to 27% at compression of 12 at the same engine speed of 3000 rpm. Ó 2007 Elsevier Ltd. All rights reserved. Keywords: EGR; Engine; Natural gas; Compression ratio; NO x 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 between natural gas and diesel engine emissions, they showed that SI natural gas engine emissions of THC, NO x, and PM were significantly lower than that of the diesel fuelled engine with a reduction of 67%, 98%, and 96%, respectively. Compared to gasoline * Corresponding author. Tel.: ; fax: address: Amr.Ibrahim@postgrads.unisa.edu.au (A. Ibrahim). engine emissions, another study showed that natural gas SI engines have the potential to achieve a reduction in CO, CO 2,NO x, and non-methane hydrocarbon emissions of 90 97%, 25%, 35 60%, and 50 75%, respectively [2]. Catania and co-workers [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 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 /$ - see front matter Ó 2007 Elsevier Ltd. All rights reserved. doi: /j.fuel

2 A. Ibrahim, S. Bari / Fuel 87 (2008) Nomenclature B cylinder bore, m E u temperature coefficient k rate constant, m 3 /kmol s L distance between cylinder head and piston, m r c compression ratio S l laminar flame speed, m/s S p mean piston speed, m/s T temperature, K X b burned gas fraction Z mole fraction Dh combustion angle, rad Dh b rapid burning angle, rad Dh d flame development angle, rad h o crank angle at start of combustion, rad m kinematic viscosity, m 2 /s Subscripts b burned u unburned Abbreviations bsfc brake specific fuel consumption EGR exhaust gas recirculation MBT maximum brake torque SCR selective catalytic reduction SI spark ignition THC total hydrocarbon TWC three way catalyst WOT wide open throttle 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 NO x emissions to satisfy the increasing emission restrictions. That usually was achieved by designing fast-burning combustion chambers and/or 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 NO x 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 NO x 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 NO x 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 NO x 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 NO x 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 NO x emissions without using exhaust gas after-treatment. The current emission reduction technologies used for the NO x 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 NO x, HC, and CO emissions. The three Table 1 Emission standards, g/kw h [5] Year Standard CO HC NO x PM 1996 Euro Euro Euro Euro

3 1826 A. Ibrahim, S. Bari / Fuel 87 (2008) way catalyst technology was developed 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 incylinder NO x 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 concept of using EGR in engines, especially for petrol and diesel engines, is not new, it is believed that natural gas SI engine operation employing stoichiometric mixture with EGR has not been fully optimized yet [4]. Many diesel and gasoline engines are converted to work on natural gas in order to satisfy some economical and environmental concerns in many countries. The number of the converted engines is expected to grow in the future. It is believed that if these engines are optimized in terms of some engine design parameters (for example, compression ratio) and operating conditions (for example, spark timing and inlet conditions), the gas engines will operate more efficiently. The improvement in engine fuel consumption does not lead only to saving in fuel cost but also it leads to reducing emissions per unit energy produced as the mass of the fuel required to produce energy is reduced. The aim of the current study is to optimize a natural gas SI engine employing cooled EGR in high inlet mixture pressure condition in terms of compression ratio and start of combustion timing in order to obtain the lowest fuel consumption accompanied with high power and low emissions. It is known that computer simulation is a major tool that can be used for design and optimization in order to save efforts, time and money. 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. 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 [7]. 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. Fig. 1. Schematic of the two-zone combustion modeling.

4 A. Ibrahim, S. Bari / Fuel 87 (2008) The flow rates in both the intake and exhaust processes were determined from quasi-steady one-dimensional compressible flow rate equations [7]: vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi _m ¼ C 1=c " # da R p p ffiffiffiffiffiffiffiffi o p t 2c RT o p o c 1 1 p c 1 u c t t ð1þ p o where _m is the mass flow rate through intake and exhaust valves, C d is the discharge coefficient (assumed to be 0.7), A R is a reference area which was selected to be equivalent to the curtain area as suggested by Heywood [7] (A R = pd v l v (t), where d v is valve diameter, l v (t) is valve lift as a function of time (or crank angle)), T o and p o are stagnation temperature and pressure upstream of the valve respectively, p t 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, p o is the intake manifold pressure, and p t is the cylinder pressure. For flow out of the cylinder through an exhaust valve, p o is the cylinder pressure, and p t is the exhaust pressure. When the flow through the valve is choked, i.e. c c 1, the mass flow rate is calculated from the following equation [7]: sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi _m ¼ C cþ1 c 1 da R p p ffiffiffiffiffiffiffiffi o 2 c ð2þ RT o c þ 1 p t p o 6 2 cþ1 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 The combustion process The following assumptions are assumed during combustion: The flame front thickness is assumed to be negligible. The cylinder pressure is assumed to be the same in the burned and unburned zones. Only the convective heat transfer mode, between the cylinder contents and the cylinder wall, is considered. The heat transfer between the two zones is neglected. For the burned zone, ten species (CO 2,H 2 O, CO, N 2, O 2, OH, NO, H, O, and H 2 ) are considered in chemical equilibrium during combustion and expansion. 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 [8] 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 cycle 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 [9] was used to solve the differential equations to determine the in-cylinder pressure and temperature The burning rate The S-shaped mass fraction burned profile, the Wiebe function, was used to determine the burning rate: " X b ¼ 1 exp a h h # mþ1 o ð3þ Dh where h is the crank angle, h 0 is the crank angle at the start of combustion, Dh is the total combustion duration (from X b =0toX b = 1), and a and m are adjustable parameters which fix the shape of the curve. Actual mass fraction burned curves have been fitted with a 5, and m 2as suggested by Heywood [7]. 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 [7]. Thus, referring to Eq. (3), putting X b = 0.5 at h = 370 enables h o to be determined at a specified combustion duration. In the current study, all results were obtained at the MBT spark timing condition The combustion angle A turbulent flame propagation model developed by Tabaczynski and coworkers [10] was used by Hires and coworkers [11] to obtain explicit relations for the flame development angle, Dh d, and the rapid burning angle, Dh b, as function of engine design and operating variables: 2=3 Dh d ¼ CðS p mþ 1=3 L ð4þ S l 10=9 2=3 Dh b ¼ C 0 B qi L ðs q p m Þ 1=3 L i u S ð5þ l where m is the kinematic viscosity, L is the distance between cylinder head and piston, S l is the laminar burning speed, S p 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

5 1828 A. Ibrahim, S. Bari / Fuel 87 (2008) the superscript (*) denotes the value at cylinder conditions, where X b = 0.5. C and C 0 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. [12]: a T u p bð3:4259x 2 S l ¼ S l0 EGR :6993x EGR þ 1:002Þ ð6þ S l0 ¼ 177:43U 3 þ 340:77U 2 123:66U 0:2297 ð7þ a ¼ 5:75U 2 12:15U þ 7:98 ð8þ b ¼ 0:925U 2 þ 2U 1:473 ð9þ where S l0 is reference burning velocity, cm/s, T u is the unburned mixture temperature, K, p is cylinder pressure, kpa, a and b are fitting coefficients, x EGR is 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 , 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 [12]. Since the dynamic viscosity of hydrocarbon-air combustion products differs little from that of air as it was demonstrated by Heywood [7], the cylinder content dynamic viscosity could be expressed using the air dynamic viscosity correlation which has the following form [7]: l ¼ 3: T 0:7 ð10þ where l is dynamic viscosity in kg/ms, and T is temperature in K. Hence, the kinematic viscosity can be determined using the well known correlation: m ¼ l q. Both Eqs. (4) and (5) are used in the present model in order to calculate the combustion duration (Dh = Dh d + Dh b ) at different operating conditions. The combustion duration is then used to determine the burned mass fraction using the Wiebe function NO formation kinetic model The extended Zeldovich mechanism [7] is used to determine the rate of change of NO mole fraction during combustion and expansion processes as follows: dz NO dt where ¼ 2r 1ð1 ðz NO =Z NO;e Þ 2 Þ 1 þðz NO =Z NO;e Þr 1 =ðr 2 þ r 3 Þ ð11þ r 1 ¼ k þ p 1 Z O;e Z N RT 2;e ð12þ b r 2 ¼ k p 2 Z NO;e Z O;e ð13þ RT b r 3 ¼ k p 3 Z NO;e Z H;e ð14þ RT b where Z is mole fraction, p is the cylinder pressure, R is the gas constant, and T b is the burned gas temperature. The subscript e refers to equilibrium value. The rate constants (k), in units of m 3 /kmol s, were calculated from Ref. [7] Knocking model Engine knock is an abnormal combustion phenomenon which can cause engine damage. The most accepted theory that explains engine knock is the auto-ignition theory [7]. The auto-ignition theory states that when the fuel air mixture in the end gas region ahead of the flame front is compressed to sufficiently high pressure and temperature, the fuel oxidation process can occur in parts or in the entire end gas region. This releases the chemical energy in the end gas region at extremely high rates resulting in high local pressures. The non-uniform pressure distribution inside the combustion chamber causes pressure waves or shock waves to propagate across the chamber causing noise which is known as knock [7]. There are two types of auto-ignition models that are used to simulate engine knocking. These are the detailed chemical kinetic models and the empirical delay time models. The chemical kinetic models describe the chemical mechanisms which govern the hydrocarbon oxidation process in the end gas region during the auto-ignition time. The auto-ignition chemical kinetic process is complex and it may be described by thousands of elementary reaction steps and species. The lack of knowledge of the elementary reaction steps and rate coefficients in addition to the long computational time required to analyze thousands of equations makes empirical delay time models more practical for engine designers [13]. The empirical delay time models usually correlate a delay time equation, which is usually in the form of Arrhenius equation, to experimental data in order to determine a simple empirical correlation for the ignition delay time. The ignition delay time represents the time required for the unburned mixture to establish necessary radicals to start the auto-ignition process. In the current study, an empirical delay time knocking model was incorporated to the combustion model in order to predict engine knocking and the crank angle at which knocking might occur. This model was developed based on the auto-ignition model developed by Soylu and Van Gerpen [13] for natural gas engines. They modeled the ignition delay time by using an equation, which was in the form of Arrhenius equation: s ¼ x 1 p x 2 expðx 3 E u =T u Þ ð15þ where s is the ignition delay time, p is the cylinder pressure, T u is the unburned mixture temperature, E u is the temperature coefficient (E u = 7000), whereas x 1, x 2, and x 3 are experimentally determined constants. Soylu and Van Gerpen [13] used the method which was proposed by Livengood and Wu [14] in order to determine the crank angle (or time) at which knocking occurs using the ignition delay correlation. This method is known as the knock Integral method, and has the following form:

6 Z 1=sdt ¼ 1 ð16þ The integration is performed from the intake valve closing time to the knocking time. The substitution of Eq. (15) to Eq. (16) leads to the following equation: Z dt=x 1 pðtþ x 2 expðx 3 E u =T u ðtþþ ¼ 1 ð17þ A. Ibrahim, S. Bari / Fuel 87 (2008) Results and discussion The presented model was used to predict the performance of a 507 cc single cylinder Ricardo E6 engine. Table 2 shows the Ricardo engine specifications Model validation In order to determine the experimental constants (x 1, x 2, and x 3 ), Eq. (17) was fitted to experimental data collected from natural SI turbocharged engine at different operating conditions. The in-cylinder pressure was measured at various operating conditions and the crank angle at which knock occurred was determined using the in-cylinder pressure data. The unburned mixture temperature was calculated using a zero-dimensional thermodynamic model. Then Eq. (17) was fitted to the experimental data and the values of the experimental constants that gave best fit to the experimental results were determined. A mathematical optimization technique known as the Steepest-Ascent method was used to optimize the auto-ignition correlation in order to determine the experimental constants. Soylu and Van Gerpen [13] demonstrated that the optimization of the knock model for 16 different operating conditions showed that the model could be simplified by fixing the value of x 1 to be and the value of x 2 to be The remaining parameter, x 3, varied from to The values of x 3 were correlated to the propane ratio in natural gas fuel, PR, and the equivalence ratio, U, as shown in Eq. (18). Although this knocking model was validated using some experimental data set, the authors concluded that this model can be used with broader generality. Further details about this model are found in Refs. [13,15,16]. x 3 ¼ð 0:575 þ 10:058PR 54:053PR 2 ÞU þ 1:456 8:703PR þ 43:615PR 2 ð18þ In the current study, the integration of Eq. (17) was performed using the calculated values of both the cylinder pressure and the unburned mixture temperature. The integration time starts from the inlet valve closing time and continues during the combustion period. If the integration is equal to unity, knocking is predicted and the corresponding crank angle is registered. Fig. 2 shows the p V diagram at inlet pressure of 98 kpa (wide open throttle, WOT, condition), stoichiometric mixture, engine speed of 2000 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 power. The engine friction power was determined experimentally using the motoring test and then it was used to calculate the engine brake power and consequently the torque. The engine torque was calculated at different engine speeds and compared with experimentally determined torque obtained by Mustafi and coworkers [17] for the engine with the same specifications as in Table 2. Fig. 3 shows that there is a very good agreement between the predicted and experimental torque, which indicates that the model has been well constructed Specifying the inlet conditions In a previous study [18], the effect of EGR on engine performance was investigated at an engine compression ratio of 8 and under several inlet conditions. The results indicated the benefits of using cooled EGR at higher inlet pressure conditions. It was found that the use of cooled EGR at an inlet mixture dilution rate of 20% and high inlet pressure of 250 kpa reduced NO emission by 80% in addition to reducing engine fuel consumption by 19 27% compared to non diluted stoichiometric mixture at an inlet pressure of 98 kpa. Thus, in order to benefit from the positive effect of employing cooled EGR at high inlet pressure, Table 2 Engine specifications Number of cylinders 1 Bore, mm 76.2 Stroke, mm Capacity, cc 507 Maximum speed, rpm 3000 Max. cylinder pressure, bar 150 Inlet valve opens, deg BTDC 9 Inlet valve closes, deg ABDC 34 Exhaust valve opens, deg BBDC 43 Exhaust valve closes, deg ATDC 8 Fig. 2. P V diagram at inlet pressure of 98 kpa, stoichiometric mixture, r c = 8, MBT spark timing, and 2000 rpm.

7 1830 A. Ibrahim, S. Bari / Fuel 87 (2008) increased to the minimum percentage of dilution that could prevent knocking at higher compression ratio as shown in Fig. 4. In the current study, all the investigations which were done at different compression ratio and speed conditions employed different amount of EGR dilution as indicated in Fig Variations of brake power and combustion rate at different compression ratio and speed conditions Fig. 3. A comparison between modeling and experimental [17] results for Ricardo natural gas engine at WOT, stoichiometric fuel air mixture, r c = 8, and MBT spark timing operation. all the current investigations on adding EGR to the inlet mixture at different engine compression ratio conditions were made at inlet pressure and temperature of 200 kpa and 333 K, respectively. In all investigations, EGR was added to a stoichiometric fuel air mixture and the percentage of exhaust gas recirculation (EGR) was calculated as a percentage from the total mass of inlet mixture Minimum EGR dilution requirements Fig. 4 shows the percentage of EGR dilution used at different compression ratio and speed conditions. The minimum value of EGR in the inlet mixture was kept at 20% in order to sustain low NO emission. However, knocking was predicted at higher compression ratio and speed conditions when EGR was equal to 20%. Therefore, EGR was Fig. 5 shows the effect of engine compression ratio variations on brake power at inlet pressure and temperature of 200 kpa and 333 K respectively and different speed conditions. The modest increase in engine brake power which occurred at lower compression ratio and lower EGR dilution rate was due to the increase of the expansion work with the increase of compression ratio. However, the increased mass of EGR which was added at constant inlet pressure in order to avoid engine knocking replaced some part of the fresh mixture and led to a decrease in engine brake power at higher compression ratio. In addition, the excessive dilution rate used at higher engine compression ratio increased the combustion duration significantly as shown in Fig. 6 and consequently slowed down the combustion rate as indicated in Fig. 7. When the compression ratio increased from 8 to 10 at constant EGR dilution of 20% and 3000 rpm, the combustion duration slightly decreased due to the increase of compression ratio which led to an increase in the unburned mixture density and the burning rate. On the other hand, when the compression ratio increased from 10 to 13 at 3000 rpm while EGR dilution increased from 20% to 30% as indicated in Fig. 4, the combustion duration increased from about 117 to 170 due to the increase of EGR dilution which reduced the in-cylinder oxygen concentration and the burning rate. Fig. 4. The percentage of EGR dilution used at inlet conditions of 200 kpa and 333 K and different speed and compression ratio conditions. Fig. 5. The effect of engine compression ratio variations on engine brake power at inlet pressure and temperature of 200 kpa and 333 K, respectively.

8 A. Ibrahim, S. Bari / Fuel 87 (2008) Fig. 6. The variations of combustion duration with compression ratio and speed at inlet conditions of 200 kpa and 333 K. Fig. 8. Start of combustion angle variations with engine compression ratio and speed at inlet conditions of 200 kpa and 333 K, and different EGR dilution rates. Both in-cylinder pressure and temperature are important variables which affect the in-cylinder mechanical and thermal stresses in addition to NO emission. The variations of both the in-cylinder pressure and temperature were studied at different compression ratio and EGR dilution rates. At a lower speed of 1000 rpm, both the maximum cylinder pressure and the maximum burned gas temperature increased with the increase of compression ratio as shown in Figs. 9 and 10, respectively. The increase of the maximum cylinder gas temperature led to an increase in NO emission by about 24% as shown in Fig. 11. At a higher speed of 3000 rpm, while the maximum cylinder pressure increased due to the increase of compression ratio, the increase of EGR dilution, which was required to avoid knocking as shown in Fig. 4, shifted the crank angle at which the maximum cylinder occurs towards the top dead centre as shown in Fig. 12. However, the maximum burned gas temperature decreased due to the increase of Fig. 7. The change of the burned mass fraction during combustion at inlet conditions of 200 kpa and 333 K, engine speed of 3000 rpm, and different compression ratio and EGR dilution conditions. The increase in combustion duration would reduce engine torque due to late combustion in the expansion stroke; therefore, the start of combustion angle was significantly advanced in order to obtain the maximum brake torque as shown in Fig. 8. The start of combustion timing was advanced from about 50 btdc to 77 btdc when EGR dilution was increased from 20% to 30% at 3000 rpm In-cylinder pressure and temperature and their effect on NO emission Fig. 9. The variation of in-cylinder pressure with crank angle at engine speed of 1000 rpm, inlet conditions of 200 kpa and 333 K, 20%EGR, and different compression ratio values.

9 1832 A. Ibrahim, S. Bari / Fuel 87 (2008) Fig. 10. The variation of in-cylinder temperature with crank angle during combustion at engine speed of 1000 rpm, inlet conditions of 200 kpa and 333 kpa, and different compression ratio values. Fig. 12. The variation of in-cylinder pressure with crank angle at engine speed of 3000 rpm and inlet conditions of 200 kpa and 333 K, and different compression ratio and EGR dilution rate conditions. Fig. 11. NO emission variations with engine compression ratio and speed at inlet mixture conditions of 200 kpa and 333 K, and different EGR dilution rates. Fig. 13. The variation of in-cylinder temperature with crank angle during combustion at engine speed of 3000 rpm, inlet conditions of 200 kpa and 333 K, and different compression ratio and EGR dilution conditions. EGR dilution as shown in Fig. 13. This decrease in the maximum burned gas temperature decreased the NO emission by about 28% as shown in Fig. 11. This indicates the strong dependence of NO emission on the maximum cylinder burned gas temperature. It also indicates the potential of EGR on reducing the maximum burned gas temperature and consequently NO emission at high compression ratio. Fig. 11 shows the NO emission variation with engine compression ratio at different engine speeds with inlet mixture conditions of 200 kpa and 333 K. Generally, NO emission change trend followed the maximum burned gas temperature change trend at each speed. The maximum burned gas temperature was affected by both the value of compression ratio and the amount of dilution. When EGR dilution rate was increased above 20% at higher compression ratio (Fig. 4) to avoid knocking, NO emission started to decrease at all speeds (Fig. 11). However, the high EGR dilution condition (30%) which was used at engine speed of 3000 rpm and compression ratio of 13 increased the start of combustion angle excessively as shown in Fig. 8 and led to an increase in the maximum burned gas temperature and consequently an increase in NO emission compared to using EGR at 27% dilution rate and compression ratio of 12 at the same engine speed. In order to make sure that the maximum cylinder pressure was below the allowable limit, the maximum cylinder pressure was monitored and indicated in Fig. 14. The increase of maximum cylinder pressure with engine

10 A. Ibrahim, S. Bari / Fuel 87 (2008) closing time. Fig. 15 shows the variations of the average exhaust temperature with engine compression ratio. The exhaust gas temperature decreases with the increase of compression ratio due to the increase of expansion ratio, which agrees well with other previous studies described by Heywood [7]. 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 Optimizing the compression ratio for minimum fuel consumption Fig. 14. Maximum cylinder pressure variations with engine compression ratio and speed at inlet conditions of 200 kpa and 333 K, and different EGR dilution rates. compression ratio was modest and it did not exceed the maximum pressure limit specified by the engine manufacturer Exhaust gas temperature variations The exhaust temperature was calculated as an enthalpyaveraged temperature using the following integration as suggested by Heywood [7]: Z Z T ex ¼ _mc p T dt= _mc p dt ð19þ where _m is the mass flow rate exiting through the exhaust valve, c p 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 Fig. 16 shows the change of brake specific fuel consumption with engine compression ratio. The excessive EGR dilution rate (up to 30%) which was used at 3000 rpm and higher compression ratio, in order to prevent engine knocking, slowed down the burning rate significantly and led to a significant increase in fuel consumption as most of the fuel was burned away from the top dead centre. The value of the compression ratio at which the minimum fuel consumption occurs varies with engine speed as shown in Fig. 16 and the value of EGR dilution which is used at the corresponding optimum compression ratio varies from 21% to 23% as it can be deduced from Fig. 4. Natural gas can be employed in diesel engines for several applications such as power generation and cogeneration applications which usually require an engine speed of 1500 rpm. For this operating speed, the optimum compression ratio that would give the minimum fuel consumption was found to be about 12. The corresponding minimum fuel consumption was found to be about 200 g/kw h with an equivalent brake thermal efficiency of about 40%. Comparable results were achieved experimentally by Nellen and Fig. 15. Average exhaust gas temperature variations with engine compression ratio and speed at inlet conditions of 200 kpa and 333 K, and different EGR dilution rates. Fig. 16. Brake specific fuel consumption variations with engine compression ratio and speed at inlet mixture conditions of 200 kpa and 333 K, and different EGR dilution rates.

11 1834 A. Ibrahim, S. Bari / Fuel 87 (2008) Boulouchos [19] who optimized a turbocharged diesel engine to work on natural gas fuel for cogeneration application at 1500 rpm. Nellen and Boulouchos achieved a high brake thermal efficiency which ranged from about 40% to 42% depending on engine load at a compression ratio of 12. They used a stoichiometric mixture with cooled EGR at a rate which ranged from 20% to 25%. 4. Conclusions A two-zone combustion model was developed in order to optimize a natural gas SI engine employing a stoichiometric mixture with EGR dilution at high inlet pressure condition in order to obtain the lowest fuel consumption accompanied with high power and low emissions. The effect of compression ratio on engine performance was studied. The following conclusions have been obtained: The use of cooled EGR with a dilution rate ranged from 20% to 30% depending on engine speed suppressed engine knocking and allowed using high inlet pressure condition (200 kpa) in a relatively high compression ratio values (up to 13). The use of high compression ratio condition (up to 13) at high speed of 3000 rpm demanded using excessive EGR dilution (up to 30%) to avoid knocking. The excessive dilution slowed the burning rate significantly and led to a significant increase in fuel consumption. The value of the compression ratio at which the minimum fuel consumption occurs varies with engine speed and the value of EGR dilution which is used at the corresponding optimum compression ratio varies from about 21% to 23% at inlet conditions of 200 kpa and 333 K. A minimum fuel consumption of about 200 g/kwh was achieved at engine speed of 1500 rpm, inlet conditions of 200 kpa and 333 K, and a compression ratio of about 12. Cooled EGR has the potential on reducing the maximum burned gas temperature and consequently NO emission at high compression ratio conditions. References [1] Baldassarri LT, Battistelli CL, Conti L, Crebelli R, De Berardis B, Iamiceli AL, et al. Evaluation of emission toxicity of urban bus engines: compressed natural gas and comparison with liquid fuels. Sci Total Environ 2006;355(1 3): [2] The Environmental Protection Agency (EPA), USA. Clean alternative fuels: compressed natural gas. fact sheet, oms/consumer/fuels/altfuels/420f00033.pdf published in 2002, [accessed ]. [3] Catania AE, D Ambrosio S, Mittica A., Spessa E. Experimental investigation of fuel consumption and exhaust emissions of a 16V pent-roof engine fueled by gasoline and CNG. Society of Automotive Engineers 2001, SAE paper no [4] The Engine Manufacturers Association, USA. The use of exhaust gas recirculation systems in stationary natural gas engines. published in 2004, [accessed ]. [5] Rabl A. Environmental benefits of natural gas for buses. J Transport Res D 2002;7: [6] Environmental Protection Agency Handbook. Control Technologies for Hazardous Air Pollutants. EPA 625/ ; [7] Heywood JB. Internal combustion engine fundamentals. New York: McGraw-Hill Book Company; [8] Ferguson CR. Internal combustion engines. Wiley & Sons; [9] Caton JA. Comparisons of instructional and complete versions of thermodynamic engine cycle simulations for SI engines. Int J Mech Eng Edu 2000;29(4). [10] Tabaczynski RJ, Ferguson, CR, Radhakrishnan, K. A turbulent entrainment model for spark-ignition engine combustion. SAE Transactions; 1977;86, SAE paper no [11] Hires SD, Tabaczynski RJ, Novak JM. The prediction of ignition delay and combustion intervals for a homogeneous charge spark ignition engine. SAE Transactions, 1978;87, SAE paper no [12] Liao SY, Jiang DM, Huang ZH, Cheng Q, Gao J, Hu Y. Approximation of flammability region for natural gas air-diluent mixture. J Hazard Mater 2005;A125:23 8. [13] Soylu S, Gerpen JV. Development of an auto-ignition sub-model for natural gas engines. J Fuel 2003;82: [14] Livengood JC, Wu PC. Correlation of auto-ignition phenomenon in internal combustion engines and rapid compression machines. In: Fifth international symposium on combustion; 1955: [15] Soylu S. Prediction of knock limited operating conditions of a natural gas engine. J Energy Conversion Manage 2005;46: [16] Soylu, S. Auto-ignition modeling of natural gas for engine modeling programs: an experimental and modeling study. PhD thesis, 2001, Iowa State University, USA. [17] Mustafi NN, Miraglia YC, Raine RR, Bansal PK, Elder ST. Spark ignition engine performance with power gas fuel: a comparison with gasoline and natural gas. J Fuel 2006;85: [18] Ibrahim A, Bari S, Bruno F. A Study on EGR utilization in natural gas SI engines using a two-zone combustion model. Soc Automotive Eng; 2007, SAE paper no [19] Nellen, C, Boulouchos, K. Natural gas engines for cogeneration: highest efficiency and near-zero-emissions through turbocharging, EGR and 3-way catalytic converter. Soc Automotive Eng; 2000, SAE paper no