EP2683920B1 - Moteur à combustion interne - Google Patents

Moteur à combustion interne Download PDF

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Publication number
EP2683920B1
EP2683920B1 EP11791343.4A EP11791343A EP2683920B1 EP 2683920 B1 EP2683920 B1 EP 2683920B1 EP 11791343 A EP11791343 A EP 11791343A EP 2683920 B1 EP2683920 B1 EP 2683920B1
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EP
European Patent Office
Prior art keywords
valve seat
crankshaft
exhaust
inlet
degrees
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Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
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EP11791343.4A
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German (de)
English (en)
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EP2683920A1 (fr
Inventor
Melkote Viraraghavachar NARASIMHAN
Varadha Iyengar LAKSHMINARASIMHAN
Varadarajan Ranganathan RAJAGOPALAN
K Kumar
Vythilingam KARUNAHARAN
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Indian Institute of Science IISC
TVS Motor Co Ltd
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Indian Institute of Science IISC
TVS Motor Co Ltd
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Publication of EP2683920A1 publication Critical patent/EP2683920A1/fr
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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01L—CYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L3/00—Lift-valve, i.e. cut-off apparatus with closure members having at least a component of their opening and closing motion perpendicular to the closing faces; Parts or accessories thereof
    • F01L3/22—Valve-seats not provided for in preceding subgroups of this group; Fixing of valve-seats
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B25/00—Engines characterised by using fresh charge for scavenging cylinders
    • F02B25/14—Engines characterised by using fresh charge for scavenging cylinders using reverse-flow scavenging, e.g. with both outlet and inlet ports arranged near bottom of piston stroke
    • F02B25/145—Engines characterised by using fresh charge for scavenging cylinders using reverse-flow scavenging, e.g. with both outlet and inlet ports arranged near bottom of piston stroke with intake and exhaust valves exclusively in the cylinder head
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B33/00—Engines characterised by provision of pumps for charging or scavenging
    • F02B33/02—Engines with reciprocating-piston pumps; Engines with crankcase pumps
    • F02B33/04—Engines with reciprocating-piston pumps; Engines with crankcase pumps with simple crankcase pumps, i.e. with the rear face of a non-stepped working piston acting as sole pumping member in co-operation with the crankcase
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
    • F02M67/00—Apparatus in which fuel-injection is effected by means of high-pressure gas, the gas carrying the fuel into working cylinders of the engine, e.g. air-injection type
    • F02M67/005—Apparatus in which fuel-injection is effected by means of high-pressure gas, the gas carrying the fuel into working cylinders of the engine, e.g. air-injection type fuel-gas mixture being compressed in a pump for subsequent injection into the engine
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
    • F02M69/00—Low-pressure fuel-injection apparatus ; Apparatus with both continuous and intermittent injection; Apparatus injecting different types of fuel
    • F02M69/10—Low-pressure fuel-injection apparatus ; Apparatus with both continuous and intermittent injection; Apparatus injecting different types of fuel peculiar to scavenged two-stroke engines, e.g. injecting into crankcase-pump chamber
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B75/00—Other engines
    • F02B75/02—Engines characterised by their cycles, e.g. six-stroke
    • F02B2075/022—Engines characterised by their cycles, e.g. six-stroke having less than six strokes per cycle
    • F02B2075/025—Engines characterised by their cycles, e.g. six-stroke having less than six strokes per cycle two

Definitions

  • IC internal combustion
  • IC engines find a variety of applications, for example, in automobiles, generators, marine applications, etc.
  • the IC engines may be classified based on the number of strokes per working cycle of the IC engine. For example, IC engines are classified as two-stroke engines or four-stroke engines.
  • various phases of the working cycle i.e., intake, compression, combustion and expansion, and exhaust
  • intake, compression, combustion and expansion, and exhaust are accomplished in two strokes of a piston by performing the intake and exhaust strokes simultaneously during the end of the expansion, and the beginning of the compression strokes, respectively.
  • the conventional two stroke engines are usually provided with ports in a cylinder wall. Due to layout of the ports in the cylinder wall and the simultaneous intake and exhaust strokes, a mixing of the charge and the combustion products may occur in the cylinder bore. Further, because of the overlapping intake and exhaust strokes, some amount of the charge may leak out from the cylinder bore during the exhaust stroke.
  • conventional four-stroke engines complete the phases of the working cycle in four strokes of the piston.
  • the four-stroke engines have the ports provided in a cylinder head and have valves provided at the ports to regulate opening and closing of the ports.
  • the ports and valves of the conventional four-stroke engines are associated with certain flow characteristics, which are prone to deficient flow.
  • the flow characteristics of the ports and valves of the conventional four-stroke engines differ significantly from the flow characteristics of piston controlled ports of two-stroke engines.
  • JPH01200013A focuses on increasing the flowing quantity of new air being supplied to the combustion chamber and thus improving the output, by setting the internal diameter of each recessed groove formed at the inner wall surface of a cylinder head, at a particular size.
  • the cylinder head has a plurality of inner wall surface portions which are respectively flat, and a pair of cylindrical recessed grooves are formed at one inner wall surface portion. Valve seating are respectively formed at the insides of cylindrical recessed grooves. Similarly air sucking valves are provided. The assembly is so set that the substantial supply of new air is started when each umbrella portion of each air sucking valve has passed out of each recessed groove.
  • DE1813906A1 discloses an exhaust valve arranged in the cylinder head of the two-stroke engine as a gas outlet in the cylinder chamber.
  • the exhaust valve is provided by control cam having wide-ranging control timing.
  • a valve lever is provided controlled by the cam.
  • the fresh air inlet is arranged at the stroke end of the cylinder as usual.
  • US4907544A discloses a two-stroke cycle internal combustion engine without crankcase scavenging, which is able to operate in a four-stroke mode at cranking and idle speeds.
  • the engine utilizes a turbocharger driven by exhaust gases to recharge the cylinder with fresh air or an air/fuel mixture.
  • An auxiliary inlet valve is responsive to pressure within the cylinder and enables the engine to operate in a four-stroke mode, thus eliminating the need for an externally driven air pump when the engine is either idling, being started, or under light loads where turbocharger boost is too low to supply sufficient flow of air or an air/fuel mixture.
  • US1963780A discloses an internal-combustion-engine power unit including a four-cycle engine, an exhaust-gas driving turbine and an air-compressor driven thereby.
  • a valve-controlled inlet opening at the upper end of the cylinder is adapted to admit part of the air during part of the suction stroke of the engine by the pumping-action of the engine-piston.
  • Ports in the cylinder-wall is adapted to be uncovered by the piston near the end of the suction stroke for introducing the rest of the air only at the end of the suction stroke and the beginning of the compression stroke.
  • Means for prematurely closing some of the air-admission members during the suction-action of the piston near the end of the working stroke is provided.
  • GB861080A discloses an internal combustion engine with direct fuel injection into a combustion chamber.
  • the engine the form of a body of revolution arranged centrally in the working piston and which is so intersected by the top surface of the piston that no throttling action or only a slight throttling action is produced at the overflow opening.
  • Means is provided at the air-admission valve for producing rotation of the air about the axis of the combustion chamber.
  • the air-admission valve is countersunk or provided with a cylindrical portion above or below its seating surface and is controlled by a cam whose highest section extends over an angle of rotation of the cam of at least 30' so that the valve remains in its position of maximum opening over this angle.
  • the subject matter described herein relates to an internal combustion engine.
  • the internal combustion engine includes at least one port and a valve provided at the port to regulate opening and closing of the port.
  • a valve seat is provided at the port for resting the valve.
  • a valve seat rim is provided at the valve seat for delaying the opening and advancing the closing of the port.
  • IC engines usually include ports for intake of charge and for the expulsion of combustion products during a combustion cycle of the IC engine.
  • Conventional two stroke IC engines having ports in a cylinder wall for intake of charge and for expulsion of the products of combustion from a cylinder bore, are susceptible to thermal loading and distortions of the ports and of the cylinder wall. Such distortions may lead to high friction between a piston and the cylinder wall during movement of the piston in the cylinder bore.
  • these engines have the benefit of faster opening and closing of ports within a short duration resulting in effective gas exchange from the cylinder, i.e. the rate of opening and closing of the ports is high, which provides for the gas exchange from the cylinder.
  • conventional four stroke IC engines have valves provided in the cylinder head for the induction of the charge, and for the expulsion of combustion products.
  • the rate of opening and closing of ports is relatively gradual as compared to the opening and closing of the ports in two-stroke engine, for the same duration.
  • it is difficult to ensure effective gas exchange in the available duration.
  • the usage of conventional valve train features and drive mechanisms of the typical four-stroke engine in the two-stroke engine to control the port area opening compromises the performance of the two-stroke engine due to constraints of valve train dynamics and the short available durations for gas exchange.
  • an internal combustion (IC) engine interchangeably referred to as an engine hereinafter, disclosed herein will help address the aforementioned drawbacks in addition to providing several other advantages over the existing two stroke engines.
  • the IC engine includes one or more ports provided in the cylinder head.
  • the cylinder head includes at least one inlet port for intake of charge into the cylinder bore, and at least one exhaust port for discharge of combustion products from the cylinder bore.
  • Each of the ports is provided with a valve to control the opening and closing of the ports. Further, the ports are provided with a valve seat, corresponding to each valve. The valve lifts and rests at the valve seat during an open and a closed position, respectively, of the port.
  • a valve seat rim is provided at at least one valve seat in the cylinder head.
  • the valve seat rim is provided at a circumference of the valve seat and is in the form of a cylindrical extension in the direction of the longitudinal axis of the valve.
  • valve seat rim delays the opening of the port and advances the closing of the ports by the valves.
  • valve seat rim is provided at an inlet valve seat, then the delayed opening and early closing of an inlet port provides a short duration for which the inlet port remains open. A shorter duration and rapid opening of the inlet port allows effective gas exchange from the combustion chamber in the available time.
  • valve seat rim may be associated with the seat of the exhaust valve.
  • a delayed opening of the exhaust port is achieved. The delay in the opening of the exhaust valve facilitates the engine in achieving a higher expansion ratio, and hence, higher fuel efficiency of the engine.
  • valve seat rim at the valve seat facilitates in achieving a high rate of increase of effective flow area between the valve and the valve seat when the seat rim is uncovered by the valve.
  • the rapid opening of the port area facilitates substantial increase in flow, for example, at the inlet valve, thereby overcoming the limitation of conventional valve train dynamics.
  • the actual valve lift durations are high but the effective lift duration is made shorter by this delay feature.
  • the port is not uncovered due to the valve rim and, hence, the actual flow of air does not occur.
  • the opening of the port takes place during the higher rate of valve lift, which helps efficient gas exchange. The same is true during the port closure.
  • Fig. 1 illustrates a sectional view of an internal combustion engine 100, according to an embodiment of the present subject matter.
  • the internal combustion engine 100 is a two stroke internal combustion engine.
  • the internal combustion engine 100 referred to as engine 100 hereinafter, includes a crankcase 102 connected to a charging device (not shown in figure).
  • the crankcase 102 houses a crankshaft 104.
  • a cylinder block 106 having a cylinder bore 108 is mounted on the crankcase 102.
  • the crankcase 102 includes an induction valve 110, for example, a reed valve.
  • the induction valve 110 allows a unidirectional induction of a scavenging fluid, such as fresh air or a lean composition of charge formed by mixture of air and fuel, into the crankcase 102.
  • the cylinder block 106 includes the induction valve 110 to induct the scavenging fluid into the crankcase 102.
  • the scavenging fluid entering the crankcase 102 is inducted into the cylinder bore 108.
  • some lubricant particles may be entrained along with the scavenging fluid.
  • a filter element (not shown in the figure) is provided in the cylinder block 106 such that the scavenging fluid inducted from the crankcase 102 into the cylinder bore 108 passes through the filter element before entering the cylinder bore 108.
  • the filter element filters any lubricant particles in the scavenging fluid in the crankcase 102 and prevents the entry of lubricant particles into the cylinder bore 108 that may otherwise be burnt during combustion Hence, the provision of the filter element facilitates low pollutants in exhaust emissions from the engine 100.
  • the filter element may be integrated with the induction valve 110, and the scavenging fluid is inducted into the crankcase 102 through the induction valve 110 and the filter element.
  • the stream of the incoming scavenging fluid may clean the filter element by drawing the lubricant particles in the filter element back into the crankcase 102. This may ensure longevity of the filter element.
  • the filter element may be provided separately from the induction valve 110 and in different locations, such as the crankcase 102 or in a passage connecting the crankcase 102 to the cylinder bore 108.
  • the cylinder bore 108 has a piston 112 disposed therein, such that the piston 112 is capable of reciprocating inside the cylinder bore 108. It may be understood that the engine 100 may have more than one cylinder bore 108. Further, the piston 112 is connected to the crankshaft 104 through a connecting rod 114 to drive the crankshaft 104. Inside the cylinder bore 108, the piston 112 has two extreme positions - a top dead centre (TDC) position when the engine 100 has completed a compression stroke and a bottom dead centre (BDC) position from where the piston 112 commences actuation at the beginning of the compression stroke.
  • TDC top dead centre
  • BDC bottom dead centre
  • the cylinder bore 108 has a plurality of transfer ports 116 provided annularly along a periphery of a cylinder wall 118.
  • the cylinder wall 118 of the cylinder block 106 defines the cylinder bore 108 therein.
  • the transfer ports 116 are disposed in the cylinder wall 118 in such a way that the transfer ports 116 are closer to the BDC position than they are to the TDC position of the piston 112.
  • Each of the plurality of transfer ports 116 is connected to the crankcase 102 through an induction passage 120. Through the transfer ports 116, the scavenging fluid is inducted into the cylinder bore 108 from the crankcase 102 during movement of the piston 112 from the TDC position to the BDC position.
  • each of the transfer ports 116 is formed in the shape of a truncated cone, for example, a right cone or an oblique cone, having an apex angle ⁇ (not shown in the figure) and having a base at an opening of the induction passage 120 into the respective transfer port 116.
  • the apex angle ⁇ may be defined as the included angle of the lateral surface of the cone, the angle being measured on a plane passing through an imaginary apex and the base of the truncated cone.
  • the apex angle is about 10 to 30 degrees to provide a swirling motion to the scavenging fluid entering the cylinder bore 108.
  • an axis of each of the transfer ports 116 is inclined to the cylinder wall 118 of the cylinder bore 108 at an angle, the angle being measured in a horizontal plane.
  • the angle between the axes of the transfer ports 116 and the wall of the cylinder bore 108 may also be measured in a vertical plane.
  • each of the plurality of transfer ports 116 formed as a truncated cone, includes a first aperture (not shown in the figure) and a second aperture (not shown in the figure).
  • the first aperture is provided at the cylinder wall 118 and the second aperture is provided at an opening of the induction passage 120 into the respective transfer port 116.
  • the diameter of the first aperture is smaller than a diameter of the second aperture.
  • the inclination and geometry of the transfer ports 116 with respect to the wall of the cylinder bore 108 provides a swirling motion of the scavenging fluid during the induction of the scavenging fluid into the cylinder bore 108.
  • the swirling motion of the scavenging fluid in the cylinder bore 108 facilitates in scavenging and purging the cylinder bore 108 of combustion products during the movement of the piston 112 from the BDC position to the TDC position. Further, opening and closing of the transfer ports 116 is controlled by the movement of the piston 112 in the cylinder bore 108.
  • a plurality of piston rings 122 are provided on the piston 112 for providing a seal between the piston 112 and the cylinder bore 108.
  • the piston rings 122 also facilitate scraping carbonaceous deposits off an inner wall of the cylinder bore 108.
  • a cylinder head assembly 124 is mounted on the cylinder block 106 to close one end of the cylinder bore 108.
  • the cylinder head assembly 124 includes a cylinder head 126 and a cylinder head cover 128.
  • a main combustion chamber 130 is formed between the cylinder head 126, the cylinder wall 118, and a crown of the piston 112 when the piston 112 is at the TDC position inside the cylinder bore 108 during a combustion cycle.
  • the cylinder head 126 includes an inlet port 132. In other embodiments, the cylinder head 126 may have more than one inlet port 132.
  • the cylinder head 126 has an exhaust port 134 for expelling combustion products from the main combustion chamber 130.
  • the cylinder head 126 may have more than one exhaust port 134.
  • the inlet port 132 is smaller in diameter than the exhaust port 134.
  • a diameter of the inlet port 132 is about 10 millimeter (mm) and a diameter of the exhaust port 134 is about 27.5 millimeter (mm).
  • the large size of the exhaust port 134 facilitates effective exhaust of the combustion products from the exhaust port 134 as well as scavenging of the combustion products from the cylinder bore 108.
  • the expulsion of the combustion products from the cylinder bore 108 is aided by the scavenging fluid entering the cylinder bore 108 through the transfer ports 116.
  • the provision of the exhaust port 134 in the cylinder head 126 of the engine 100 allows uniflow scavenging of the cylinder bore 108 by the scavenging fluid. Such a scavenging facilitates a thorough purging of the cylinder bore 108 and enhances fuel economy and efficiency of the engine 100.
  • an auxiliary combustion chamber 136 is formed in the cylinder head 126 of the engine 100.
  • the auxiliary combustion chamber 136 is provided with an ignition element 138 to achieve combustion of the charge in the auxiliary combustion chamber 136.
  • more than one ignition element 138 may be provided in the auxiliary combustion chamber 136 and one or more ignition elements 138 may be provided in the main combustion chamber 130, to facilitate the combustion of the charge.
  • a partial combustion, referred to as first combustion, of charge takes place in the auxiliary combustion chamber 136.
  • the auxiliary combustion chamber 136 opens into the main combustion chamber 130 such that during the expansion stroke, a flame front produced by the combustion of charge in the auxiliary combustion chamber 136 spreads into the main combustion chamber 130 and the combustion of charge is substantially completed in the main combustion chamber 130.
  • the substantially complete combustion of charge in the main combustion chamber 130 is interchangeably referred to as second combustion of charge.
  • the auxiliary combustion chamber 136 is about 70% of the total combustion chamber volume and the main combustion chamber 130 is about 30% of the total combustion chamber volume.
  • the transfer ports 116 are uncovered and the swirling scavenging fluid scavenges the combustion products. Further, during the movement of the piston 112 from the BDC to the TDC, the scavenging fluid is compressed in the cylinder bore 108.
  • the auxiliary combustion chamber 136 is in fluid communication with a fuel supply pump 140.
  • the fuel supply pump 140 is a reciprocating type pump and includes an auxiliary piston 142 reciprocating inside an auxiliary bore 144.
  • the auxiliary piston 142 has a small skirt length to reduce contact surface of the auxiliary piston 142, and hence friction between the auxiliary piston 142 and a wall of the auxiliary bore 144.
  • the auxiliary bore 144 is formed in an auxiliary cylinder block 146 of the fuel supply pump 140.
  • the auxiliary cylinder block 146 is formed in the cylinder block 106 of the engine 100, and hence the fuel supply pump 140 is integrated in the cylinder block 106.
  • the fuel supply pump 140 is integrated with the crankcase 102 of the engine 100.
  • the fuel supply pump 140 is integrated with the cylinder head 126 of the engine 100.
  • the fuel supply pump 140 further includes an auxiliary crankshaft 148 housed in an auxiliary crankcase 150.
  • the auxiliary piston 142 is operably coupled to the auxiliary crankshaft 148 through an auxiliary connecting rod 152 such that a rotational motion of the auxiliary crankshaft 148 is transformed into a reciprocatory motion of the auxiliary piston 142.
  • the auxiliary crankshaft 148 of the fuel supply pump 140 obtains a drive from the crankshaft 104 of the engine 100.
  • the auxiliary crankshaft 148 obtains the drive from the crankshaft 104 through a chain drive assembly (not shown in the figure).
  • other drive mechanisms may also be used to provide a drive from the crankshaft 104 to the auxiliary crankshaft 148.
  • an end of the auxiliary crankshaft 148 has an auxiliary sprocket wheel (not shown in the figure) mounted and fixed thereon.
  • a crankshaft sprocket wheel (not shown in the figure) is mounted on an end of the crankshaft 104 to provide a drive to the auxiliary sprocket wheel of the fuel supply pump 140.
  • a phase difference may be provided between the drive of the crankshaft 104 and the drive of the fuel supply pump 140 so that the induction of the charge into the auxiliary combustion chamber 136 is achieved substantially before the piston 112 reaches the TDC position in the cylinder bore 108.
  • the fuel supply pump 140 includes an auxiliary cylinder head (not shown in the figure) mounted on the auxiliary cylinder block 146.
  • the auxiliary cylinder head includes an auxiliary inlet port 154 and an auxiliary exhaust port 156.
  • the auxiliary inlet port 154 is connected to a fuelling device (not shown in the figure), such as a carburettor, and receives a charge of fuel and air from the fuelling device through a charging valve 158.
  • the charging valve 158 is a reed valve, which allows a unidirectional flow of the charge from the fuelling device into the auxiliary bore 144 of the fuel supply pump 140 through the auxiliary inlet port 154.
  • a small quantity of the charge is received by the fuel supply pump 140 from the fuelling device.
  • the charge is richer in composition as compared to the corresponding stoichiometric composition of the charge for an operation, for example, full load operation or part load operation, of the engine 100.
  • the auxiliary bore 144 of the fuel supply pump 140 includes a fuel inlet port (not shown in the figure) connected to the fuelling device for the induction of charge.
  • the fuel inlet port is formed in the auxiliary cylinder block 146 slightly above a BDC position of the auxiliary piston 142 in the auxiliary bore 144. The charge is inducted into the fuel supply pump 140 when the auxiliary piston 142 is approaching the BDC position in the auxiliary bore 144. Hence, the reciprocating motion of the auxiliary piston 142 regulates the opening and closing of the fuel inlet port for the induction of charge.
  • the charge entering the auxiliary bore 144 is pressurized by the reciprocating motion of the auxiliary piston 142.
  • the pressurization of the charge facilitates substantial atomization and disintegration of the charge before it is supplied into the auxiliary combustion chamber 136 for combustion. Such pressurization assists in complete burning of the charge and hence improves the efficiency of the engine 100.
  • the fuel supply pump 140 is a low-pressure pump and has a compression ratio of about 3:1.
  • the compression ratio may be understood as ratio of volume of auxiliary bore 144 when the auxiliary piston 142 is at the BDC position to the volume of the auxiliary bore 144 when the auxiliary piston 142 is at the TDC position.
  • a volumetric capacity of the fuel supply pump 140 is about 10% of a volumetric capacity of the engine 100.
  • the capacity of the fuel supply pump 140 is in a range of about 10 cubic centimeters (cc) to 15 cc.
  • the auxiliary exhaust port 156 is connected to an inlet passage 160 leading to the inlet port 132 of the engine 100.
  • the fuel supply pump 140 inducts a pressurized charge of air and fuel for combustion into the auxiliary combustion chamber 136 through the inlet passage 160 and the inlet port 132.
  • the inlet passage 160 is formed in the cylinder head 126 in substantial proximity of the main combustion chamber 130 and the auxiliary combustion chamber 136. In said embodiment, a substantially complete vaporization of the fuel in the pressurized charge is achieved in the inlet passage 160 because of the proximity to the combustion chambers 130 and 136.
  • the auxiliary exhaust port 156 includes a discharge valve (not shown in the figure) to regulate the induction of pressurized charge from the fuel supply pump 140 into the inlet passage 160.
  • the discharge valve is a pressure-regulated check valve and is calibrated in such a way that it allows the pressurized charge to enter the inlet passage 160 when the pressure of the charge is above a predetermined value. During the induction of charge from the fuelling device into the fuel supply pump 140 and during the pressurization of charge, the discharge valve remains closed.
  • the fuel supply pump 140 inducts the charge into the auxiliary combustion chamber 136 when the piston 112 is compressing the air inducted in the main combustion chamber 130 through the transfer ports 116.
  • the piston 112 is said to be in a compression stroke and is approaching the TDC position in the cylinder bore 108.
  • the main combustion chamber 130 has the scavenging fluid, such as fresh air or a lean composition of charge, while the auxiliary combustion chamber 136 has a small quantity of relatively rich composition of the charge as compared to that in the main combustion chamber 130 creating layers or strata of the charge with different compositions of air and fuel.
  • a plurality of strata of charge is formed in the main combustion chamber 130 and the auxiliary combustion chamber 136.
  • the different strata of charge differ in the composition of the air and fuel. This allows the engine 100 to operate on an overall lean composition of charge. Hence, the overall fuel consumption of the engine 100 is low.
  • the engine 100 may include only the main combustion chamber 130 for the combustion of the charge.
  • the fuel supply pump 140 inducts charge into the main combustion chamber 130.
  • the cylinder head assembly 124 includes a valve train assembly 162.
  • the valve train assembly 162 is provided in the cylinder head 126 and housed inside the cylinder head assembly 124.
  • the valve train assembly 162 includes an inlet valve 164 and an exhaust valve 166.
  • the inlet valve 164 is provided at the inlet port 132 in the cylinder head 126
  • the exhaust valve 166 is provided at the exhaust port 134.
  • An inlet valve seat and an exhaust valve seat are provided at the inlet port 132 and the exhaust port 134, respectively.
  • the valves 164 and 166 rest at their respective valve seats during a closed position of the respective ports 132 and 134.
  • the inlet valve 164 and the exhaust valve 166 regulate the opening and closing of their respective ports, i.e., the inlet port 132 and the exhaust port 134, at which they are provided.
  • a valve seat rim (not shown in the figure) is provided at least at one of the exhaust valve seat or the inlet valve seat in the cylinder head 126.
  • the valve seat rim is cylindrical, provided at a circumference of the valve seat and extends in a direction of a longitudinal axis of the valve 164 or 166 towards the cylinder bore 108.
  • the valve seat rim is provided at the inlet valve seat as well as the exhaust valve seat.
  • valve seat rim is formed integrally with the cylinder head 126.
  • valve seat rim is formed as a bore in the cylinder head 126, extending to the valve seat.
  • valve seat rim may be formed separately from the cylinder head 126 and disposed at the valve seat.
  • valve seat rim at the valve seat facilitates in achieving a high rate of increase of an effective flow area between the valve and the valve seat, thereby achieving a substantially rapid lifting and seating of the valve. Further, the extension of the valve seat, as a result of the presence of the valve seat rim, delays the opening of the port and advances the closing of the ports by the valves.
  • the valve seat rim is described in further detail in Fig. 2 and 3 .
  • valve train assembly 162 includes a camshaft 168 to actuate the inlet valve 164 and the exhaust valve 166.
  • the inlet valve 164 is operatively coupled to the camshaft 168 via a first push rod 170 and a first rocker arm 172.
  • the exhaust valve 166 is operatively coupled to the camshaft 168 via a second push rod 174 and a second rocker arm 176 (partially shown in Fig. 1 ).
  • first rocker arm 172 and the second rocker arm 176 include a first follower 178 and a second follower (not shown in figure), respectively.
  • the first follower 178 and the second follower are roller type followers.
  • the inlet valve 164 and the exhaust valve 166 are spring loaded valves.
  • the camshaft 168 includes a plurality of cam lobes (not shown in the figure), corresponding to each of the inlet valve 164 and the exhaust valve 166, to actuate the valves 164 and 166.
  • the camshaft 168 may obtain a drive from the engine 100, for example, through a chain drive assembly.
  • a camshaft sprocket wheel (not shown in the figure) mounted on the camshaft 168 receives a drive from the crankshaft 104 through the chain drive assembly.
  • the chain drive assembly is a step-less chain drive assembly, that is, the crankshaft 104 provides a drive to the camshaft 168 for actuating the inlet valve 164 and the exhaust valve 166 and also provides a drive to the auxiliary crankshaft 148 for driving the fuel supply pump 140, without any reduction in the drive ratio.
  • This facilitates synchronization of the rotation of the crankshaft 104, the opening and closing of the inlet valve 164 and the exhaust valve 166, as well as the rotation of the auxiliary crankshaft 148 of the fuel supply pump 140, for effective functioning of the engine 100.
  • the chain drive assembly may include a plurality of chain tensioners (not shown in the figures).
  • the chain tensioners maintain proper tension in the chain to reduce vibrations and noise generated by the chain during operation.
  • the chain tensioners achieve a compensation for the wear that the chain undergoes and, hence, facilitates avoiding errors, such as tooth-skip, during the transmission of drive through the chain.
  • the chain tensioners also help in reducing synchronization errors between the crankshaft 104, the camshaft 168, and the auxiliary crankshaft 104, which may occur due to loss in tension of the chain.
  • the inlet valve 164 and the exhaust valve 166 may be electronically actuated valves.
  • Fig. 2 illustrates a magnified sectional view of the exhaust port 134 of the engine 100, according to an embodiment of the present subject matter.
  • an exhaust valve seat 202 is provided at the exhaust port 134.
  • the lifting and the seating of the exhaust valve 166 at the exhaust valve seat 202 regulate the opening and the closing of the exhaust port 134.
  • an exhaust valve seat rim 204 is formed at a circumference of the exhaust valve seat 202, extending along a longitudinal axis 206 of the exhaust valve seat 202 in the direction of the cylinder bore 108.
  • the exhaust valve seat rim 204 is cylindrical in shape.
  • the exhaust valve seat rim 204 may be formed integrally with the exhaust valve seat 202 in the cylinder head 126, or may be formed separately and disposed at the exhaust valve seat 202.
  • the exhaust valve seat rim 204 has a length of about 5% to about 15% of a valve lift of the exhaust valve 166.
  • the valve lift of a valve can be understood as the maximum distance, measured from the respective valve seat, traversed by the valve when lifted off the valve seat during operation.
  • the length of the exhaust valve rim 204 is about 1 millimeter (mm) measured from the exhaust valve seat 202.
  • the operation of the exhaust valve 166 may be understood with respect to Figs. 1 and 2 .
  • the actuation of the second follower of the second rocker arm 176 by the corresponding lobe of the camshaft 168 causes the second push rod 174 to actuate the exhaust valve 166.
  • the exhaust valve 166 is lifted off its seat in the cylinder head 126 to open the exhaust port 134.
  • the opening of the exhaust port 134 allows the expulsion of the combustion products from the cylinder bore 108.
  • the piston 112 can be of a relatively small skirt length. Since, the piston 112 of the engine 100 has a small skirt length and has less area of contact with the cylinder wall 118, thereby reducing friction between the piston 112 and the cylinder wall 118 and ensuring longevity and durability of the engine 100. Further, the piston 112 with small skirt length has a small weight and therefore less inertia. Hence, the engine 100 expends less thrust of combustion in overcoming the inertia of the piston 112. In addition, with the provision of the exhaust port 134 in the cylinder head 126, the thermal loading and distortions due to the combustion products occurs in the cylinder head 126 and not at the cylinder wall 118, hence, ensuring durability of the engine 100.
  • a delay in the opening and an advance in the closing of the exhaust port 134 is achieved, which provides a short duration of opening of the exhaust port 134.
  • the short duration of the opening of the exhaust port 134 facilitates expulsion of the combustion products from the cylinder bore 108 of the engine 100, such as in the case of two stroke engines where the exhaust stroke is completed in a short duration.
  • the delay in the opening of exhaust port 134 helps the engine 100 in achieving a high expansion ratio, thereby allowing the engine 100 to extract large amount of energy from the combustion of the charge. The engine 100, hence, achieves high fuel efficiency.
  • the high expansion ratio also results in low pressure inside the cylinder bore 108 at the beginning of the exhaust stroke of the piston 112.
  • a pressure gradient between the cylinder bore 108 and an exhaust manifold (not shown) is also low, facilitating less noise during expulsion of the combustion products from the cylinder bore 108 through the exhaust port 134.
  • Fig. 3 illustrates a magnified sectional view of the inlet port 132 of the engine 100, according to an embodiment of the present subject matter.
  • the inlet port 132 is formed in the cylinder head 126 and the opening and closing of the inlet port 132 is regulated by the inlet valve 164.
  • the lifting and seating of the inlet valve 164 at an inlet valve seat 302 achieves the opening and closing, respectively, of the inlet port 132.
  • an inlet valve seat rim 304 is provided at the inlet valve seat 302.
  • the inlet valves seat rim 304 is formed at the inlet valve seat 302 in a similar manner as the exhaust valve seat rim 204 is formed at the exhaust valve seat 202.
  • a length of the inlet valve seat rim 304 measured from the inlet valve seat 302 may be about 25% to about 40% of a valve lift of the inlet valve 164.
  • the length of the inlet valve seat rim 304 measured from the inlet valve seat 302 is about 1.5 millimeter (mm).
  • the operation of the inlet valve 164 is explained, as follows, in relation with Fig. 1 and Fig. 3 .
  • the first follower 178 is operably coupled to the corresponding lobe of the camshaft 168 for actuating the first rocker arm 172 and the first push rod 170.
  • the actuation causes the first push rod 170 to push the inlet valve 164.
  • the inlet valve 164 is pushed, it is lifted off the inlet valve seat 302 in the cylinder head 126 to open the inlet port 132.
  • the inlet port 132 is opened, the pressurized charge from the fuel supply pump 140 enters the auxiliary combustion chamber 136 through the inlet passage 160 and the inlet port 132.
  • the provision of the inlet valve seat rim 304 delays the opening and advances the closing of the of the inlet port 132, which provides a short duration of the opening of the inlet port 132. Such a short duration of opening of the inlet port 132 allows an appropriate amount of charge to be inducted into the auxiliary combustion chamber 136.
  • the inlet valve seat rim 304 facilitates in achieving a high rate of increase of effective flow area between the inlet valve 164 and the inlet valve seat 302, when the inlet valve 164 uncovers the inlet valve seat rim 304.
  • Fig. 4 illustrates a valve timing diagram 400 of the engine 100, according to an embodiment of the present subject matter.
  • the description of the valve timing diagram 400 is provided in relation to Fig. 1 , 2, and 3 for a two stroke internal combustion engine.
  • the positions of the various components of the engine 100 and the events are described in terms of degrees of rotation of the crankshaft 104.
  • the crankshaft 104 rotates.
  • the displacement of the piston 112, at any position between the TCD and the BDC in the cylinder bore 108 is expressed in terms of an angular rotation of the crankshaft 104 in degrees.
  • the TDC and the BDC have been taken as reference positions.
  • the TDC position and the BDC positions are depicted as positions A and D, respectively, on the valve timing diagram 400.
  • the exhaust valve 166 starts to lift from the exhaust valve seat 202 to open the exhaust port 134.
  • the actual lifting of the exhaust valve 166 and opening of the exhaust port 134 is achieved at about 110 to 150 degrees after TDC position, depicted as position B in the valve timing diagram 400.
  • the opening of the exhaust port 134 the combustion products, which were formed during the immediately preceding phase of the working cycle of the engine 100, are discharged from the cylinder bore 108 through the exhaust port 134.
  • the discharge of the combustion products is aided by a pressure difference between a pressure inside the cylinder bore 108 and a pressure in the exhaust manifold.
  • the opening of the exhaust port 134 marks the beginning of the exhaust stroke.
  • the piston 112 crosses and uncovers the transfer ports 116.
  • the transfer ports 116 are uncovered at about 135 to 145 degrees of rotation of the crankshaft 104 from the TDC position.
  • the opening of the transfer ports 116 marks the beginning of the intake stroke. With the opening of the transfer ports 116 the scavenging fluid, such as air or a lean charge of air and fuel, enters into the cylinder bore 108.
  • the scavenging fluid is transferred from the crankcase 102 to the cylinder bore 108 through the induction passages 120 and the transfer ports 116 as a result of the movement of the piston 112 towards the BDC position and the resulting compression of the charge in the crankcase 102.
  • the scavenging fluid entering the cylinder bore 108 is in a swirling motion.
  • the scavenging fluid entering the cylinder bore 108 facilitates the scavenging of the combustion products through the exhaust port 134 in the cylinder head 126, thereby achieving uniflow scavenging and thorough purging of the cylinder bore 108.
  • the exhaust port 134 and the transfer ports 116 are both open.
  • the intake or induction of the scavenging fluid occurs simultaneously along with the expulsion of the combustion products from the cylinder bore 108.
  • the piston 112 crosses the BDC position and moves towards the TDC position due to inertia, and as a result the rotation of the crankshaft 104 continues beyond 180 degrees.
  • the exhaust port 134 is fully open and the transfer ports 116 are open as well. This condition is depicted as position E in the valve timing diagram 400. As the exhaust port 134 is fully open, the removal of the combustion products takes place at a high rate.
  • the piston 112 covers the annular transfer ports 116 at about 210 to 230 degrees of rotation of the crankshaft 104 after the TDC position.
  • the transfer ports 116 are closed at about 30 to 50 degrees of rotation of the crankshaft 104 after the BDC position.
  • the exhaust port 134 closes. In another embodiment, the exhaust port 134 closes at about 260 to 280 degrees of rotation after the TDC position. According to an aspect, the exhaust stroke takes place for about 120 to 160 degrees of the rotation of the crankshaft 104. Shown as position G in the valve timing diagram 400, the closure of the exhaust port 134 marks an end of the exhaust stroke and a commencement of the compression stroke.
  • the movement of the piston 112 towards the TDC position compresses the still swirling scavenging fluid in the main combustion chamber 130 being formed between the piston 112, the cylinder head 126, and the cylinder wall 118.
  • the compression of the swirling scavenging fluid causes turbulence in the main combustion chamber 130. Such turbulence facilitates mixing and combustion of the charge as will be explained later with reference to the combustion phase.
  • an increase in a temperature and pressure of the scavenging fluid occurs. As can be seen from the valve timing diagram 400, the compression stroke takes place for about 60 to 110 degrees of the rotation of the crankshaft 104.
  • the inlet port 132 starts to open.
  • the inlet valve 164 starts to lift from the inlet valve seat 302.
  • the actual opening of the inlet port 132 occurs at about 260 to 290 degrees of rotation of the crankshaft 104 after the TDC position, i.e., at about 80 to 105degrees of rotation after the BDC. This instance is depicted as position H in the valve timing diagram 400.
  • the charge from the fuel supply pump 140 starts to enter the auxiliary combustion chamber 136.
  • the charge entering the auxiliary combustion chamber 136 may include a small quantity of a rich composition of charge.
  • the turbulence created in the main combustion chamber 130 by the compressed scavenging fluid may spread into the auxiliary combustion chamber 136. The turbulence helps in thorough mixing of the charge in the auxiliary combustion chamber 136 and facilitates good combustibility of the fuel in the charge.
  • the scavenging fluid such as air or a lean composition of charge
  • present in the main combustion chamber 130 facilitates in formation of strata of different compositions of air and fuel in the charge in the main combustion chamber 126 and the auxiliary combustion chamber 136, thereby achieving stratification of the charge.
  • the stratification of the charge allows the engine 100 to operate on an overall lean composition of charge. Hence, the overall fuel consumption of the engine 100 is low.
  • the charge on opening of the inlet port 132, the charge enters directly into the main combustion chamber 130 and the combustion of charge is achieved in the main combustion chamber 130.
  • the inlet port 132 is fully open and the induction of the charge takes place at a fast rate. This is shown as position I in the valve timing diagram 400. According to an embodiment, the intake of charge takes place for about 60 to 90 degrees of rotation of the crankshaft 104.
  • the inlet port 132 closes. In an embodiment, the inlet port 132 closes at about 320 to 325 degrees of rotation after the TDC position. Shown as position J in the valve timing diagram 400, the closure of the inlet port 132 concludes the injection of the charge into the auxiliary combustion chamber 136. According to an embodiment, the injection of charge into the auxiliary combustion chamber 136 takes place for about 20 to 45 degrees of rotation of the crankshaft 104.
  • an end of the compression stroke takes place at about 325 to 360 degrees of rotation and is depicted by position K in the valve timing diagram 400.
  • the ignition element 138 such as a spark plug, provided in the auxiliary combustion chamber 136 fires at about 325 to 360 degrees of rotation after the TDC position, i.e. at about 10 to 30 degrees of rotation before the TDC position. In another embodiment, the ignition element 138 fires at 340 to 350 degrees of rotation of the crankshaft after the TDC position.
  • the ignition element 138 ignites the charge in the auxiliary combustion chamber 136. In the auxiliary combustion chamber 136, a partial combustion of the charge occurs and as a result of the compact and small size of the auxiliary combustion chamber 136, a flame front of the burning charge travels quickly into the main combustion chamber 130 where the combustion of the charge is substantially completed.
  • the combustion of the charge produces expanding combustion products that propel the piston 112 towards the BDC position.
  • the expansion stroke begins at about 10 degrees of rotation of the crankshaft after the TDC position.
  • the exhaust port 134 opens and marks the end of the expansion stroke.
  • the expansion stroke ends at about 120 to 140 degrees of rotation of the crankshaft from the TDC position.
  • the expansion stroke takes place for about 110 to 150 degrees of the rotation of the crankshaft 104.
  • the provision of the exhaust valve seat rim 204 at the exhaust valve 202 provides a delay in the opening of the exhaust port 134 on the completion of the expansion stroke.
  • the expansion stroke is of a longer duration as compared to the compression stroke, which leads to a high expansion ratio of the engine 100.
  • the high expansion ratio facilitates the engine 100 in achieving high fuel efficiency and low pressure of the combustion products in the cylinder bore 108 at the beginning of the exhaust stroke.
  • the low pressure of the combustion products helps in substantially noise-less expulsion of the combustion products from the cylinder bore 108 during the exhaust stroke.
  • the provision of the inlet valve seat rim 304 results in high rate of change of the effective flow area when the inlet valve 164 uncovers the inlet valve seat rim 304 and, hence the flow deficiency associated with the conventional inlet valves is substantially prevented. Further, the provision of the inlet valve seat rim 304 provides for a short duration of effective opening of the inlet port 132.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Combustion Methods Of Internal-Combustion Engines (AREA)
  • Output Control And Ontrol Of Special Type Engine (AREA)

Claims (12)

  1. Moteur à combustion interne (100) comprenant :
    un bloc cylindre (106) présentant au moins un alésage de cylindre (108) défini par une paroi de cylindre (118), le bloc cylindre (106) comprenant,
    un piston (112) se déplaçant alternativement dans ledit alésage de cylindre (108) pour entraîner un vilebrequin (104) ;
    une pluralité d'orifices de transfert (116) prévus en anneau dans la paroi de cylindre (118) le long d'une périphérie de la paroi de cylindre (118) ; et
    une tête de cylindre (126) montée sur le bloc cylindre (106), la tête de cylindre (126) comprenant :
    au moins un orifice d'entrée (132) ;
    au moins un orifice d'échappement (134) ;
    une soupape d'entrée (164) prévue au niveau de l'orifice d'entrée (132) pour réguler une ouverture et fermeture de l'orifice d'entrée (132) ;
    une soupape d'échappement (166) prévue au niveau de l'orifice d'échappement (134) pour réguler une ouverture et fermeture de l'orifice d'échappement (134) ;
    un siège de soupape d'entrée (302) prévu au niveau de l'orifice d'entrée (132) pour le repos de la soupape d'entrée (164) ;
    un siège de soupape d'échappement (202) prévu au niveau de l'orifice d'échappement (134) pour le repos de la soupape d'échappement (166) ;
    caractérisé par
    un bord de siège de soupape d'entrée (304) prévu au niveau du siège de soupape d'entrée (302) pour permettre le retard d'ouverture et pour avancer la fermeture de l'orifice d'entrée (132) ; et
    un bord de siège de soupape d'échappement (204) prévu au niveau du siège de soupape d'échappement (202) pour permettre le retard d'ouverture et avancer la fermeture de l'orifice d'échappement (134), et dans lequel
    un levage de soupape maximum de la soupape d'entrée (164) du siège de soupape d'entrée apparaît pendant le mouvement du piston d'une position de point mort bas à une position de point mort haut à environ 120 à 125 degrés de rotation du vilebrequin (104) après la position de point mort bas, dans lequel un levage de soupape maximum de la soupape d'échappement (166) du siège de soupape d'échappement (202) apparaît pendant le mouvement du piston (112) de la position de point mort bas à la position de point mort haut à environ 20 à 30 degrés de rotation du vilebrequin (104) après la position de point mort bas, et
    dans lequel le piston (112) découvre la pluralité d'orifices de transfert (116) pour admettre le fluide de balayage à environ 130 à 150 degrés de rotation du vilebrequin (104) après la position de point mort haut et couvre la pluralité d'orifices de transfert (116) à environ 210 à 230 degrés de rotation du vilebrequin (104) après la position de point mort haut, dans lequel la soupape d'entrée (164) reste fermée alors que la pluralité d'orifices de transfert (116) est ouverte.
  2. Moteur à combustion interne (100) selon la revendication 1, dans lequel le bord de siège de soupape d'entrée (304) et le bord de siège de soupape d'échappement (204) sont prévus à une circonférence du siège de soupape d'entrée (302) et du siège de soupape d'échappement (202) respectivement et dans lequel le bord de siège de soupape d'entrée (304) et le bord de siège de soupape d'échappement (204) sont cylindriques, s'étendant le long d'un axe longitudinal de la soupape d'entrée (164) et la soupape d'échappement (166) respectivement.
  3. Moteur à combustion interne (100) selon la revendication 1, dans lequel le bord de siège de soupape d'entrée (304) et le bord de siège de soupape d'échappement (204) s'étendent à une longueur d'environ 5 % à environ 40 % du levage de soupape d'entrée et du levage de soupape d'échappement respectivement.
  4. Moteur à combustion interne (100) selon la revendication 1, dans lequel le bord de siège de soupape d'entrée (304) et le bord de siège de soupape d'échappement (204) s'étendent à une longueur d'environ 1 millimètre à 1,5 millimètre mesurée depuis le siège de soupape d'entrée (302) et le siège de soupape d'échappement (202) respectivement.
  5. Moteur à combustion interne (100) selon la revendication 1, dans lequel le bord de siège de soupape d'entrée (304) et le bord de siège de soupape d'échappement (204) sont prévus au niveau du siège de soupape d'entrée (302) et du siège de soupape d'échappement (202) respectivement.
  6. Moteur à combustion interne (100) selon la revendication 1, dans lequel la soupape d'entrée (164) ouvre l'orifice d'entrée (132) du moteur à combustion interne (100) pour admettre une charge à environ 240 à 290 degrés de rotation du vilebrequin (104) après la position du point mort haut et ferme l'orifice d'entrée (132) à environ 280 à 330 degrés de rotation du vilebrequin (104) après la position de point mort haut.
  7. Moteur à combustion interne (100) selon la revendication 1, dans lequel la soupape d'échappement (166) ouvre l'orifice d'échappement (134) du moteur à combustion interne (100) pour évacuer des produits de combustion à environ 110 à 140 degrés de rotation du vilebrequin (104) après la position de point mort haut et ferme l'orifice d'échappement (134) à environ 240 à 280 degrés de rotation du vilebrequin (104) après la position de point mort haut.
  8. Moteur à combustion interne (100) selon la revendication 1, comprenant en outre un élément d'allumage (138) pour allumer une charge, dans lequel l'élément d'allumage (138) s'allume à environ 325 degrés de rotation du vilebrequin (104) après la position point mort haut.
  9. Moteur à combustion interne (100) selon la revendication 1, dans lequel une course de détente du moteur à combustion interne (100) commence à environ 10 degrés de rotation du vilebrequin (104) après la position de point mort haut et se termine à environ 120 à 140 degrés de rotation du vilebrequin (104) après la position de point mort haut.
  10. Moteur à combustion interne (100) selon la revendication 1, dans lequel une course de détente se déroule pendant environ 110 à 150 degrés de rotation du vilebrequin (104) et une course de compression se déroule pendant environ 60 à 110 degrés de rotation du vilebrequin (104).
  11. Moteur à combustion interne (100) selon la revendication 1, dans lequel la pluralité d'orifices de transfert (116) est formée sous la forme d'un cône tronqué présentant un angle d'apex alpha et présentant une base au niveau d'une ouverture du passage d'admission dans la pluralité d'orifices de transfert (116), dans lequel l'angle d'apex est l'angle inclus de la surface latérale du cône, et dans lequel un axe de chacun de la pluralité d'orifices de transfert (116) est incliné vers la paroi de cylindre (118) de l'alésage de cylindre (108) à un angle, dans lequel l'angle est mesuré dans un plan horizontal.
  12. Moteur à combustion interne selon la revendication 11, dans lequel l'angle d'apex alpha est d'environ 10 à 30 degrés.
EP11791343.4A 2011-03-07 2011-09-14 Moteur à combustion interne Active EP2683920B1 (fr)

Applications Claiming Priority (2)

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IN671CH2011 2011-03-07
PCT/IN2011/000637 WO2012120529A1 (fr) 2011-03-07 2011-09-14 Moteur à combustion interne

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Citations (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1963780A (en) * 1930-10-29 1934-06-19 Bois Marcel Du Internal combustion engine power unit
GB861080A (en) * 1959-02-25 1961-02-15 Daimler Benz Ag Improvements relating to internal combustion engines with direct fuel injection
DE1813906A1 (de) * 1968-12-11 1970-07-09 Wolf Dipl Ing Karl Zweitakt-Brennkraftmaschine mit Benzineinspritzung
US4907544A (en) * 1989-04-06 1990-03-13 Southwest Research Institute Turbocharged two-stroke internal combustion engine with four-stroke capability

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Publication number Priority date Publication date Assignee Title
GB100873A (en) * 1916-02-21 1916-07-20 Rhys Morgan Improvements in Two-stroke Internal Combustion Engines.
JPS5910343Y2 (ja) * 1980-03-03 1984-04-02 日産自動車株式会社 過給エンジン
JPH01200013A (ja) * 1987-10-12 1989-08-11 Toyota Motor Corp ２サイクル内燃機関
DE3741720A1 (de) * 1987-12-09 1989-06-22 Bayerische Motoren Werke Ag Hubventil mit ventilsitz einer kraftmaschine
JPH0299705A (ja) * 1988-10-06 1990-04-11 Mazda Motor Corp エンジンのバルブシート構造
ITTO20060131A1 (it) * 2006-02-24 2007-08-25 Fiat Auto Spa Dispositivo di controllo della movimentazione di una valvola in particolare di una valvola di aspirazione di un motore a combustione interna
WO2009011145A1 (fr) * 2007-07-16 2009-01-22 Joho Corporation Système pour faire varier l'angle d'ouverture de soupape total par une levée variable

Patent Citations (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1963780A (en) * 1930-10-29 1934-06-19 Bois Marcel Du Internal combustion engine power unit
GB861080A (en) * 1959-02-25 1961-02-15 Daimler Benz Ag Improvements relating to internal combustion engines with direct fuel injection
DE1813906A1 (de) * 1968-12-11 1970-07-09 Wolf Dipl Ing Karl Zweitakt-Brennkraftmaschine mit Benzineinspritzung
US4907544A (en) * 1989-04-06 1990-03-13 Southwest Research Institute Turbocharged two-stroke internal combustion engine with four-stroke capability

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EP2683920A1 (fr) 2014-01-15
WO2012120529A1 (fr) 2012-09-13
WO2012120529A8 (fr) 2012-11-08

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