Improved Design of Air Flow for a Two Stroke Internal Combustion Engine without Scavenging Problems to Promote Cleaner Combustion

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1 International Journal of Engineering and Applied Sciences (IJEAS) ISSN: , Volume-5, Issue-4, April 2018 Improved Design of Air Flow for a Two Stroke Internal Combustion Engine without Scavenging Problems to Promote Cleaner Combustion Hari Prakash V, Madhan Gopal M M, Antony Prasanth A Abstract In the present work, an attempt is made to reduce scavenging problems by developing a new model of two-stroke spark-ignition engine. This model allows flow of fresh air through intake valves positioned at the bottom of the cylinder and exit of burnt gases through exhaust ports situated at the top of cylinder. The exhaust ports are closed by the piston as it moves towards bottom dead Centre following which gasoline is injected minimizing the possibility of mixing of fuel with outgoing exhaust gases. During expansion the piston unravels the exhaust ports and the burnt gases escape to the atmosphere due to pressure difference. Due to low density at high temperatures, the exhaust gases naturally rise against the direction of gravity reducing the possibility of mixing with incoming fresh air. A comparative study of fuel distribution inside the cylinder, showed greater distribution when injection takes place against gravity than along gravity, which promotes cleaner combustion. The combustion analysis was done using Diesel-RK software and flow analysis was done using ANSYS FLUENT. Index Terms New two stroke engine model; scavenging minimization, intake valves and exhaust ports; conversion of 4 stroke to 2 stroke. I. INTRODUCTION The purpose of internal combustion engines is the production of mechanical power from the chemical energy contained in the fuel. In internal combustion engines, as distinct from external combustion engines, this energy is released by burning or oxidizing the fuel inside the engine. The fuel-air mixture before combustion and the burned Products after combustion are the actual working fluids. The work transfers which provide the desired power output occur directly between these working fluids and the mechanical components of the engine. Because of their simplicity, ruggedness and high power to weight ratio, internal combustion engines have found wide application in transportation (land, sea, and air) and power generation. Based on working cycle IC engines are classified as four stroke and two stroke cycle. In a four stroke IC engine the four strokes of the cycle are intake, compression, power, and exhaust. Each corresponds to one full stroke of the piston; therefore, the complete cycle requires two revolutions of the crankshaft to complete. The cycle of the four-stroke of the piston (the suction, compression, power and exhaust strokes) is completed in two strokes in the case of a two-stroke engine Hari Prakash V, B.tech scholar, School of Mechanical Engineering, SASTRA University, Tamil nadu, India. Madhan Gopal M M, B.tech scholar, School of Mechanical Engineering, SASTRA University, Tamil nadu, India. Antony Prasanth, B.tech scholar, School of Mechanical Engineering, SASTRA University, Tamil nadu, India. producing more power than four stroke engines. The existing two stroke engines are more compact than the four-stroke engines, and they are lightweight and less costly. In addition, the torque produced on the crankshaft is more uniform because the power is produced during every alternate stroke of the piston. The air flow system in two stroke engines are governed by the scavenging process. There are mainly three types of scavenging methods direct/crankcase/cross flow; back flow/loop and uni flow. In case of crankcase scavenging method, the air is drawn into the crankcase due to the suction created by the upward stroke of the piston. On the down stroke of the piston it is compressed in the crankcase, The compression pressure is usually very low, being just sufficient to enable the air to flow into the cylinder through the transfer port when the piston reaches near the bottom of its down stroke. The air thus flows into the cylinder, where the piston compresses it as it ascends, till the piston is nearly at the top of its stroke. The compression pressure is increased sufficiently and the cycle continues. Two stroke engines with an exhaust valve mounted in the cylinder head are known as uni flow scavenged engines. This is because the flow of scavenging air is in one (uni) direction. Some 2 stroke engines do not have exhaust valves. They are fitted with exhaust ports located just above the scavenge ports. As the piston uncovers the exhaust ports on the power stroke, the exhaust gas starts to leave the cylinder. When the scavenge ports are uncovered, scavenge air, loops around the cylinder and pushes the remaining exhaust gas out of the cylinder. This type of engine is known as a loop scavenged engine. The process of scavenging mainly comprises of, pushing out the burnt gases out of the cylinder and drawing in a fresh charge of air or air/fuel mixture for the next cycle. Scavenging in two stroke engines has been an age old problem. Short-circuiting and mixing has been two major shortcomings associated with scavenging systems. If scavenging is incomplete, the suction stroke will suck a mixture of exhaust gases and fresh charge instead of complete fresh charge of clean air. This will lead to incomplete combustion emitting significant amount of particulate matter, un-burnt hydrocarbons, carbon mono-oxide (CO) and Nitrous oxides (Nox).Another problem is that it allows expulsion of some of the fresh-air charge directly to the exhaust leading to increase in fuel consumption. During the recent decades, new factors for change have become important and now significantly affect engine design and operation. These factors are, first, the need to control the automotive contribution to urban air pollution and, second, the need to achieve significant improvements in automotive 11

2 Improved Design of Air Flow for a Two Stroke Internal Combustion Engine without Scavenging Problems to Promote Cleaner Combustion fuel consumption. With depletion of oil resources, the cost of fuel is rising steadily which demanded the need for fuel efficient engines. In addition, It became difficult to cost effectively achieve the emission standards that are being followed in the current world with two stroke engines. This has led to a gradual decrease in the production of two stroke engines for automotive applications. This work deals with improved air flow design of a fuel injected, two stroke cycle engine. The focus of this design is to avoid mixing of fuel with burnt gases and escape of fuel during the exhaust stroke. The proposed design comprises of an engine model with valve and port configuration. The valves are used for intake that are located at the bottom of the cylinder and the burnt gases are sent out through the ports that are located at the top of the cylinder. In addition, a comparative study of fuel distribution when the fuel is injected against the gravity and when the fuel is injected along the direction of gravity inside the cylinder was performed. It was observed that a greater distribution of fuel inside the combustion chamber takes place when injection is done against gravity than along gravity. This in turn reduces the contribution towards fuel film formation on piston head thereby reducing soot formation and enhancing cleaner combustion. II. REVIEW OF LITERATURE There are ample literatures related to the optimization of various operating parameters of internal combustion engines. CFD simulations and two dimensional analysis soft wares are primarily used for optimization. Studies related to development methodology for an internal combustion engine models, include numerical approaches that are validated by experimental procedures. Hoag et al. (2006) provided a concise reference volume on engine design and mechanical development processes for engineers serving the engine industry. It provided basic information and references to guide more in-depth study. Radoslav Plamenov (2011) work gives an insight into the complete design a real engine, having into account all necessary calculations concerning with kinematics, dynamics and strength calculation of basic details. This gave an insight into valve designs and gas forces acting on the piston during combustion. The extract from the works of Seeman (2012), Clive lewis (2007), Hareesh et al. provided an insight into existing numerical approaches and capabilities of existing softwares.these extracts provided a clear vision for devising the approach methodology. Yerrennagoudaru et al. (2014) performed CFD study on IC engine model that was created using ANSYS Design modeler and simulation using ANSYS Fluent. The study also states that the use of Computational Fluid Dynamics (CFD) along with optimization tools can help shorten the design optimization cycle time. Traditional approach of experiments using flow bench testing is very costly as well as time consuming. María Isabel Lamas Galdo and Carlos G. Rodríguez Vidal (2012) describes scavenging as the process by which the fresh charge displaces the burnt gas from the cylinder and it plays a very important role in the performance and efficiency of two-stroke engines. In the study, a CFD analysis was carried out to study the scavenging process of two-stroke engines. This study shows that CFD predictions yield reasonably accurate results that allow improving the knowledge of the fluid flow characteristics. J. Ma et al. presented a comprehensive study and analysis of the 2-stroke uniflow direct injection gasoline engine. Effects of intake port design and engine geometry on the 2-stroke scavenging process and maximum engine performance were investigated by means of 3-D CFD and 1-D engine simulation. The work also stated the uniflow as the most efficient scavenging method for the 2-stroke cycle operation. Also, the uniflow 2-stroke engine can be operated at high boost by closing the exhaust valves earlier. Martin S oder (2013) performed a numerical investigation of flows related to internal combustion engines. The work listed different type of flows that occur within the cylinder and their corresponding relation the geometrical specifications. Abdullah et al. (2008) performed CFD modelling of fuel injection, spark ignition, combustion and pollutant formation to simulate these phenomena and processes in a compressed natural gas direct injection engine. The details of the engine geometry chosen for the calculations were given along with and explanation for the modelling choices made. Yogesh et al. carried out In cylinder cold flow CFD simulation of four stroke petrol engine using hybrid approach of ANSYS fluent. The work states that the central challenge in design is the complex fluid dynamics of turbulent reacting flows with moving parts through the intake/exhaust manifolds, valves, cylinder, and piston. The time scales of the intake air flow, fuel injection, liquid vaporization, turbulent mixing, species transport, chemistry, and pollutant formation all overlap, and need to be considered simultaneously. Intake and Exhaust Flows: Borde et al. (2015) performed CFD analysis of an intake manifold for boosted engine application. This work gave an insight into variation of pressure with respect to CA and runner dimensions in the intake manifold. Fig. 2.1 Variation of pressure with CA and runner dimensions in intake manifold Yin et al. (2014) describes the effect of compression ratio, expansion ratio, supercharge ratio, and excess air coefficient on the performance of a spark ignition engine operating on natural gas. This study provided an insight to effect of expansion ratio on end temperatures and pressures. Vaibhav Kabsuri (2013) provided an insight view of the manifold as what exactly is happening with the turbulence parameter. The project analyzes the characteristics of the air-fuel mixture taking into account the inlet velocity of the 12

3 mixture, other parameters and most importantly the size/dimensions of the intake manifold. This extract related to air flow design provided formulas used for verification of intake manifold velocity. V (14.7 Vg T ch 180) / (520 P (180 )) (2.1) g V g Where = gas velocity- fixed in ft. /min T = Temperature in Rankine (Intake temperature is generally 528 Rankine And Exhaust temperature is 1032 Rankine) ch = Charging efficiency = 85% P = Pressure of gas in psi (14.7psi for intake and 29-35psi for exhaust) 180 = Duration of valve opening V g For stationary engines <= ft. /min for intake valve V g and <= ft. /min for exhaust valves. John B.L Heywood, in his studies provides a relation exhaust manifold pressures and intake manifold pressures at different engine RPMs. International Journal of Engineering and Applied Sciences (IJEAS) ISSN: , Volume-5, Issue-4, April 2018 film may get enough heat from the piston-wall to evaporate, and then decrease the air cooling effect and the compression ratio. Catapano et al. in his work states that Spray/wall interaction has a significantly influence on the mixture formation process in gasoline direct injection (GDI) engines. Moreover, the fuel wall film and the resulting delayed evaporation of the liquid fuel are the main sources of soot formation in the internal combustion engines. Chang et al. in his work states that there is an urgent need for the reduction of liquid fuel impingement on the piston top, because a liquid fuel film vaporizes more slowly than airborne droplets, resulting in high hydrocarbon emissions. Oslen in his work states that, if the fuel is not fully vaporized and properly mixed with the air in the engine s cylinders during the combustion process, part of this fuel may go out of the cylinders as unburned hydrocarbons. The carbon deposits can also cause cold start and drivability problems as the engine warms up because they can actually act as a sponge by momentarily absorbing some of the fuel that is need for proper combustion. III. PRINCIPLE OF OPERATION OF THE MODEL The principle of working of the proposed model is explained below. 1. Initially the piston is at the end of the compression stroke. Fuel is injected and spark initiates combustion and the piston starts moving towards BDC. 2. As the piston uncovers the exhaust ports, due to pressure difference between the atmosphere and cylinder, the burnt gases inside the cylinder starts moving out. 3. When the pressure inside the cylinder drops below the pressure at inlet, the inlet air valves are opened. Due to the pressure difference the inlet air aids in pushing out the Exhaust gases. 4. As the piston covers the ports the inlet air valves are closed. Now the cylinder is already filled with air as the inlet air has to cover the length of the cylinder to push out the exhaust gases. The cycle is repeated. Fig. 2.2 Exhaust manifold and intake manifold pressures at different engine RPMs Effects of Rate Of Fuel Evaporation On Combustion: Wisłocki et al. performed different fuel spray injection video observations and its corresponding effects on various combustion parameters. Alger et al. (2001), used optical access engine to image the liquid film evaporation off the piston of a simulated direct injected gasoline engine. Mie scattering images show the liquid exiting the injector probe as a stream and directly impacting the piston top. Schlieren imaging was used to show the fuel vaporizing off the piston top late in the expansion stroke and during the exhaust stroke. Emissions tests showed that the presence of liquid fuel on in-cylinder surfaces increases engine-out hydrocarbon emissions. Habchi et al. (1999) states that for stratified cold operation, the spray impingement on piston may yield to film formation and large HC and soot exhaust emissions. Moreover, liquid Fig. 3.1 working model of the proposed engine 13

4 Improved Design of Air Flow for a Two Stroke Internal Combustion Engine without Scavenging Problems to Promote Cleaner Combustion IV. METHODOLOGY A. INITIAL APPROACH TO THE PROBLEM The approach is done by modification of air flow of a 4 stroke gasoline engine to that of the proposed two stroke model. This is because of the following reasons. The technical specifications of a real time gasoline engine (Royal Enfield 4 stroke super meteor engine) was available in the internet. Table 1 Initial conversion of valve timing diagram from 4 stroke to 2 stroke PROPOSED 2 OPERATING REFERENCE STROKE STROKE (DEGREES) (DEGREES) Suction Compression Expansion 105(from TDC) 120(from TDC) Exhaust Compression and expansion stroke duration for the proposed model were maintained almost the same as the reference engine by altering the intake and exhaust stroke duration. This lead to the formulation of the objectives to ensure the feasibility of the proposed model. Why is fuel injector mandatory according to this idea? Will sufficient amount of air be injected through the valves in the stipulated time? Is it feasible to send out the exhaust gases within the stipulated time? Has the cylinder pressure dropped below the inlet pressure during inlet valve opening? Will forced induction be necessary to achieve the working of the cycle? Will there be a mixing of exhaust gases with inlet fresh air? B. DETERMINATION OF AIR AND FUEL MASS AND INTAKE VELOCITY Fig. 4.1 Technical details of Royal Enfield super meteor engine Four stroke engines use valves for intake. Since the work deals with modification of air flow by introducing valves for intake of fresh air, the existing intake valve timing of the 4 stroke was altered to that for a two stroke engine with reduced valve timing. Real world two stroke engines have ports and valves according to their design and altering them appeared to be cumbersome. To obtain the air flow and thermodynamic parameters of the four stroke engine, the geometrical parameters (valve timing, stroke, bore, compression ratio) of Royal Enfield Super Meteor Engine were fed into a software called Diesel R K and a reference engine was developed. For initial progress the valve timings for different strokes were fixed to alter the other parameters according to the air flow needed. The valve timing diagram of the 4 stroke was modified for the proposed 2 stroke model. CH. 8 The chemical formula for pure octane fuel is 18 C H 12.5 O 8 CO 9 H O (4.2) The stoichiometric ratio is 15.1:1. However chemical analysis for perfect gasoline/air is C 8H 11.85O 2 8CO 2 7.7H 2O (4.2) 15.4 The stoichiometric air fuel ratio is 14.7:1.However under wide open throttle conditions it is considered as 14.6:1.In present work the engine is run at a constant 5000 RPM. As no throttling is done wide open throttle conditions are considered for calculations. The total mass of air is found from the mass of fuel supplied and equivalent ratio. Fig. 4.2 A/F ratio and mass of fuel supplied in reference engine Mass of Air Calculation: 1.001=(A/F ACTUAL) /(A/F STOICHIOMETRIC) (4.3) = A/F ACTUAL 3 Mass of fuel (reference engine details) = kg 4 Actual air supplied = kg Total flow through intake = mass of air + mass of fuel = kg kg 4 = kg By applying flow rate equation, 14

5 Π/4 d2 4 V time = kg (4.4) d= 2.66 cm Velocity = m/s. In order to verify if the velocity of gases are within the limits at inlet and exhaust flows the following verification is made from literature study. V (14.7 Vg T ch 180) / (520 P (180 )) g (4.5) V g Where = gas velocity- fixed in ft. /min T = Temperature in Rankine (Intake temperature is generally 528 Rankine and Exhaust temperature is 1032 Rankine) ch = Charging efficiency = 85% P = Pressure of gas in psi (14.7psi for intake and 29-35psi for exhaust) 180 = Duration of valve opening V g For stationary engines <= ft. /min for intake valve and V g <= ft. /min for exhaust valves. International Journal of Engineering and Applied Sciences (IJEAS) ISSN: , Volume-5, Issue-4, April 2018 Iteration 1: The position of the transfer port was initially fixed at 1 cm and 7 cm wide from the exhaust port opening point. Almost the same amount of air and fuel ( g of gasoline) as in the model engine was supplied (equivalence ratio 1.001).The pressure drop at the end of 1 cm of the exhaust port was 6 bar (not atmospheric). Hence opening the inlet valve would not permit the flow of fresh air. Iteration 2: So the model was modified in such a way that the exhaust port measured 2 cm from the opening point. The width of the port was extended to 20 cm from 7 cm.the amount of fuel supplied was g per cycle (slightly lesser than that of the reference engine) and A/F equivalent ratio was maintained at For Intake valves, V = ( ) / ( g 270) V g = 6853 ft. /min Which is less than ft. /min as per literature study. For Exhaust valves, V = ( ) / ( g 290) = ft. /min Fig. 4.4 Pressure CA diagram of crankcase scavenged model ( g fuel) Which is also less than ft. /min according to literature study. Thus m/s was verified to be suitable to be used in intake calculations. C. EXHAUST CALCULATIONS The focus of exhaust calculations was to design the dimensions of the exhaust port that would be capable of sending out all the combusted gases out of the cylinder within 120 degree of exhaust timing and correspondingly attain a cylinder pressure that is less than atmospheric. The point at which exhaust port should open is 6 cm from TDC. So the existing crankcase scavenging model of 2 stroke is modified in such a way that the exhaust port opens at 6 cm from the top dead Centre. The dimensions were fixed by narrowing down to the required dimensions through several iterations. Fig. 4.5 Pressure at the end of 2 cm height of exhaust port ( g fuel) The pressure at the end of 2 cm was found to be 0.83 bar which is less than atmospheric. Iteration 3: Not able to achieve the exact quantity of fuel supplied as of that in reference engine the amount of fuel supplied was g per cycle, a little higher than that of the reference engine with air fuel equivalence ratio at Fig. 4.3 Exhaust port location and dimension in model engine Fig. 4.6 Pressure CA diagram of crankcase scavenged model ( g fuel) 15

6 Improved Design of Air Flow for a Two Stroke Internal Combustion Engine without Scavenging Problems to Promote Cleaner Combustion Fig. 4.8 Velocity curve of exhaust gases Fig. 4.7 Pressure at the end of 2 cm height of exhaust port ( g fuel) The pressure at the end of 2 cm was found to be 0.75 bar. (Below atmospheric pressure) The port of dimension 2 cm height and 20 cm width is able to reduce the pressure at the end of expansion stroke to a pressure less than the atmospheric. Now the check for exhaust port to be sufficient to send out the kg of exhaust gases. Two separate integrals are to done, one for velocity and another for area. Curve fitting from the values of Iteration 2 was selected for calculation (closest to ).\ Fig. 4.9 Area curve of exhaust ports The velocity is integrated for the time interval of 40 degrees, ( ) 0 x x x x x x dx ( x 6 / 7) (x / 6) ( x / 5) ( x / 4) (x / 3) 50342(x / 2) 2 10 ( x) ( ) ( m ) ( ) ( ) The equation of area is integrated for the length of the port, (0.1x 5 10 ) dx ( x / 2) 5 10 ( x) 2 10 (0.2) 10 2 (10 ) 15 =0.002 m2 Mass of gases sent out was found by Mass = density velocity area time = kg. (4.6) Now this was cross verified by taking instantaneous product of velocity and area for definite intervals.the sum was multiplied by time and density. 2 0 Fig Product of instantaneous velocity and instantaneous area Fig CA vs. exhaust gas velocity The product of velocity*area*density*time for 0.5 degree was kg. Which signifies that, assuming that the cylinder is completely filled with air, the total mass of air flowing out during the exhaust stroke through the exhaust port of 2cm height and 20 cm width is kg. In both the cases the exhaust port is capable of sending out mass greater than that of the incoming charge. 16

7 A. INTAKE CALCULATIONS The piston consists of a set of piston rings. In our model the piston rings plays an important role in altering the CA degree allotted for each stroke. International Journal of Engineering and Applied Sciences (IJEAS) ISSN: , Volume-5, Issue-4, April 2018 T T 75 o C c e For aluminum alloys. Thus the thickness of the piston is found out by the following calculations (4.7) H C HCV m BP( kw) H m o sec (4.8) Fig Effect of piston rings on intake duration Calculation Of : 360 o x x sec Here x denotes time for one complete revolution 3 5 m ( ) / BP H kw t H / (12.56 k ( T T )) t h c e h 2.87mm Thus the distance was found out to be a minimum of 2.87 mm (0.3cm approximately) Method 2: The thickness can also be found out by the following method, t PD (4.9) h 2 (3 ) / (16 ) t Fig Cross sectional view of a piston Table 4.2 material properties of the piston METHOD 1: METHOD 2: Method 1: For maximum gas pressure of 4.68 MPa t h =6.91 mm or 7 mm. Thus an intermediate value of 5mm was chosen as the distance between piston top and the first compression ring. This 5 mm gives an additional 10 degree to the intake stroke. So, the intake has to be modified to 90 degree instead of 80 degree. The diameter required for the valves is found out by the following calculation. The number of valves is initially assumed to be one and the calculation was carried out. It was found that it was impossible to make a valve of such diameter to fit into the cylinder bore. Hence the number of valves (n) has to be 3. Time = 90 degree = sec Mass of the air inducted = 2 4 ( / 4) d V time n Kg Thus we get d = 2.66 cm. Now the intake is developed for 2.66 cm port size and 90 degree valve timing. Fig Formula expansion 17

8 Improved Design of Air Flow for a Two Stroke Internal Combustion Engine without Scavenging Problems to Promote Cleaner Combustion Iteration 1: First design was made for an inlet pressure of 1.3 bar forced induction as it was idealized before. Effect Of Clearance Volume Between Piston Diameter And Bore: The clearance between piston and bore from literature study is found to be between 0.1 mm to 0.3 mm. In case of the 4 stroke engine the combustion chamber is completely sealed. This the compression ratio is not altered by the position of piston rings. However in case of the proposed model, the combustion chamber gets sealed only after the piston ring reaches the end of exhaust port towards TDC which causes a small reduction in the compression ratio. Bore=7 cm Fig Velocity variation with CA for 1.3 bar inlet 3 valves of diameter 26.6 mm forced induction at 1.3 bar was 4 able to pump in Kg of air. Required amount to be sent inside during inlet is 3.65 X 10-4 kg. Iteration 2: Pressure 1.1 bar. Accounting in density variations with pressure P R T (4.10) / ( ) = kg/ m Piston diameter = 7 cm- 0.2 mm =69.8 mm. Difference in area= 3.14/4 (bore 2 piston diameter 2) =21.9 mm2 (4.11) Clearance volume = (height from piston ring 21.9 mm2) (4.12) = 5 mm 21.9 mm2 = mm2 = cm2 Maximum volume of the cylinder = 3.14/4 D2 L = cm3 (4.13) Compression ratio theoretical = 7.5 = cm3 / Minimum volume of the cylinder Minimum volume of the cylinder = cm3 Actual minimum volume= cm3 + clearance volume (4.14) Fig Velocity variation with CA for 1.1 bar inlet The mass of air through the inlet is X 10-4 kg. (Required is 3.65 X 10-4 kg) Iteration 3: Natural aspiration. = cm3. Actual compression ratio = A. GRAVITY EFFECTS ON CRANK AND SLIDER MECHANISM The arrangement piston controlled exhaust ports at the top is achieved by motion of piston against the gravity during expansion stroke. This in simpler terms is inversion of the slider crank mechanism present in the conventional engine. The effects on piston as it moves opposite to that of in conventional engine was studied. It was found to be the same to that in conventional mechanism. The setup was simulated in ADAMS Software as shown below. The dimensions and materials used are as follows: Fig Velocity variation with CA for natural aspiration Amount of air through inlet in natural aspiration is X 10-4 kg, which is less than 3.65 X 10-4 kg (Required amount).thus sufficient amount of air can be injected through the valves in the stipulated time at 1.1 bar. Piston: width 7 cm, 3 cm height, 2 cm thickness (Aluminum) Connecting rod: 16.5 cm length, 2 cm thickness, 2 cm width (steel) Crank rod: cm length, 2 cm thickness, 2 cm width (steel) Rpm: 5000 force: N (46.8 bar at 5000 RPM) 18

9 International Journal of Engineering and Applied Sciences (IJEAS) ISSN: , Volume-5, Issue-4, April 2018 Fig Velocity graph for inverted crank slider mechanism Fig Force analysis representation on slider crank mechanism Gravity Not Reversed: Fig Acceleration graph inverted crank slider mechanism Fig Normal gravity settings Fig Velocity graph for normal crank slider mechanism B. GRAVITY EFFECTS ON FUEL INJECTION AND AIR FLOW As the exhaust ports are situated at the top of cylinder and intake occurs through bottom induction, there is very less chance of mixing of burnt exhaust gases mixing with incoming fresh charge. Since the combustion simulations have been done using a crankcase scavenging system, the heat transfer to the intake system cannot be considered for calculations as they design is completely different in the proposed model (crank case scavenging: ports for intake, proposed: valves for intake).so assuming the value obtained from crankcase system would be approximately be close to that of the intake manifold the difference in density of dry air was studied. The average temperature at crankcase wall was 60 degree Celsius and corresponding air density is kg/m3 at atmospheric pressure (natural aspiration).the temperature of cylinder at the end of combustion is 1977 degree Celsius and the corresponding density is kg/m3 at atmospheric pressure (0.85 bar at end of exhaust stroke).this ensures that the burnt gases have very less possibility to mix with incoming fresh air. Fig Acceleration graph for normal crank slider mechanism Gravity Reversed: Fig Temperature variation with CA Fig Setting for gravity reversed in ADAMS Fig Temperature at the end of exhaust port opening 19

10 Improved Design of Air Flow for a Two Stroke Internal Combustion Engine without Scavenging Problems to Promote Cleaner Combustion The temperature of exhaust manifold is 727 degree Celsius and corresponding pressure of air is 0.6 bar at the time of complete opening of exhaust ports. As the gases of reduced density from combustion chamber pass to the exhaust port, the pressure at the combustion chamber is 0.85 bar which is higher than that in the exhaust port. This shall ensure no backflow of burnt gases back into the combustion chamber. Fig Exhaust manifold pressure variation with CA Fig Pressure contour for engine placed along gravity The inlet velocity highest was 60.7 m/s which is equal to the velocity of inlet as per mathematical model of m/s. Against gravity: Fig Exhaust manifold temperature variation with CA Change in flow parameters due to inversion of cylinder: Fig Velocity contour for engine placed against gravity First air flow velocity into the cylinder was analyzed, with inversion of flow direction and then followed by inversion of direction of fuel injection. Rate at which air was supplied was kg/sec for seconds (5000 RPM and 90 degree inlet).this ensures supply of required kg of air. Exhaust port was at atmospheric pressure. Along Gravity: Fig Pressure contour for engine placed against gravity This shows there is a marginal change in inlet velocity due to change in gravity direction. Density of air being 1.2 kg/m3, the effect of gravity is very less. Fig Velocity contour for air flow along gravity Fuel Injection: The fuel has to be injected precisely at 3 cm from the BDC to ensure sealed combustion chamber and avoidance of scavenging completely. The flow of fuel into the cylinder, Without the presence of air was analyzed for 6.6*E-4 seconds at a rate of kg/sec which equals kg of fuel getting injected in 20 degrees at constant 200 bar pressure in both the cases. 20

11 International Journal of Engineering and Applied Sciences (IJEAS) ISSN: , Volume-5, Issue-4, April 2018 Along Gravity: intake manifold as it is capable of almost the same amount of air as that of forced induction at 1.1 bar. C. Is it feasible to send out the exhaust gases within the stipulated time? YES The 120 degree of exhaust timing with exhaust port dimensions of 2cm height and 20cm width is sufficient to send out all the combusted mass inside the cylinder. The pressure at the end of 2cm was found to be 0.83 bar which is less than atmospheric. The amount of gases that can be sent out was found to be kg (assuming continuous flow without change in density) which is surplus to the required amount to be sent out. and also the cylinder pressure dropped below the inlet pressure during inlet valve opening. Fig Velocity contour for fuel injection along gravity Against Gravity: D. Has the cylinder pressure dropped below the inlet pressure during inlet valve opening? YES The intake pressure at the point where intake valve opens was measured to be 0.98 bar (natural aspiration) and above 1.1 bar (forced induction 1.1 bar).in both cases the pressure at the time of induction is higher is less than that of the cylinder pressure (0.83 bar). Fig Velocity contour for fuel injection against gravity The decrease in velocity is 12 m/s. However from the contours it is seen that by inverting the engine the distribution of fuel is over a larger area after impact on cylinder head at a very short duration ( seconds and at a RPM of 5000). Fig. 5.1 Inlet pressure variation for forced induction at 1.1 bar V. RESULTS AND DISCUSSIONS The address to various objectives listed was done and the following inferences have be incurred. A. Is fuel injector mandatory according to this idea? Yes. According to the proposed concept the fuel has to be injected at the point where the top of the piston reaches the start of exhaust port. This shall provide a sealed combustion chamber eliminating all possibilities of short circuiting and mixing. In the absence of fuel injection there will be mixing of inlet fresh air fuel mixture with that of burnt charge at the end of expansion stroke and that will promote scavenging. B. Will sufficient amount of air be injected through the valves in the stipulated time? YES. From the graphs of intake calculation, we find that 90 degree of intake stroke is sufficient to send in the required amount of air when 3 valves of diameter 2.66 cm are used and the manifold remains unaltered to that of the model engine. Perfect induction would be forced induction at 1.1 bar. However natural aspiration can also be used by tuning the Fig. 5.2 Inlet pressure variation for Natural aspiration E. Will forced induction be necessary to achieve the working of the cycle? NOT MANDATORILY The mass of air inducted using forced induction at 1.1 bar was is X 10-4 kg. (Required is 3.65 X 10-4 kg) and in case of natural aspiration 3.60 X 10-4 kg was inducted. So a little tuning of intake manifold would ensure the use of natural aspiration instead of forced induction. F. Will there be a mixing of exhaust gases with inlet fresh air? ALMOST ELIMINATED The average temperature at crankcase wall was 60 degree Celsius and corresponding air density is kg/m3 at atmospheric pressure (natural aspiration).the temperature of 21

12 Improved Design of Air Flow for a Two Stroke Internal Combustion Engine without Scavenging Problems to Promote Cleaner Combustion cylinder at the end of combustion is 1977 degree Celsius and the corresponding density is kg/m3 at atmospheric pressure (0.85 bar at end of exhaust stroke).this ensures that the burnt gases have very less possibility to mix with incoming fresh air. In addition to that, from the contours of fuel it is seen that by fuel injection from the bottom head of the engine, the distribution of fuel is over a larger area than that from the top of the cylinder head, after impact on cylinder head at a very short duration ( seconds and at a RPM of 5000).This shall enhance greater mixing between air and fuel during compression enhancing cleaner combustion. The velocity of fuel was reduced by 12 m/s during fuel injection from bottom of the cylinder, which in turn signifies lesser impact on cylinder head reducing the thin film fuel layer formation on cylinder head over a period of time reducing the possibility of soot formation on the piston head to certain extent. VI. CONCLUSIONS The results of air flow design validate the possibility of reduction of scavenging and promotion of cleaner combustion which remained the prime focus of the present work. The final CA diagram is shown below. Fig. 6.1 Final conversion CA diagram Table 6.1 Final Conversion table of valve timing diagram OPERATION REFERENCE ENGINE (DEGREES) PROPOSED 2 STROKE (DEGREES) Suction Compression Expansion 105 (FROM TDC) 120 (FROM TDC) Exhaust REFERENCES [1] Abdullah, Shahrir, Wendy Hardyono Kurniawan and Azhari Shamsudeen, (2008) [2] Numerical Analysis of the Combustion Process in a Compressed Natural Gas Direct Injection Engine, Department of Mechanical and Materials Engineering, National University of Malaysia, D.E., UKM Bangi, Malaysia. 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