Six Stroke Ic Engine Essay Example for Free

Six Stroke Ic Engine Essay 1. INTRODUCTION A diesel engine is an internal combustion engine that uses the heat of compression to initiate ignition to burn the fuel, which is injected into the combustion chamber during the final stage of compression. Diesel engines have wide range of utilization for automobiles, locomotives marines and co-generation systems. However, large problem is still related to undesirable emission. The six-stroke engine is a type of internal combustion engine based on the four-stroke engine but with additional complexity to make it more efficient and reduce emissions. Two different types of six-stroke engine have been developed: In the first approach, the engine captures the heat lost from the four-stroke Otto cycle or Diesel cycle and uses it to power an additional power and exhaust stroke of the piston in the same cylinder. Designs use either steam or air as the working fluid for the additional power stroke. The pistons in this type of six-stroke engine go up and down three times for each injection of fuel. There are two power strokes: one with fuel, the other with steam or air. The currently notable designs in this class are the Crower Six-stroke engine invented by Bruce Crower of the U.S. ; the Bajulaz engine by the Bajulaz S.A. company of Switzerland; and the Velozeta Six-stroke engine built by the College of Engineering, at Trivandrum in India. The second approach to the six-stroke engine uses a second opposed piston in each cylinder that moves at half the cyclical rate of the main piston, thus giving six piston movements per cycle. Functionally, the second piston replaces the valve mechanism of a conventional engine but also increases the compression ratio. The currently notable designs in this class include two designs developed independently: the Beare Head engine, invented by Australian Malcolm Beare, and the German Charge pump, invented by Helmut Kottmann. To improve exhaust emissions from diesel engines, a new concept of Six Stroke Engine has been proposed. This engine has a second compression and combustion processes before exhaust process. [pic] Fig 1 Diesel engine sectional view Fig 2 Ideal Otto cycle [pic] Fig 3 Pressure- Volume diagrams for dual cycle As the fuel in one cycle was divided into two combustion processes and the EGR (Exhaust Gas Recirculation) effect appeared in the second combustion process, the decreased maximum cylinder temperature reduced Nitrous Oxide (NO) concentration in the exhaust gas. It was further confirmed that soot formed in the first combustion process was oxidized in the second combustion process .Therefore, a six stroke diesel engine has significant possibilities to improve combustion process because of its more controllable factors relative to a conventional four-stroke engine. Since the cylinder temperature before the second combustion process is high because of an increased temperature in the first combustion process, ignition delay in the second combustion process should be shortened. In addition, typically less desirable low cetane number fuels might also be suitable for use in the second combustion process, because the long ignition delays of these fuels might be improved by increased cylinder temperatures from the first combustion process. Methanol was chosen as the fuel of the second combustion. The cetane number of methanol is low and it shows low ignitability. However, since methanol will form an oxidizing radical (OH) during combustion, it has the potential to reduce the soot produced in the first combustion process. [pic] Fig 4 Comparison of 4 stroke and 6 stroke cycle 2. BAJULAZ SIX STROKE ENGINE The majority of the actual internal combustion engines, operating on different cycles have one common feature, combustion occurring in the cylinder after each compression, resulting in gas expansion that acts directly on the piston (work) and limited to 180 degrees of crankshaft angle. According to its mechanical design, the six-stroke engine with external and internal combustion and double flow is similar to the actual internal reciprocating combustion engine. However, it differentiates itself entirely, due to its thermodynamic cycle and a modified cylinder head with two supplementary chambers: Combustion, does not occur within the cylinder within the cylinder but in the supplementary combustion chamber, does not act immediately on the piston, and it’s duration is independent from the 180 degrees of crankshaft rotation that occurs during the expansion of the combustion gases (work). The combustion chamber is totally enclosed within the air-heating chamber. By heat exchange through the glowing combustion chamber walls, air pressure in the heating chamber increases and generate power for an a supplementary work stroke. Several advantages result from this, one very important being the increase in thermal efficiency. IN the contemporary internal combustion engine, the necessary cooling of the combustion chamber walls generates important calorific losses. 2.1 Analysis: Six-stroke engine is mainly due to the radical hybridization of two- and four-stroke technology. The six-stroke engine is supplemented with two chambers, which allow parallel function and results a full eight-event cycle: two four-event-each cycles, an external combustion cycle and an internal combustion cycle. In the internal combustion there is direct contact between air and the working fluid, whereas there is no direct contact between air and the working fluid in the external combustion process. Those events that affect the motion of the crankshaft are called dynamic events and those, which do not effect are called static events. [pic] Fig 5 Prototype of Six stroke engine internal view 1. Intake valve, 2.Heating chamber valve, 3.Combustion chamber valve,4. Exhaust valve, 5.Cylinder, 6.Combustion chamber, 7. Air heating chamber, 8.Wall of combustion chamber, 9.Fuel injector and 10.Heater plug. 2.1.1 Analysis of events [pic] Fig 6 Event 1: Pure air intake in the cylinder (dynamic event) 1. Intake valve. 2. Heating chamber valve 3. Combustion chamber valve. 4. Exhaust valve 5. Cylinder 6. Combustion chamber. 7. Air heating chamber. 8. Wall of combustion chamber. 9. Fuel injector. 10. Heater plug. [pic] Fig 7 Event 2: Pure air compression in the heating chamber. Event 3: Keeping pure air pressure in closed chamber where a maximum heat exchange occurs with the combustion chambers walls, without direct action on the crankshaft (static event). [pic] Fig 8 Event 4: Expansion of the Super heat air in the cylinder work (dynamic Event). [pic] Fig 9 Event 5: Re-compressions of pure heated air in the combustion chamber (dynamic event). Events 6: fuel injection and combustion in closed combustion chamber, without direct action on the crankshaft (static event). [pic] Fig 10 Events 7: Combustion gases expanding in the cylinder, work (dynamic event). [pic] Fig 11 Events 8: Combustion gases exhaust (dynamic event). [pic] Fig 12 Six-stroke engine cycle diagram: 2.1.2 External combustion cycle: (divided in 4 events): No direct contact between the air and the heating source. e1. (Event 1) Pure air intake in the cylinder (dynamic event). e2. (Event 2) Compression of pure air in the heating chamber (dynamic event). e3. (Event 3) Keeping pure air pressure in closed chamber where a maximum heat exchange occurs with the combustion chambers walls, without direct action on the crankshaft (static event). e4. (Event 4) Expansion of the super heated air in the cylinder, work (dynamic event). 2.1.3 Internal combustion cycle: (divided in 4 events) Direct contact between the air and the heating source. I1. (Event 5) Re-compression of pure heated air in the combustion chamber (dynamic event) I2. (Event 6) Fuel injection and combustion in closed combustion chamber, without direct action on the crankshaft (static event). I3. (Event 7) Combustion gases expanding in the cylinder, work (dynamic event). I4. (Event 8) Combustion gases exhaust (dynamic event). 2.2 Constructional details: The sketches shows the cylinder head equipped with both chambers and four valves of which two are conventional (intake and exhaust). The two others are made of heavy-duty heat-resisting material. During the combustion and the air heating processes, the valves could open under the pressure within the chambers. To avoid this, a piston is installed on both valve shafts which compensate this pressure. Being a six-stroke cycle, the camshaft speed in one third of the crankshaft speed. The combustion chambers walls are glowing when the engine is running. Their small thickness allows heat exchange with the air-heating chamber, which is surrounding the combustion chamber. The air-heating chamber is isolated from the cylinder head to reduce thermal loss. Through heat transfer from the combustion chamber to the heating chamber, the work is distributed over two strokes, which results in less pressure on the piston and greater smoothness of operation. In addition, since the combustion chamber is isolated from the cylinder by its valves, the moving parts, especially the piston, are not subject to any excessive stress from the very high temperatures and pressures. They are also protected from explosive combustion or auto-ignition, which are observed on ignition of the air-fuel mixture in conventional gas or diesel engines. The combustion and air-heating chambers have different compression ratio. The compression ratio is high for the heating chamber, which operates on an external cycle and is supplied solely with pure air. On the other hand, the compression ratio is low for the combustion chamber because of effectively increased volume, which operates on internal combustion cycle. The combustion of all injected fuel is insured, first, by the supply of preheated pure air in the combustion chamber, then, by the glowing walls of the chamber, which acts as multiple spark plugs. In order to facilitate cold  starts, the combustion chamber is fitted with a heater plug (glow plug). In contrast to a diesel engine, which requires a heavy construction, this multi-fuel engine, which can also use diesel fuel, may be built in a much lighter fashion than that of a gas engine, especially in the case of all moving parts. Injection and combustion take place in the closed combustion chamber, therefore at a constant volume, over 360 degrees of crankshaft angle. This feature gives plenty of time for the fuel to burn ideally, and releases every potential calorie (first contribution to pollution reduction). The injection may be split up, with dual fuel using the SNDF system (Single Nozzle, Dual Fuel). The glowing walls of the combustion chamber will calcite the residues, which are deposited there during fuel combustion (second contribution to pollution reduction). As well as regulating the intake and exhaust strokes, the valves of the heating and the combustion chambers allow significantly additional adjustments for improving efficiency and reducing noise. 2.3 Factors Contributing To the Increased Thermal Efficiency, Reduced Fuel Consumption, and Pollutant Emission 1. The heat that is evacuated during the cooling of a conventional engine’s cylinder head is recovered in six-stroke engine by air-heating chamber surrounding the combustion chamber. 2. After intake, air is compressed in the heating chamber and heated through 720 degrees of crankshaft angle, 360 degrees of which in closed chamber (external combustion). 3. The transfer of heat from thin walls of the combustion chamber to the air heating chambers lowers the temperature, pressure of gases on expansion and exhaust (internal combustion). 4. Better combustion and expansion of gases that take place over 540 degrees of crankshaft rotation, 360 ° of which is in closed combustion chamber, and 180 ° for expansion. 5. Elimination of the exhaust gases crossing with fresh air on intake. In the six stroke-engines, intake takes place on the first stroke and exhaust on the fourth stroke. 6. Large reduction in cooling power. The water pump and fan outputs are reduced. Possibility to suppress the water cooler. 7. Less inertia due  to the lightness of the moving parts. 8. Better filling of the cylinders on the intake due to the lower temperature of the cylinder walls and the piston head. 9. The glowing combustion chamber allows the finest burning of any fuel and calcinate the residues. 10. Distribution of the work: two expansions (power strokes) over six strokes, or a third more than the in a four-stroke engine. Since the six-stroke engine has a third less intake and exhaust than a four stroke engine, the depression on the piston during intake and the back pressure during exhaust are reduced by a third. The gain in efficiency balances out the losses due to the passage of air through the combustion chamber and heating chamber valves, during compression of fresh and superheated air. Recovered in the six-stroke engine By the air-heating chamber surrounding the combustion. Friction losses, theoretically high er in the six-stroke engine, are balanced by a better distribution of pressure on the moving parts due to the work being spread over two strokes and the elimination of the direct combustion. 3. DUAL FUEL SIX STROKE ENGINE 3.1 Working The cycle of this engine consists of six strokes: 1. Intake stroke 2. First compression stroke 3. First combustion stroke 4. Second compression stroke 5. Second combustion stroke 6. Exhaust stroke [pic] Fig 13 Concept of a Six-stroke diesel engine 3.1.1 Intake or Suction stroke To start with the piston is at or very near to the T.D.C., the inlet valve is open and the exhaust valve is closed. A rotation is given to the crank by the energy from a flywheel or by a starter motor when the engine is  just being started. As the piston moves from top to bottom dead centre the rarefaction is formed inside the cylinder i.e. the pressure in the cylinder is reduced to a value below atmospheric pressure. The pressure difference causes the fresh air to rush in and fill the space vacated by the piston. The admission of air continues until the inlet valve closes at B.D.C. 3.1.2 First Compression stroke Both the valves are closed and the piston moves from bottom to top dead centre. The air is compressed up to compression ratio that depends upon type of engine. For diesel engines the compression ratio is 12-18 and pressure and temperature towards the end of compression are 35-40 kgf/cm2 and 600-700 0C 3.1.3 First combustion stroke This stroke includes combustion of first fuel (most probably diesel) and expansion of product of combustion. The combustion of the charge commences when the piston approaches T.D.C. Here the fuel in the form of fine spray is injected in the combustion space. The atomization of the fuel is accomplished by air supplied. The air entering the cylinder with fuel is so regulated that the pressure theoretically remains constant during burning process. In airless injection process, the fuel in finely atomized form is injected in combustion chamber. When fuel vapors raises to self ignition temperature, the combustion of accumulated oil commences and there is sudden rise in pressure at approximately constant volume. The combustion of fresh fuel injected into the cylinder continues and this ignition is due to high temperature developed in engine cylinder. However this latter combustion occurs at approximately constant pressure. Due to expansion of gases piston moves downwards. The reciprocating motion of piston is converted into rotary motion of crankshaft by connecting  rod and crank. During expansion the pressure drop is due to increase in volume of gases and absorption of heat by cylinder walls. 3.1.4 Second compression stroke Both the valves are closed and the piston moves from bottom to top dead centre. The combustion products from the first compression stroke are recompressed and utilized in the second combustion process before the exhaust stroke. In typical diesel engine combustion the combustion products still contains some oxygen. 3.1.5 Second combustion stroke This stroke includes combustion of second fuel having low cetane (Cetane number of fuel is defined as percent volume of cetane (C16H34) in a mixture of cetane and alpha-methyl-naphthalene that produces the same delay period or ignition lag as the fuel being tested under same operating conditions on same engine). The combustion of the charge commences when the piston approaches to TDC. The second fuel injected into recompressed burnt gas can be burnt in the second combustion process. In other words combustion process of the second fuel takes place in an internal full EGR (Exhaust Gas Recirculation) of the first combustion. This second combustion process was the special feature of the proposed Six Stroke DI Diesel Engine. 3.1.6 Exhaust stroke The exhaust valve begins to open when the power stroke is about to complete. A pressure of 4-5 kgf/cm2 at this instant forces about 60% of burnt gases into the exhaust manifold at high speed. Much of the noise associated with automobile engine is due to high exhaust velocity. The remainder of burnt gases is cleared of the swept volume when the piston moves from TDC to BDC. During this stroke pressure inside the cylinder is slightly above the atmospheric value. Some of the burnt gases are however  left in the clearance space. The exhaust valve closes shortly after TDC. The inlet valve opens slightly before the end of exhaust stroke and cylinder is ready to receive the fresh air for new cycle. Since from the beginning of the intake stroke the piston has made six strokes through the cylinder (Three up And Three down). In the same period crank shaft has made three revolutions. Thus for six stroke cycle engine there are two power strokes for every three revolutions of crank shaft. 3.2 Performance analysis 3.2.1 Modification over four stroke diesel engine This six-stroke diesel engine was made from a conventional four-stroke diesel engine with some modification. A sub-shaft was added to the engine, in order to drive a camshaft and injection pumps. The rotation speed of the sub-shaft was reduced to 1/3 of the rotation of an output shaft. To obtain similar valve timings between a four-stroke and a six-stroke diesel engine, the cam profile of the six-stroke diesel engine was modified. In order to separate the fuels, to control each of the injection timings and to control each injection flow rate in the first and the second combustion processes, the six-stroke diesel engine was equipped with two injection pumps and two injection nozzles. The injection pumps were of the same type as is used in the four-stroke diesel engine. The nozzle is located near the center of a piston cavity, and has four injection holes. For the six-stroke diesel engine, one extra nozzle was added on the cylinder head. This extra nozzle was of the same design as that of the four-stroke engine. [pic] Fig 14 Volume –Angle diagram for six stroke engine Diesel fuel for the first combustion process was injected through this extra nozzle, and methanol for the second combustion process was injected through the center nozzle. Here, we denoted the injection timing of the four  stroke diesel engine as Xi. The injection timings of the first and second combustion strokes for the six-stroke diesel engine are shown as Xi I and Xi II, respectively. Crank angle X was measured from the intake BDC. In the six-stroke engine, crank angle of the first combustion TDC is 180 degrees. The second combustion TDC is 540 degrees. Specifications of the test engines are shown in Table 1. The conventional four-stroke diesel engine that was chosen as the basis for these experiments was a single cylinder, air cooled engine with 82 mm bore and 78 mm stroke. The six-stroke engine has the same engine specifications except for the valve timings. However, the volumetric efficiency of the six-stroke engine showed no significant difference from that of the four-stroke engine. Characteristics of the six-stroke diesel engine were compared with the conventional four-stroke diesel engine. In this paper, the engine speed (Ne) was fixed at 2,000 rpm. Cylinder and line pressure indicators were equipped on the cylinder head. NO concentration was measured by a chemiluminescence’s NO meter, and soot emission was measured by a Bosch smoke meter. The physical and combustion properties of diesel fuel and methanol are shown in Table. 2. Since combustion heats of diesel fuel and methanol are different, injection flow rates of the first and the second combustion processes are defined by the amount of combustion heat. Here, the supplied combustion heat for the first combustion process is denoted by QI. The second combustion stroke is denoted by QII. The ratio of QII to Qt (Qt = QI+QII) supplied combustion heat per cycle) is defined as the heat allocation ratio ÃŽ ±H: ÃŽ ±H = QII = QII QI +QII Qt Table 1. Specifications of the test engine: Four stoke Six stroke Diesel Engine Diesel Engine Engine type DI, Single cylinder, Air cooled, OHV Bore x Stroke [mm] 82 x 78 Displacement [cc] 412 Top Clearance [mm] 0.9 Cavity Volume [cc] 16 Compression ratio 21 Intake Valve Open100 BTDC70 BTDC Intake valve Close1400 BTDC1450 BTDC Exhaust Valve Open1350 ATDC1400 ATDC Exhaust Valve Close120 ATDC30 ATDC Valve Overlap 220 100 Rated power 5.9 kW /3000rpm Base Engine - Table 2. Physical and combustion properties of diesel fuel and methanol: | |Diesel Fuel |Methanol | |Combustion heat [MJ/kg] |42.7 |19.9 | |Cetane number |40-55 |3.0 | |Density [kg/m2] |840 |793 | |Theoretical air-fuel ratio |14.6 |6.5 | 3.3 Performance of six stroke diesel engine 3.3.1 Comparison with four stroke diesel engine A four-stroke engine has one intake stroke for every two engine rotations. For the six-stroke engine, however, the intake stroke took place once for every three engine rotations. In order to keep the combustion heat per unit time constant, the combustion heat supplied to one six-stroke cycle should be 3 or 2 times larger than that of the four-stroke engine. There are many ways to compare performance between the four-stroke and six-stroke engines. For this paper, the authors have chosen to compare  thermal efficiency or SFC at same output power. If the thermal efficiency was the same in both engines, the same output power would be produced by the fuels of equivalent heats of combustion. Therefore, in order to make valid comparison, fuels supplied per unit time were controlled at the same value for both engines and engine speeds were kept constant. In this section, fuel supplied for the engines was only a diesel fuel. Performance of the six-stroke engine was compared with that of the four-stroke engine under various injection timings. Detailed conditions for comparison of the four-stroke and six-stroke engines are listed in Table. 3. The heat allocation ratio of the six-stroke engine was set at ÃŽ ±H = 0.5. Injection flow rate of fuel was Qt4 = 0.50 KJ/cycle for the four-stroke engine and Qt6 = 0.68 KJ/cycle for the six stroke engine. For six stroke engine, it meant that the amount of 0.34KJ was supplied at each combustion process. At the viewpoint of combustion heat, 0.75 KJ/cycle of heat should be supplied for the six stroke engine to make the equivalence heat condition. However diesel fuel of 0.68 KJ/cycle was supplied here because of difficulties associated with methanol injection. Injection timing of the four-stroke engine was changed from 160 degrees (200BTDC) to 180 degrees (TDC). For six -stroke engine, the injection timing of the first combustion process was fixed to 165 degrees (15 °BTDC) or 174 degrees (6 °BTDC), and the second injection timing was changed from 520 degrees (2000 BTDC) to 540 degrees (TDC). [pic] Fig 15 Valve timing diagram four stroke engine Table 3. Detailed conditions of comparison between the four stroke and six stroke diesel engines and performance of engine | |Four Stroke |Six Stroke | |Engine Parameters |Diesel Engine |Diesel Engine | |Engine Speed Ne [rpm] |2007 |2016 | |Supplied combustion heat per cycle | | | |Qt [KJ/cycle] |0.50 |0.68 | |Supplied combustion heat per unit time Ht [KJ/s] | | | | |8.36 |7.62 | |Intake air flow per cycle | | | |Ma [mg/cycle] |358.7 |371.4 | |Injection quantity per cycle | | | |Mf [mg/cycle] |11.8 |16 | | | | | |Excess air ratio ÃŽ » |2.40 |1.83 | |Intake air flow per unit time | | | |Ma [g/cycle] |6.00 |4.16 | |Injection quantity per unit time | | | |Mf [g/sec] |0.197 |0.179 | |Brake torque Tb [N-m] |15.52 |15.28 | |Brake power Lb [KW] |3.26 |3.24 | |BSFC. b [ g / KW-h] |217.9 |520.3 | |IMEP Pi [Kgf / cm2] |5.94 |4.37 | |Indicated torque Ti [N-m] |19.10 |18.71 | |Indicated power Li [KW] |4.01 |3.75 | |ISFC bi [g / KW-h ] |177.2 |163.3 | Indicated torque of the six-stroke engine is almost same level with that of the four-stroke engine under various injection timings. NO concentration in exhaust gas of the six-stroke engine was lower than that of the four-stroke engine. NO emissions from both engines were reduced by the retard of injection timing. The effect of retard in the second injection timing of the six-stroke engine was similar to that of the retard in the four-stroke engine. For the six-stroke engine, from the comparison between Xi I = 165 degrees (15 °BTDC) and Xi I = 174 degrees (6 °BTDC), it seemed that the NO reduction effect appeared with the timing retard in the first combustion process. Soot emission in the exhaust gas of the four-stroke engine was low level and it was not affected by the timing retard of injection. However, the level of soot emission from the six-stroke engine was strongly affected by the timing of the second injection. When the injection timing was advanced from 528 degrees (12 ° BTDC), it was confirmed that the soot emission was lower than that of the four-stroke engine. From numerical analysis, it was considered that the soot formed in the first combustion process was oxidized in the second combustion process. On the contrary, when the injection timing was retarded from 528 degrees (12 ° BTDC), soot emission increased with the timing retard. Then, it was considered that the increased part of the soot was formed in the second combustion process because an available period for combustion was shortened with the retard of injection timing. Experimental conditions were Xi = Xi I = 170 degrees (100 BTDC) and XiII=530 degrees (100 BTDC). The heat allocation ratio of six stroke engine was ÃŽ ±H=0.5. The cylinder temperature and heat release rate were calculated from the cylinder pressure. The pattern of heat release rate in the first combustion stroke of the six-stroke engine was similar to that of the heat release rate of the four-stroke engine. It was the typical combustion pattern that contained a pre-mixed combustion and diffusion combustion. On the other hand, since an increase of cylinder temperature in the second combustion process was caused by the compression of the burned gas formed in the first combustion stroke, a pre-mixed combustion in the second combustion process was suppressed by a short ignition delay. The maximum cylinder temperature in the first combustion process was lower than that in the four-stroke engine. It was caused by smaller amount of fuel which was injected in the first combustion process. Considering these results, it was proved that NO concentration in the exhaust gas was reduced by the decrease of the maximum cylinder temperature in the first co mbustion process and EGR effect in the second combustion process. The performance of these two engines could be compared by Table. 3. Since BSFC of the six-stroke engine obtained by the brake power suffered, SFC is compared with ISFC for the Xi = 163 degree (170 BTDC), ISFC of the four-stroke engine was 177.2 g/KW-h. On the other hand, for the Xi I = 165 degrees (15 ° BTDC) and Xi II = 523 degrees (170 BTDC), I.S the six-stroke engine was 163.3 g/KW-h. i.e. ISFC of the six-stroke engine was slightly lower than that of the four-stroke engine. It was considered that this advantage in ISFC was caused by a small cut-off ratio of constant pressure combustion. Because, in the six-stroke engine proposed here, the fuel divided into two combustion processes resulted in a short combustion period of each combustion process. Furthermore, in the reduction of NO emission, the six-stroke engine was superior to the four-stroke engine. 3.3.2 Effect of heat allocation ratio Injection conditions were Xi I = 170 degrees (1000 BTDC) and Xi II = 530 degrees (100 BTDC). Both fuels in the first and second combustion processes were diesel fuel. Total fuel at the combustion heat basis was Qt = 0.68 KJ/cycle. It meant a high load in this engine because the total excess air ratio was 1.83 as previously shown in Table 3. The maximum value of the indicated torque appeared around ÃŽ ±H = 0.5 NO concentration in exhaust gas was reduced by an increase of heat allocation ratio. In other words, NO emission decreased with an increase of the fuel of the second combustion process. In the case of ÃŽ ±H = 0.5, there is a relatively long ignition delay in the first combustion process and pre-mixed combustion was the main combustion phenomena in it. NO of high concentration was formed in this pre-mixed combustion process. On the other hand, in the case of ÃŽ ±H = 1, diffusion combustion was the main combustion phenomena and NO emission was low. Soot emission in exhaust gas increased with an increase of heat allocation ratio. Since the injection flow rate in the second combustion process increased with an increase of the heat allocation ratio, the injection period increased with an increase of the heat allocation ratio. It caused the second combustion process to be long, and unburnt fuel that was the origin of soot remained after the second combustion process. The heat release rates on ÃŽ ±H = 0.15 and ÃŽ ±H = 0.85. For ÃŽ ±H =0.15, since injection flow rate in the first combustion process was high and injection period in it was long, the combustion period in the first combustion process became long as compared with case of ÃŽ ±H = 0.85. On the other hand, for ÃŽ ±H = 0.85, the combustion period in the second combustion process became long as compared with case of ÃŽ ±H=0.15. It was also observed that the long combustion periods in both the first and second combustion were caused by the long diffusion combustion. Further, diffusion combustion was the main combustion phenomena of the second combustion process. When the heat allocation ratio was 0.85, the ratio of heat release rates between the first and second combustion should be 15: 85, however the actual ratio obtained from the figure was 46: 54. This inconsistency was caused from the drift of the base lines of the heat release diagrams. For ÃŽ ±H = 0.15, the actual ratio of heat release rates was 73: 27 with the similar reason. The cylinder temperature for the ÃŽ ±H = 0.15 condition was higher than that of the ÃŽ ±H = 0.85 condition. This could be explained as follows. In the first combustion stroke, since the injection flow rate of ÃŽ ±H = 0.15 was higher than that of ÃŽ ±H = 0.85, the combustion temperature for the ÃŽ ±H = 0.15 condition was higher than that of ÃŽ ±H = 0.85. In the second compression stroke, since the high temperature burned gas was re-compressed, the temperature of ÃŽ ±H = 0.15 was also higher than that of ÃŽ ±H = 0.85. As a result, the temperature at the beginning of the second combustion stroke was high in ÃŽ ±H = 0.15 condition as compared with ÃŽ ±H = 0.85 condition. At the later stage of the second combustion, however, the opposite relationship between these two temperatures were observed, because the injection flow rate of the second combustion process was low in ÃŽ ±H = 0.15 condition. The maximum temperatures in the first and second combustion process decreased with an increase of the heat allocation ratio. Then, it could be concluded that the reduction of NO concentration with the heat allocation ratio, was caused by the decrease of the cylinder temperature. 3.4 Performance of the dual fuel six stroke diesel engine 3.4.1 Comparison with diesel fuel six stroke engine Operating conditions of comparison between the diesel fuel and the dual fuel six-stroke engines are shown in Table. 4. Experimental conditions were Xi I= 170 degrees (100 BTDC), Xi II = 530 degrees (10o BTDC) and ÃŽ ±H = 0.5. In dual fuel six-stroke engine, diesel fuel and methanol were supplied into first and second combustion process, independently. Combustion heats supplied per one cycle of the diesel fuel and dual fuel six-stroke engines were same. The combustion heat supplied per one cycle was selected as Qt = 0.43 KJ/cycle under the middle load condition. Performance of the dual fuel six-stroke engine was compared with the diesel fuel six-stroke engine under various injection timings in the second combustion process. Indicated torques of both engines was revealed constant around 15 N-m. As a result, it could be concluded that states of combustion of the diesel fuel and the dual fuel six-stroke engines had similar contributions on the engine performance. NO emissions from the dual fuel six-stroke engine were lower than those of the diesel fuel six-stroke engine. This effect appeared prominently at the advanced injection timing of the second combustion. Further, NO concentrations of both engines were reduced by the injection timing retard in the second combustion. [pic] Fig 16 Torque- Angle diagram for six stroke engine Soot emission in the exhaust gas of the diesel fuel six stroke engines increased with a retard of the injection timing in the second combustion. For the dual fuel six-stroke engine, the exhaust level of soot was very low under various injection timings of the second combustion process. Soot was formed clearly by the combustion of diesel fuel in the first combustion process and it was oxidized in the second combustion process. Considering these results, it was possible to estimate that soot was almost oxidized by methanol combustion in the second combustion process. This estimation is supported by a dual fuel diesel engine operated with diesel fuel methanol. The combustion heat supplied per one cycle was selected as Qt = 0.68 KJ/cycle under the high load condition. Indicated torques of both engines was also revealed constant around 20 N-m. NO concentration had the same tendency as the cases of the middle load. Soot emission level of the diesel fuel six-stroke engine was high in this high load condition. For the dual fuel six-stroke engine, however, soot was very low under various injection timings of the second combustion process. The performance of these engines was compared in Table. 4. For the second combustion process, since combustion heats of diesel fuel and methanol were different, injection quantities of both engines were different. BSFC and ISFC of the dual fuel six-stroke engine was sensibly higher than that of the diesel fuel engine. To compare the performance of these engines, injection quantity of both engines was defined by an amount of combustion heat, and SFC should be calculated from it. As a result, indicated specific heat consumption of the diesel fuel six-stroke engine was 5.59 MJ/KW-h, and that of the dual fuel six-stroke engine was 5.43 MJ/KW-h. For the high load conditions shown in Table. 5, the similar advantage of the dual fuel six-stroke engine was observed. Table 4. Detailed conditions of comparison between the diesel fuel and dual fuel diesel engines and performance of engines under ÃŽ ±H = 0.5 and middle load | |Diesel Fuel Six Stroke Diesel |Dual Fuel Six Stroke | | |Engine |Diesel Engine | |Engine Speed Ne [rpm] |2016 |2003 | |Supplied combustion heat per cycle | | | |Qt [KJ/cycle] |0.43 | | |Injection quantity per cycle |5.0 | | |(First Combustion Stroke) |(Diesel Fuel) | | |Mf1 [mg/cycle] | | | |Injection quantity per cycle |5.0 |10.7 | |(Second Combustion Stroke) |(Diesel Fuel) |(Methanol) | |Mf2 [mg/cycle] | | | |Excess air ratio ÃŽ » |2.98 |3.15 | |Brake torque Tb [N-m] |3.14 |3.14 | |Brake power Lb [KW] |0.66 |0.66 | |B.S.F.C. b [ g / KW-h] | 610.9 |952.9 | |I.M.E.P. Pi [Kgf / cm2] |3.43 |3.53 | |Indicated torque Ti [N-m] |16.70 |15.12 | |Indicated power Li [KW] |3.1 |2.77 | |I.S.F.C. bi [g / KW-h ] |130.1 |198.4 | |Indicated specific heat consumption | | | |bi’ [MJ /KW-h] |5.59 |5.43 | In order to confirm the advantage of dual fuel six-stroke engine, the performance of these engines was compared with four-stroke engine as shown in Table. 6. NO concentrations of the diesel fuel and the dual fuel six-stroke engines were improved with 85 90% as compared with that of the four-stroke engine. Soot emission of the diesel fuel six-stroke engine was much higher than that of the four-stroke engine. However, for the dual fuel six-stroke engine, soot level was very low. Furthermore, the indicated specific heat consumption of the diesel fuel and dual fuel six-stroke engine were lower than that of the four-stroke engine. Especially, for the dual fuel six-stroke engine, the indicated specific heat consumption was improved with 15% as compared with that of the four stroke engine. From these results, it could be confirmed that the dual fuel six-stroke engine was superior to the diesel fuel six-stroke engine, and also it was superior to the four-stroke engine. Table 6. Percentage improvements of exhaust emission and specific heat consumption | |Four Stroke Diesel | Six Stroke Diesel Engine|Dual Fuel Six Stroke Engine | | |Engine | | | |NO [ppm] | |113 |90.5 | |( % improvement) |768 |(85.3%) |(88.2%) | |Soot [%] | |28.8 |0 | |(%improvement) |6.8 |(- 323.5%) |(100%) | |Indicated specific heat consumption bi’ | | | | |[MJ/KW-h] |7.51 |6.61 |6.37 | |(% improvement) | |(12.0%) |(15.2%) | Table 5. Detailed conditions of comparison between the diesel fuel and dual fuel diesel engine and performance of engines under ÃŽ ±H =0.5 and high load | | Six Stroke Diesel Engine |Dual Fuel Six Stroke Engine | |Engine Speed Ne [rpm] |2016 |2006 | |Supplied combustion heat per cycle | | | |Qt [kJ/cycle] |0.68 | | |Injection quantity per cycle |8.0 | | |(First Combustion Stroke) |(Diesel Fuel) | | |Mf1 [mg/cycle] | | | |Injection quantity per cycle |8.0 |17.2 | |(Second Combustion Stroke) |(Diesel Fuel) |(Methanol) | |Mf2 [mg/cycle] | | | |Excess air ratio ÃŽ » |1.86 |1.93 | |Brake torque Tb [N-m] |6.18 |6.08 | |Brake power Lb [kW] |1.52 |1.5 | |B.S.F.C. b [ g / kW.h] | 504.0 |777.7 | |I.M.E.P. Pi [kgf / cm2] |4.56 |4.75 | |Indicated torque Ti [N-m] |21.68 |20.38 | |Indicated power Li [kW] |3.45 |2.98 | |I.S.F.C. bi [g / kW.h ] |155.5 |236.2 | |Indicated specific heat consumption | | | |bi’ [MJ /kW.h] |6.61 |6.37 | 3.4.2 Effect of injection timing Performance of the dual fuel six-stroke engine under various injection timings in the second combustion process was investigated on middle and high load. Experimental conditions were Xi I = 170 degrees (100 BTDC) and ÃŽ ±H = 0.5. Performance of the dual fuel six-stroke engine under both load conditions had the similar tendency with the timing retard. NO concentrations in the high load condition were higher than those of the middle load condition. However, soot emission levels of both load conditions were extremely low under various injection timings of the second combustion. 3.4.3 Effect of heat allocation ratio Performance of the dual fuel six-stroke engine under various heat allocation ratios was investigated on middle and high load. Injection conditions were Xi I = 170 degrees (100 BTDC) and Xi II = 530 degrees (100 BTDC). Since the combustion heat of methanol was low, experimental range of heat allocation ratio was limited by the smooth operation of the engine. Only the range from ÃŽ ±H = 0.25 to 0.75 (on Qt = 0.43 KJ/cycle), and from ÃŽ ±H = 0 to 0.5 (on Qt = 0.68 KJ/cycle) could be tested.. Indicated torque increased with an increase of the heat allocation ratio. NO concentration in exhaust gas was reduced with an increase of the heat allocation ratio. Soot was very low, irrespective of the methanol flow rate. Even if the load condition was high, it was concluded that soot was practically eliminated by a small amount of methanol in the second combustion process (8% of total fuel). 4. ADVANTAGES OF SIX STROKE OVER FOUR STROKE ENGINES The six stroke is thermodynamically more efficient because the change in volume of the power stroke is greater than the intake stroke, the compression stroke and the Six stroke engine is fundamentally superior to the four stroke because the head is no longer parasitic but is a net contributor to – and an integral part of – the power generation within exhaust stroke. The compression ration can be increased because of the absent of hot spots and the rate of change in volume during the critical combustion period is less than in a Four stroke. The absence of valves within the combustion chamber allows considerable design freedom. 4.1 Main advantages of the duel fuel six-stroke engine: 4.1.1 Reduction in fuel consumption by at least 40%: An operating efficiency of approximately 50%, hence the large reduction in specific consumption. the Operating efficiency of current petrol engine is of the order of 30%. The specific power of the six-stroke engine will not be less than that of a four-stroke petrol engine, the increase in thermal efficiency compensating for the issue due to the two additional strokes. 4.1.2 Two expansions (work) in six strokes: Since the work cycles occur on two strokes (3600 out of 10800 ) or 8% more than in a four-stroke engine (1800 out of 720 ), the torque is much more even. This lead to very smooth operation at low speed without any significant effects on consumption and the emission of pollutants, the combustion not being affected by the engine speed. These advantages are very important in improving the performance of car in town traffic. 4.1.2 Dramatic reduction in pollution: Chemical, noise and thermal pollution are reduced, on the one hand, in proportion to the reduction in specific consumption, and on the other, through the engine’s own characteristics which will help to considerably lower HC, CO and NOx emissions. Furthermore, it’s ability to run with fuels of vegetable origin and weakly pollutant gases under optimum conditions, gives it qualities which will allow it to match up to the strictest standards. 4.1.3 Multifuel: Multifuel par excellence, it can use the most varied fuels, of any origin (fossil or vegetable), from diesel to L.P.G. or animal grease. The difference in inflammability or antiknock rating does not present any problem in combustion. It’s light, standard petrol engine construction, and the low compression ration of the combustion chamber; do not exclude the use of diesel fuel. Methanol-petrol mixture is also recommended. 5. CONCLUSIONS The performance of the dual fuel six-stroke engine was investigated. In this dual fuel engine, diesel fuel was supplied into the first combustion process and methanol was supplied into the second combustion process where  the burned gas in the first combustion process was re-compressed. The results are summarized as follows. 1. Indicated specific fuel consumption (ISFC.) of the six-stroke engine proposed here is slightly lower than that of the four-stroke engine (about 9% improvement). NO and soot emissions from the six-stroke engine was improved as compared with four-stroke engine under advanced injection timings in the second combustion stroke. 2. For the dual fuel six-stroke engine, the timing retard and an increase of heat allocation ratio in the second combustion stroke resulted in a decrease of the maximum temperatures in the combustion processes. It caused the reduction of NO emission. 3. For the dual fuel six-stroke engine, soot was practically eliminated by a small amount of methanol in the second combustion process. 4. From the comparison of the performance between the dual fuel six-stroke and the four-stroke engine, it was concluded that indicated specific heat consumption of the dual fuel six-stroke engine was improved with 15% as compared with the four-stroke engine. NO concentration of the dual fuel six-stroke engine was improved with 90%. Furthermore, soot emission was very low in the dual fuel six-stroke engine. 5. As the fuel in one cycle was divided into two combustion processes and the EGR effect appeared in the second combustion process, the decreased maximum cylinder temperature reduced NO concentration in the exhaust gas It was further confirmed that soot formed in the first combustion process was oxidized in the second combustion process .Therefore, a six stroke DI diesel engine has significant possibilities to improve combustion process because of its more controllable factors relative to a conventional four-stroke engine. Considering these results, it was confirmed that the dual fuel six-stroke engine was superior to the four-stroke engine. 6. REFERENCES 1. Tsunaki Hayasaki, Yuichirou Okamoto, Kenji Amagai and Masataka Arai â€Å"A Six-stroke DI Diesel Engine under Dual Fuel Operation â€Å"SAE Paper No 1999-01-1500 2. S.Goto and K.Kontani, A Dual Fuel Injector for Diesel Engines, SAE paper, No. 851584, 1985 3. â€Å"Internal Combustion Engines â€Å"A book by Mathur Sharma. 4. â€Å"Internal Combustion Engines† Tata McGraw-hill publications, Author V Ganesan 7. NOMENCLATURE Ne : Engine speed X : Crank angle Xi : Injection timing of the four-stroke diesel engine ÃŽ ±H : Heat allocation ratio Q : Supplied combustion heat Qt : Supplied combustion heat per cycle P : Cylinder pressure V : Cylinder volume Vs : Stroke volume Pi : Indicated mean effective pressure (LM.E.P) Ti : Indicated torque Li : Indicated power Tb : Brake torque Lb : Brake power Ht : Supplied combustion heat per unit time Ma : Intake air flow per cycle Ma : Intake air flow per unit time Mf : Injection quantity per cycle Mi : Injection quantity per unit time ÃŽ » : Excess air ratio b : Brake specific fuel consumption (B.S.F.C.) bl : Indicated specific fuel consumption (I.S.F.C.) bi : Indicated specific heat consumption SUBSCRIPTS I: first combustion stroke II: second combustion stroke 4: four-stroke diesel engine 6: six-stroke diesel engine