Gasoline Mixture Combustion in Direct Fuel Injection SI Two-Stroke Engine

Transcription

1 Archivum Combustionis ol. 30 (00) no. 4 Gasoline Mixture Combustion in Direct Fuel Injection SI Two-Stroke Engine W. Mitianiec, Ł. Rodak, M. Forma Cracow University of Technology, Al. Jana Pawla II 37, Krakow, Poland tel: ; Fax: ; wmitanie@usk.pk.edu.pl, lrodak@pk.edu.pl, marianforma@wp.pl Process of gasoline mixture combustion in two-stroke engines is much more complicated than in four stroke engines. In order to decrease fuel consumption and exhaust gas emission, particularly hydrocarbons, two-stroke engines should have direct fuel injection systems. The paper presents results of both simulation and experimental tests on a small power gasoline two-stroke engine. The special high pressure injection system was elaborated in Cracow University of Technology for engine with capacity 5 cm 3. The most important factor in such engine is small time for fuel evaporation and obtaining a homogeneous mixture in the combustion chamber. The gas flow in the cylinder (tumble and swirl) causes a certain deviation of fuel stream during injection and therefore location of the injector has big significance on the combustion process. The paper shows results of fuel distribution in the cylinder for different crank angles obtained from spatial simulation (KIA3) and propagation of the flame during combustion process. Experimental work was carried out on the engine in order to determine influence of injection parameters on engine work and combustion process. On the basis of pressure traces at different loads and rotational speeds heat release rates were calculated and ibe functions were found. The paper presents the engine experimental stand and results of decreasing of important exhaust gases components at the same engine working parameters as in carburetted engine.. Problems of direct fuel injection in spark ignition two-stroke engines Combustion speed can be increased by higher turbulence of the charge, which is caused by special design of the combustion chamber. The results of simulation were inputs for designing of the experimental injection system. The engine was fuelled by conventional automotive direct fuel injection FSI and was not completely suitable for such small engine. The standard carburetted engine was tested on the laboratory stand to achieve working parameters as a data, which have been compared with the data obtained from engine equipped with direct fuel injection system. The engine with direct fuel injection indicated much smaller specific fuel consumption and very low exhaust gas emission, for example

2 378 W. MITIANIEC, Ł. RODAK, M. FORMA HC and CO in comparison to the carburetted version. Applying of direct injection of fuel in a two-stroke engine decreases HC emission several times. In the carburetted engine HC emission is much higher than in four-stroke engine. Start of fuel injection begins after closing of the exhaust port and therefore there are any escapes of fuel to the outflow port during scavenge process. Turbulence and high gas motion in the form of tumble in the combustion chamber cause better mixing of the fuel with the air and quicker vaporization of fuel droplets despite of short time between beginning of fuel injection and ignition. The proper location of the injector was found after many CFD simulations by using KIA [, 3] and by using GT- Power program (0D and D). The industrial two-stroke engine Robin EC was equipped with the standard automotive injector from olkswagen FSI system working with high injection pressure 55 bar. The control electronic system was prepared in CUT in Labview environment. The combustion process is better controlled with direct fuel injection than in carburetted version. Besides of HC emission also amount of carbon monoxide can be reduced in comparison with standard engine. Reduction of CO is higher at higher rotational speed quite different as in carburetted version. Obviously, by reduction of fuel escaping to the exhaust port during scavenge process, the break specific fuel consumption (bsfc) for direct fuel injection version is decreased about 30% and is near to specific fuel consumption in SI four stroke engine, however without complicated control system as in automotive engines.. Goal and scope of work The research work carried out by authors on small capacity two-stroke engine with direct fuel injection is one of the first in Poland. In the world only one industrial application of direct fuel injection in outboards high power multi-cylinder two-stroke engines from Mercury Marine (in marine boats) is known. The main advantages of applying of such engines are compactness, high unit power, lightness and possibility of working in different position. There are many small two-wheel and three wheel vehicles driven by carburetted two-stroke engines in the world. Limitation of exhaust gas emission and fuel consumption of such engines is a main object of recent research works concerning such engines. The goal of the author s work was to define the injector position in the combustion chamber with regard of the air-fuel mixture motion and orientation of fuel spray and determination of these factors on the combustion process and exhaust gas emission. Such relations could be simply done by CFD simulations by different injector orientation in the combustion chamber. For that case KIA3 program was used as common used tools for solving combustion and fuel injection problems in internal combustion engines. The fundamental goal of the work was determination of following aspects: decreasing of specific fuel consumption and pollutants emission, influence of the injector location in respect to possible ignition of the air-fuel mixture

3 Gasoline Mixture Combustion in Direct Fuel Injection SI Two-Stroke Engine 379 and influence of gas motion on mixture distribution by using CFD simulation, comparison of working parameters of DFI and carburettor two-stroke engine, comparison of exhaust gas emission in the both engine versions, determination of heat release rate on experimental pressure measurements and ibe function for zero-dimensional models of two-stroke engines with direct fuel injection, because such functions are not met in literature, influence of fuel injection pressure on working parameters and exhaust gas emission, The work objects given above were realized by simulation tests with using of the CFD KIA3 program with source code and GT-Power software, by the experimental work on 5 cm 3 capacity engine Robin EC (Fuji Heavy Industries) and theoretical considerations concerning to the mathematical model of heat release rate in the direct injection two-stroke engine. Most of the work was realized by experimental research on dynamometer stand with the modified engine equipped with the prototype of high pressure injection system. The goal of the measurements carried out on the test stand was to determine: fuel and air mass flow rate, torque, optimal injection pressure, injection timing, volumetric concentration of components of exhaust gases by using of gas analyzers, pressure measurements in the cylinder for analysis of combustion process. The scope of the experimental work contained the following tasks: choice of suitable injector after research of fuel mass flow rate, influence of pressure injection on work engine parameters, influence of injection timing on fuel consumption and exhaust gas emission, determination of load performances for chosen rotational speeds, carrying out of indicating diagram in the cylinder and crankcase. determination of instability of combustion process, analysis of experimental results. 3. Determination of heat release in DFI engine Thermodynamic state of combustion engines is presented widely in literature e.g. by Heywood [9], Blair [4] and Mitianiec [6, 7]. Fluid dynamics of fuel injection process is still under development and was a subject of the work of Feath [7], Ghandi [8], Ikeda [] or Melton [5]. Combustion processes and charge turbulence are described e.g. by Chen and Kim [5], Spalding [0] or Higelin []. The engine cylinder after closing of transfer and exhaust ports is a close thermodynamic system. The First Law of Thermodynamics for such closed system states that heat delivered to the charge dq R increases internal energy du and some has been lost through heat transfer dq c to the cylinder walls and coolant at the same time and work done by the piston dw.

4 380 W. MITIANIEC, Ł. RODAK, M. FORMA d QR dqc dqvap = du + dw () The term d Qvap can be neglected because in spark ignition engines almost whole fuel is in vapor state during combustion process. Increment of internal energy in the close system is a function of increment of temperature. du = mc dt () v The specific heat at constant volume c v is a function of temperature and also of the gas properties. During combustion process temperature and properties of the charge change rapidly. Thus c v is not a constant value and should be determining every time during measurement or calculation. p = mrt (3) R c v = (4) k After differentiation of the state equation of close system at assumption of constant mass of charge m and gas constant R : pd + dp = mrdt dt pd + dp = (5) mr After substitution to the equation of internal energy pd dp d U = (6) k Average work done on the piston during time interval is: After substitution to the main formula (): d W = pd (7) pd + dp d QR dqc = + pd (8) k If the combustion process had not occurred, then the compression or expansion process would have continued in a normal fashion. The polytropic process is taking place with relationship defined by:

5 38 Gasoline Mixture Combustion in Direct Fuel Injection SI Two-Stroke Engine = n p const (9) 0 = + d np d p n n = 0 + d np p d p (0) In imaginary process is one of non-heat addition dq R is zero and the First Law of Thermodynamics could be rewritten to calculate the heat loss dq c by substituting of formula (9) and the following dependence can be written [4]: ( ) + + = n n c k p Q d () On the assumption that the heat loss is the same and continues during the combustion process the following equation on the heat release rate (HRR) can be determined: + = k p p dq n R () Rasseweiler and Withrow [] proposed another simpler form: + = k p p dq n R (3) Specific heat at constant volume c v is a function of temperature T and can be determined on JANAF Tables [3] and can be expressed as: z A z A z A z A z A A c p = (4) where 000 T / z = and coefficients of the polynomial amount as follows: A =-0.97 A =9.775 A 3 =-7.68 A 4 =.466 A 5 =0.0

6 38 W. MITIANIEC, Ł. RODAK, M. FORMA In order to determine temperature in the cylinder during combustion process, the mass of charge should be known. In the calculations the authors took this value from CFD calculations of the whole engine work cycle. The gas constant R changes during combustion, however it is close to the gas constant of the air. On these simple assumptions the gas temperature can be calculated. Polytropic exponent n is obtained from equation (9) by measured pressure values during whole combustion period D a b from the state and state and calculated volumes at these states. Total released heat Q t can be determined by integration of heat release rate at combustion period: Q t = a a dq da R (5) where index signifies state as the beginning of combustion and state signifies the end of combustion. The current burned mass ratio x z is determined as follows: x z a QRda = Q (6) t In the engine theory of combustion engines the ibe [] function is used for determination of combustion process based on approximation of measured pressure on real engines. This function is commonly used in literature [4, ] and research work for analysis of combustion process in piston engines. ibe function in most cases takes the following form: m a a x z = exp a (7) Da b The angle a is the current angle of crankshaft position, a is an angle of crankshaft position, when combustion process begins, on the contrary Da b is total angle of crankshaft rotation during combustion process. The coefficients a and m are determined by continuous change their values until the shape of ibe function reaches measured fuel mass burned ratio. ibe functions is a common used approximation function in definition of heat release rate during combustion process and is regarded as the best approximation function for determination of ratio of burned fuel. It is needed in 0-dimensional model used in computer program simulations for analysis of whole engine system. Determination of coefficients in ibe function for verification of combustion model applied in own ICE models and GT- Power software, used in this work, was needed.

7 Gasoline Mixture Combustion in Direct Fuel Injection SI Two-Stroke Engine CFD simulation of the work of DFI two-stroke engine Till now it is not possible exactly to determine: the propagation of fuel jet, fuel droplets evaporation and thermodynamic parameters in the combustion chamber. In the work CFD simulation had the objective to determine an influence of gas motion on the fuel jet behavior, distribution of the air-fuel mixture and combustion process in the combustion chamber. The main target of such research was to define the injector position and determination of initial injection timing for experimental tests. Simulation of engine work and calculations of whole thermodynamic processes in the direct fuel injection two-stroke engine were carried out for different engine rotational speeds. Result of CFD simulation concerns to engine work at 500 rpm with fuel dose 5 mg per one cycle. Simulation was carried out by CFD program Kiva3v. The initial parameters (pressure and temperature) in the crankcase and transfer ports were taken from the experiment on the real carburetted engine. At outlet port the pressure was assumed as constant ambient pressure equal bar. The main problem in direct fuel injection two stroke engine is deflection of injected fuel stream by the air stream during compression process and it influences on the local air fuel ratio. Loop scavenge process and squeezing effects cause so called tumble motion of the gas inside the combustion chamber. elocity vectors of the gas in the hemisphere combustion chamber of Robin CE engine is shown in Fig. at 4 deg CA BTDC. The gas motion strongly influences on the injected fuel jet from the injector, which was mounted on the exhaust side. cm from the central axis of the cylinder with inclination 0 deg to that axis. Fig. Charge velocity vectors in the cylinder at 4 deg CA BTDC 4.. Injection process Evaporation of the gasoline fuel is very rapid after small period from the beginning of fuel injection and depends on value of pressure injection and temperature of the charge. Because of high injection pressure about 6 MPa the fuel jet contains small fuel droplets, which quickly evaporate as a result of high temperature about 600 K. The calculated mean Sauter diameter D 3 amounted 0 mm. The gas motion deflects both liquid fuel jet and

8 384 W. MITIANIEC, Ł. RODAK, M. FORMA fuel vapors. For combustion process it is important to achieve a mixture of air and fuel vapors near stoichiometric ratio. Mass ratio of gasoline vapors in the charge is shown in Fig. a and b at 80 and 6 deg CA BTDC, respectively and before the ignition in Fig. 3a and 3b at 4 and 7 deg CA BTDC. Simulation shows a deflection of vapor stream from the geometrical direction of the injector. The liquid and the vapor fuel jets are rotated in the combustion chamber in direction to the spark plug and thus the correct ignition of the mixture depends on the advance angle of the injection process. a) b) Fig. Distribution of gasoline vapours at a) 80 deg and b) 6 deg CA BTDC a) b) Fig. 3 Distribution of gasoline vapours at a) 4 deg and b) 7 deg CA BTDC 4.. Combustion process Theoretically development of the flame during combustion process should be spherical. However, as a result of gas motion and local distribution of the air excess ratio the real burned zone has the shape of a distorted sphere moved to the side of the exhaust port. The ignition was initiated by additional energy given by spark plug in the central axis near the

9 Gasoline Mixture Combustion in Direct Fuel Injection SI Two-Stroke Engine 385 cylinder head wall. Combustion model in KIA3 takes into account chemical species, such as C8H7, O, O, N, N, CO, CO, H, H, NO, HO and OH. The program calculates chemical species by solving of 4 chemical kinetics reactions with Zeldovich mechanism and dissociation reactions. Fig. 4 Charge temperature at 5 deg CA BTDC and 0 deg CA ATDC a) b) Fig. 5 Formation of hydroxyl radicals at 5 deg CA BTDC and 0 deg CA ATDC Propagation of the burned mixture by representation of temperature is shown in Fig. 4a and 4b at 5 deg CA BTDC and 0 deg CA ATDC, respectively. Simulation showed that combustion process is incomplete. Flame propagation has not a spherical shape besides the spark plug is located in the central cylinder axis. It is caused by charge motion, which deflects the flame shape and local air-fuel ratio. On the cylinder opposite side of the exhaust port the mixture is too rich and does not take part in the combustion process. Combustion of the mixture begins after formation radicals group such as HC, OH, H, O and others. Figures 5a and 5b present the mass distribution of hydroxyl radicals OH in the combustion chamber at 5 deg CA BTDC and 0 deg CA ATDC, respectively. The most concentration of OH radicals takes place in the centre of burned zone at temperature about 500 K.

10 386 W. MITIANIEC, Ł. RODAK, M. FORMA 4.3. Discussion of results Both simulation of injection process and combustion process showed a big influence of gas motion in a tumble form on the fuel propagation and deviation of the combustion flame. The proper position of the injector should enable a filling of the combustion chamber with mixture near stoichiometric value and at the moment of ignition of a combustible mixture should reach the spark plug. By applying of direct fuel injection the charge in the combustion chamber is not homogeneous and is rather close to a stratified charge. Total air excess coefficient l should be higher as.0 in order to consume oxygen by hydrocarbons during combustion process. Besides high amount of fresh air in some regions, there is no possibility to burn whole amount of fuel in regions, where the local l is below.0. Simulation was carried out for different positions of the injector in the combustion chamber. In the paper were shown the results for best position for the engine Robin. Location of the injector on the exhaust port side enables near uniform propagation of fuel vapors and droplets to the spark plug and filling the combustion chamber by combustible mixture. Location of the injector on the opposite side causes a flow of fuel to the side of the exhaust port and a significant increase of enrichment of mixture in this area. Combustion process is initiated by radicals such OH, HC, H and others and highest temperature about 500 K takes place in the centre of the burning zone, which causes formation of NOx. The flame does not reach areas, where is no combustible mixture. Simulation research showed incomplete combustion process resulting of local air-fuel ratios. In simple models of combustion processes in the piston engines the charge turbulence is not taken into account and turbulent combustion velocity is only estimated. 5. Experimental stand The simplified diagram of engine stand is presented in Fig.6 showing the main units of whole injection and flow system. Full electronic control of engine work by programmer of eddy current dynamometer was fulfilled by software working in Labview environment. Mass flow of fuel was determined by a special measurement unit for small dose of fuel, because industrial system was not so precise. Low pressure pump delivered the fuel to the high pressure pump and pressure was adjusted by the industrial multi-range pressure controller. alue of injection pressure was set from 35 to 65 bars. Oscillation of pressure was reduced by applying of high pressure vessel with volume about 0.5 l. Excess of fuel returned to low pressure system between the fuel filter and the low pressure pump. Fuel pressure was observed from the manometer. The air mass flow rate was determined by flow meter (heated wire). The software in the computer enabled change of the fuel dose by change of the injection time. In dependence of engine speeds the beginning of fuel injection changed from 00 to 60 deg BTC of piston position. All control signals were determined in relation to the

11 Gasoline Mixture Combustion in Direct Fuel Injection SI Two-Stroke Engine 387 crank position in BDC by using the encoder directly mounted on the end of the crankshaft. For determination of volumetric concentration of main exhaust gas components a typical exhaust analyzer Arcon with NDIR system was used, but it measured also concentration of NO. In standard engine start of ignition was set 0 deg BTDC, but in the modified engine the ignition can be changed. 6. Results of experimental test The experimental tests were carried out both for standard carburetted engine and for high pressure direct fuel injection engine. During experimental investigations the influence of injection timing and fuel pressure level for different loads and engine rotational speeds was registered. Preliminary results of the tests showed advantages and disadvantages of applied feeding system. Fig. 6 Diagram of experimental stand with DFI two-stroke engine

12 388 W. MITIANIEC, Ł. RODAK, M. FORMA The tests concerns also measurement of cylinder pressure for different loads and rotational speeds. For that reason the piezo-optical sensor from Optrand company was used, which was directly mounted in the wall of cylinder head. Another Kistler piezo-resistive sensor measured pressure inside engine crankcase. All electric signals from sensors were transformed by amplifiers to the computer in the function of crank angle (encoder) and variations of pressure can be performed on the graphs. Figure 7 presents pressure traces in the cylinder and crankcase of two-stroke engine with direct fuel injection at 00% throttle opening for engine speed 3300 rpm. It is observed inequality of the pressure maximum from cycle to cycle. The maximum pressure in the cylinder changes from.3 to.68 MPa. It is caused by existing of the rest of exhaust gases and not complete cleaning of the cylinder by the fresh air during the scavenge process. The rest of exhaust gases stays in the cylinder after closing of the ports as a result of incomplete scavenging by fresh air, when both exhaust and transfer ports are opened. Strong dynamic effect of pressure waves in the exhaust system takes place in two-stroke engines, which stops the outflow of the gases. Fig. 7 Inequality of pressure in engine with direct fuel injection at 750 rpm 50% WOT Therefore the injected fuel does not form a proper mixture composition, which enables good combustion process. Mean temperature of the charge during combustion was calculated on the base of pressure measurement and general formula of gas state: p T = (8) mr It can be noticed in Fig. 8 that the maximum of pressure takes place at 0 deg ATDC of piston position, however ignition takes place at constant value 0 deg BTDC of ignition advance. Combustion process is elongated in comparison to four stroke engines. Maximum pressure a little bit exceeds.5 MPa and is typical value for industrial two-stroke engines and combustion process takes place without knocking. The crankcase pressure variation is shown also in Fig.8, which changes from to 0.35 MPa.

13 Gasoline Mixture Combustion in Direct Fuel Injection SI Two-Stroke Engine 389 Fig. 8 Cylinder and crankcase pressure traces at 3400 rpm and WOT When the transfer ports start to open the pressure in the cylinder is higher than in the crankcase. The exhaust gases from the cylinder flow to the transfer ports causing a rapid increase of pressure in the crankcase. Figure 9 presents variation of pressure of the cylinder charge and temperature, which was calculated on the base of eq. (9) in the range of combustion process. For considered rotational speed 3400 rpm the maximum of temperature takes place about 30 deg ATDC and insignificantly exceeds 000 K. Low value of the maximal temperature is caused by longer time of combustion process and lower heat released rate. Combustion process is retarded by big amount of exhaust gases in the fuel mixture. Fig. 9 Cylinder temperature and pressure during combustion process at 3400 rpm

14 390 W. MITIANIEC, Ł. RODAK, M. FORMA On the base of observation of many pressure traces achieved from the indicated measurements it was not found any abnormal combustion process. In a support of the pressure indicating diagram and formulas ( 7) presented in the section the heat released rate and total heat released were calculated. One example of heat released rate and total heat released in a function of crank angle for two-stroke engine with direct fuel injection is presented in Fig. 0 for engine rotational speed 3400 rpm. Fig. 0 Heat release rate and total heat release at 3400 rpm The burning period amounts 60 deg of crankshaft rotation and maximum of heat release rate amounts 6 J/deg and total heat released reaches value 330 J. ariations of volume specific heat and specific heat ratio are shown in Fig. for the same rotational speed. Combustion process in the DFI two-stroke engine runs with non-monotonically heat released rate and maximal value of HRR takes place at maximum of pressure (0 deg ATDC). At this process the changeable HRR is caused by burning of non-homogenous mixture with different local air excess coefficient l. On these figures indexes and represent the beginning and the end of combustion process. For the simulation of combustion process the ibe function is widely used and in some computer programs the parameters a and m are needed. Approximation functions of the combustion process for direct fuel injection twostroke engines are not met in the literature. Equations (6) and (7) enable to obtain the approximated ibe function for different loads and rotational speeds. It is not possible to find the exact ibe function for the real combustion process. Two examples of mass burned rate and approximated ibe function are presented in Fig. and 3 at 750 rpm and 3400 rpm, respectively.

15 Gasoline Mixture Combustion in Direct Fuel Injection SI Two-Stroke Engine 39 Fig. ariations of specific heat at constant volume and specific heat ratio during combustion process at 3400 rpm Fig. Fuel mass burned ratio and approximated ibe function in combustion process at 750 rpm At lower rotational speed 750 rpm the burning period amounted to 36 deg CA (.8 ms) and is shorter than at higher rotational speed 3400 rpm where the burning period amounted to 55 deg CA (.69 ms). The selected coefficients in ibe function amount as follows: n=750 rpm a=5.9 m=3 n=3400 rpm a=6.0 m=.6

16 39 W. MITIANIEC, Ł. RODAK, M. FORMA Fig. 3 Fuel mass burned ratio and approximated ibe function in combustion process at 3400 rpm At lower rotational speeds of DFI two-stroke engine the burning process is quicker than at higher speeds, where that process is more softly. For carburetted version the measured brake mean effective pressure is slightly higher than for DFI system in lower rotational speeds. Engine mean effective pressure variation is shown in Fig. 4 for both versions. For the sake of applied commercial injector tests of DFI engine were carried out in smaller engine range of rotational speed ( rpm). For standard engine and DFI engine maximum torque takes place at the same rotational speed about 3500 rpm. At the same scavenge timing the volumetric efficiency is the same for both version. In the DFI engine torque value depends on amount of injected fuel, which controls the air-fuel ratio. The mean effective pressure for both engine versions does not exceed 3.75 bar. Fig. 4 Comparison of mean effective pressure for standard carburetted engine and engine with direct fuel injection at WOT

17 Gasoline Mixture Combustion in Direct Fuel Injection SI Two-Stroke Engine 393 By applying of high pressure direct fuel injection, atomizing of fuel is enough for quick evaporation of droplet, despite of shorter time than in four stroke engine. This phenomenon enables full combustion of fuel. The DFI engine with almost the same power as the carburetted engine consumes lower amount of fuel. Fig. 5 presents comparison of specific fuel consumption for both engine versions. It is observed a considerable decrease of bsfc (brutto specific fuel consumption) for DFI engine at whole range of rotational speed. Fig. 5 Comparison of specific fuel consumption for standard carburetted engine and engine with direct fuel injection at WOT Decreasing of fuel consumption reaches almost 35% in comparison to standard version and is close to values of four stroke engines. Injected fuel to the cylinder is fully consumed during combustion process and in comparison to the carburetted engine the mixture does not flow into exhaust port during scavenge process. Comparison of mass emission of hydrocarbons for both versions is shown in Fig.6. Fig. 6 Comparison of hydrocarbon emission in two-stroke engine for carburetted version and direct fuel injection version

18 394 W. MITIANIEC, Ł. RODAK, M. FORMA For carburetted engine HC emission reaches 60 g/kwh and is ten times higher than for DFI engine, where highest value of HC emission amounts 9 g/kwh in lower rotational speed. Lowest value of HC concentration amounted about 00 ppm at 3700 rpm (without catalytic converter) and is near homologation requirements for car engines. Like the hydrocarbon emission DFI engine emits also lower amounts of carbon monoxide, which is caused by better combustion process. Dose of injected fuel enables achieving of different range of air-fuel ratios. For carburetted version the air excess coefficient changed for different rotational speeds in the range l= For DFI version l value was constant and amounted.05 by changing fuel dose for each measurement point. Comparison of CO emission is shown in Fig. 7 both for the carburetted and DFI engine. Fig. 7 Comparison of carbon monoxide emission in exhaust gases in two-stroke engine for carburetted version and direct fuel injection version In the carburetted engine an increase of CO emission with increasing of engine rotational speed takes place in comparison to the DFI engine, where at higher rotational speed the concentration of CO amounted %. Existence of high amount of CO as a result of incomplete fuel combustion is caused by stratification of mixture. Simulation results carried out in KIA program show existence of rich mixture areas near the cylinder wall opposite to the exhaust port. Because of internal exhaust gas recirculation (IEGR) in twostroke engines the emission of NOx is lower than in four stroke engines. However in twostroke engine with direct fuel injection one observes an insignificant increase of volumetric concentration of NO at higher engine speeds. Comparison of NO emission for both engines is presented in Fig. 8. Maximum value of NO emission does not exceed.5 g/kwh (volumetric concentration is below 300 ppm), which is several times lower than in SI four stroke engine. olumetric concentration of CO for both versions oscillated in the range 9 %. Reaching 4% of CO as in four stroke engines is not possible on this stage of investigations, because of higher emission level of CO. For this reason the current goal is increasing of thermal efficiency by achieving of complete combustion process.

19 Gasoline Mixture Combustion in Direct Fuel Injection SI Two-Stroke Engine 395 Fig. 8 Comparison of NO emission in exhaust gases in two-stroke engine for standard version and direct fuel injection version 7. Conclusions and remarks The engine was fitted with high pressure standard automotive injector applied in DFI four-stroke engines, which is not suitable for small capacity two-stroke engine. For this reason the engine work parameters can be improved and exhaust gas emission can be decreased by applying of an injector with smaller fuel mass flow rate. However on the carried out calculations and experimental work some conclusions can be presented.. Direct fuel injection does not assure higher engine power for two-stroke engine. The gasoline engine consumes higher amount of fuel (about 30%) per time unit in comparison to DFI engine, but big amount of the mixture flows into the exhaust system. For DFI engine injection process takes place after closing of the exhaust ports and only unburned fuel is delivered to the outflow system. Thus for the same global air-fuel ratio in the cylinder the engine has almost the same power for both versions, however with direct fuel injection system the engine has lower bsfc. For the constant timing of transfer and exhaust ports the same air is delivered to the crankcase and cylinder. Propagation of the fuel jet in DFI engine depends strongly on the gas motion in the combustion chamber, which is caused by the scavenge process and squeezing effect.. The modified engine almost ten times reduces hydrocarbons emission as an effect of elimination of escaping of fuel mixture during the scavenge process. olumetric concentration of HC changes with rotational speeds and can reach values below 00 ppm. The HC emission in automotive spark ignition port injection four stroke engines without catalytic converter is about.7 g/kwh for l= and for the tested engine HC emission reaches g/kwh.

20 396 W. MITIANIEC, Ł. RODAK, M. FORMA 3. The high pressure direct fuel injection decreases significantly specific fuel consumption above 30% in comparison to carburetted engine, mainly as a result of lower fuel escape to the exhaust port during scavenge process and also as a consequence of charge stratification in the combustion chamber. Higher indicated efficiencies are obtained for stratified mixtures in the piston engines. In the tested engine total efficiency amounted near 8.5% at 3500 rpm. 4. High pressure injection increases velocity of fuel droplets. Gas motion and high velocity of fuel stream takes effect on the breaking of droplets. Smaller fuel droplets have possibility of quicker vaporization and oxidation and for this reason the combustion process is improved. Concentration of CO in exhaust gases is less than in the carburetted engine and is still higher than in the same capacity SI four stroke engines. On the other hand NOx emission (..5 g/kwh) slightly increases with rotational speed in comparison to the carburetted engine and is several time lower than in automotive MPI SI four-stroke, where NOx emission is about 4 g/kwh. 5. On the base of pressure measurements in the cylinder the unknown parameters of the approximate ibe function were matched for some rotational speeds. This approximation function is needed in the most 0-dimensional models of combustion process in the piston engines. The time of fuel combustion increases with rotational speed. Combustion process is determined by engine rotational speed, injected fuel mass, injection timing and charge turbulence in the combustion chamber. Knowing of the real combustion time of given fuel dose and the parameters of ibe function the heat release rate can be calculated in the computer program with 0-dimensional combustion model. References [] Agawral A., Assanis D., Multi-Dimensional Modeling of Natural Gas Ignition Under Compression Ignition Conditions Using Detailed Chemistry, SAE Paper [] Amsden A.A. et al: KIA: A Computer Program for Two- and Three-Dimensional Fluid Flows with Chemical Reactions and Fuel Sprays. Rept. LA-045-MS, 985. [3] Amsden A.A., O Rurke P.J., Butler T.D., KIA-II A Computer Program for Chemically Reactive Flows with Sprays, Los Alamos National Lab., LA-560-MS, 989 [4] Blair G. P., Design and Simulation of Four-Stroke Engines, SAE R-6, Warrendale, 980 [5] Chen Y.S., Kim S.W.: Computation of Turbulent Flows Using an Extended k-e turbulence closure model. NASA CR-7904, 987. [6] Elshnawi M., Teodorczyk A., alidation of detailed reaction mechanisms for simulations of combustion systems with gas injection, Journal of Kones, ol. 9, No -, Warsaw-Gdansk, 00. [7] Faeth G. M., Evaporation and Combustion of Sprays. Prog. Energy Comb. Sci., ol. 9. pp -76, 983. [8] Ghandi J.B.et al.: Investigation of the Fuel Distribution in a Two-Stroke Engine with an Air-Assisted Injector. SAE Paper SAE Int. Congress & Exp., Detroit, 994. [9] Heywood J. B., Internal Combustion Engine Fundamentals, Mc Graw-Hill, 988 [0] Heywood J. B., Combustion and its modeling in Spark Ignition Engines, International Symposium COMODIA 94, Paper C94-P00, 994

21 Gasoline Mixture Combustion in Direct Fuel Injection SI Two-Stroke Engine 397 [] Higelin P., Mounaim-Rouselle C., Pajot O., Robinet C., Moreau B., Flame Propagation in an SI Engine with a Non-Conventional Igniter, The Fourth International Symposium COMODIA 98, Paper C98-P34, 998 [] Ikeda Y., Nakajima T., Kurihara N.: Spray Formation of Air-Assist Injection for Two-Stroke Engine. SAE Paper SAE Int. Congress & Exposition, Detroit, 995 [3] JANAF Thermochemical Tables, National Bureau of Standards Publication NSRDS-NBS37, 97 [4] Kuo T., Reitz R.: Three-dimensional computations of combustion in premixed-charge and direct-injected two-stroke engines. SAE paper 9045, 99. [5] Melton L.A.: Exciplex-Based apor/liquid isualization Systems Appropriate for Automotive Gasolines. Report #, Contract #7, Texas High Education Coordination Board. November 990. [6] Mitianiec W.: Wtrysk paliwa w silnikach dwusuwowych małej mocy. Polska Akademia Nauk. Kraków, 999. [7] Mitianiec W., Jaroszewski A.,Modele matematyczne procesów fizycznych w silnikach spalinowych małej mocy, Ossolineum, Wroclaw-Warszawa-Krakow, 993 [8] Nakano D., Suzuki T., Matsui M., Gas Engine Ignition System for Long-Life Spark Plugs, SAE Paper / , SETC Graz, 004. [9] Sendyka B., Noga M., Propagation of flame whirl at combustion of lean natural gas charge in a chamber of cylindrical shape, Combustion Engines, 007-SC, PTNSS, Bielsko Biała 007 [0] Spalding D.B., Combustion and Mass transfer, Pergamon Press, 979 [] Rassweiler G.M., Withrow L., Motion Pictures of Engine Flames Correlated with Pressure Cards, SAE Paper 8003, Warrendale, 980 [] ibe I. I., Brennverlauf und Kreisproceß von erbrennungs-motoren, EB Technik, Berlin, 970