EP2683920A1

EP2683920A1 - Internal combustion engine

Info

Publication number: EP2683920A1
Application number: EP11791343.4A
Authority: EP
Inventors: Melkote Viraraghavachar NARASIMHAN; Varadha Iyengar LAKSHMINARASIMHAN; Varadarajan Ranganathan RAJAGOPALAN; K Kumar; Vythilingam KARUNAHARAN
Original assignee: Indian Institute of Science IISC; TVS Motor Co Ltd
Current assignee: Indian Institute of Science IISC; TVS Motor Co Ltd
Priority date: 2011-03-07
Filing date: 2011-09-14
Publication date: 2014-01-15
Anticipated expiration: 2031-09-14
Also published as: WO2012120529A1; WO2012120529A8; EP2683920B1

Abstract

The subject matter described herein relates to an internal combustion engine (100). The internal combustion engine (100) comprises at least one port (132, 134), and a valve (164, 166) provided at the port (132, 134) to regulate opening and closing of the port (132, 134). A valve seat (202, 302) is also provided at the port (132, 134) for resting the valve (164, 166). Further, a valve seat rim (204, 304) is provided at the valve seat (202, 302) to delay opening and to advance closing of the port (132, 134).

Description

INTERNAL COMBUSTION ENGINE

TECHNICAL FIELD

The subject matter described herein, in general, relates to internal combustion (IC) engines.

BACKGROUND

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.

Conventionally in the two stroke engines, various phases of the working cycle, i.e., 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. To achieve the intake of charge during the intake stroke and to expel the combustion product during the exhaust stroke, 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.

In contrast to the two-stroV.e engines, 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. Usually, 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.

SUMMARY

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

l regulate opening and closing of the port. A valve seat is provided at the port for resting the valve. Further, a valve seat rim is provided at the valve seat for delaying the opening and advancing the closing of the port.

These and other features, aspects, and advantages of the present subject matter will be better understood with reference to the following description and appended claims. This summary is provided to introduce a selection of concepts in a simplified form. This summary is not intended to identify key features or essential features of the claimed subject matter, nor is it intended to be used to limit the scope of the claimed subject matter.

BRIEF DESCRIPTION OF DRAWINGS

The above and other features, aspects, and advantages of the subject matter will be better understood with regard to the following description, appended claims, and accompanying drawings where:

Fig. 1 illustrates a front sectional view of an IC engine, according to an embodiment of the present subject matter.

Fig. 2 illustrates a front sectional view of an exhaust port of the IC engine, according to an embodiment of the present subject matter.

Fig. 3 illustrates a front sectional view of an inlet port of the IC engine, according to an embodiment of the present subject matter.

Fig. 4 illustrates a valve timing diagram of the IC engine, according to an embodiment of the present subject matter.

DETAILED DESCRIPTION

Internal combustion (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. Further, 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. In contrast, 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. Hence, it is difficult to ensure effective gas exchange in the available duration. Further, 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.

The embodiments of 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. In one embodiment, 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.

According to an embodiment of the present subject matter, a valve seat rim is provided at at least one valve seat in the cylinder head. In said embodiment, 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.

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 trie closing of the ports by the valves. In case the 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.

In one implementation, the valve seat rim may be associated with the seat of the exhaust valve. In case of the rim being provided at an exhaust valve seat, 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.

In addition, the provision of the 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. During the initial lift duration of the valve the port is not uncovered due to the valve rim and, hence, the actual flow of air does not occur. Further, 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. In one example, 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. Further, a cylinder block 106 having a cylinder bore 108 is mounted on the crankcase 102.

In one embodiment, the crankcase 102 includes an induction valve 110, for example, a reed valve. According to an implementation, 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. In another embodiment, the cylinder block 106 includes the induction valve 110 to induct the scavenging fluid into the crankcase 102.

During operation of the engine 100, the scavenging fluid entering the crankcase 102 is inducted into the cylinder bore 108. At the time of the induction of the scavenging fluid into the cylinder bore 108 from the crankcace 102, some lubricant particles may be entrained along with the scavenging fluid. In an embodiment, 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 frcri the engine 100.

In one embodiment, 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. In such a case, 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.

It may be understood that in other embodiments, 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.

Further, 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 !ms 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.

According to an embodiment of the present subject matter, 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. In said embodiment, 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.

According to an embodiment, 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 a (not shown in the figure) and having a base at an opening of the induction passage 120 into the respective transfer port 116. The ape.\ angle a 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. In said embodiment, the apex angle is about 10 to 30 degrees to provide a swirling motion to the scavenging fluid entering the cylinder bore 108. Further, in said embodiment, 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. In another embodiment, 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.

In said embodiment, 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. According to an embodiment, 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 Ι· 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.

Further, 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.

Further, 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. In an embodiment, the cylinder head 126 includes an inlet port 132. In other embodiments, the cylinder head 126 may have more i! tii ne inlet port 132. Further, the cylinder head 126 has an exhaust port 134 for expelling combustion products from the main combustion chamber 130. In other embodiments, the cylinder head 126 may have more than one exhaust port 134. In an embodiment, the inlet port 132 is smaller in diameter than the exhaust port 134. In said embodiment, 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.

According to an embodiment, 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. In other embodiments, 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.

In one implementation, 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. In an embodiment, 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.

During the movement of the piston 112 from the TDC to the BDC, at about the end of the expansion stroke, 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.

In addition, the auxiliary combustion chamber 136 is in fluid communication with a fuel supply pump 140. In one embodiment, the fuel supply pump 140 is a reciprocating type pump and includes an auxiliary piston 142 reciprocating inside an auxiliary bore 144. In an embodiment, 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. Such a design of the auxiliary piston 142 ensures durability of the fuel supply pump 140. The auxiliary bore 144 is formed in an auxiliary cylinder block 146 of the fuel supply pump 140. In an embodiment, 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. In another embodiment, the fuel supply pump 140 is integrated with the crankcase 102 of the engine 100. In yet another embodiment, 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.

In an implementation, the auxiliary crankshaft 148 of the fuel supply pump 140 obtains a drive from the crankshaft 104 of the engine 100. In one embodiment, the auxiliary crankshaft 148 obtains the drive from the crankshaft 104 through a chain drive assembly (not shown in the figure). Similarly, other drive mechanisms may also be used to provide a drive from the crankshaft 104 to the auxiliary crankshaft 148.

In case of a chain drive assembly providing the drive 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. In said embodiment, 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.

Further, 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. In an embodiment, 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. In one example, 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. In one implementation, a small quantity of the charge is received by the fuel supply pump 140 from the fuelling device. In said implementation, the charge is richer in composition as conpared to the corresponding stoichiometric composition of the charge for an operation, for example, full load operation or part load operation, of the engine 100.

In another embodiment, 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. In said embodiment, 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. Further, in one embodiment, 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. In an embodiment, a volumetric capacity of the fuel supply pump 140 is about 10% of a volumetric capacity of the engine 100. In said embodiment, the capacity of the fuel supply pump 140 is in a range of about 10 cubic centimeters (cc) to 15 cc.

Further, 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. In one embodiment, 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.

In one embodiment, the auxiliary exhaust port 156 includes a discharge valve (not shown in the figure) to regulate the ind".ction of pressurized charge from the fuel supply pump 140 into the inlet passage 160. in .n implementation, 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.

In operation, 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. During this phase, the piston 112 is said to be in a compression stroke and is approaching the TDC position in the cylinder bore 108. Accordingly, during this phase, 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.

Due to this stratification of charge, 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. It will be understood that the engine 100 may include only the main combustion chamber 130 for the combustion of the charge. Jn such a case, the fuel supply pump 140 inducts charge into the main combustion chamber 130.

In addition, the cylinder head assembly 124 includes a valve train assembly 162. In one embodiment, the valve train assembly 162 is provided in the cylinder head 126 and housed inside the cylinder head assembly 124. In one implementation, the valve train assembly 162 includes an inlet valve 164 and an exhaust valve 166. In said embodiment, the inlet valve 164 is provided at the inlet port 132 in the cylinder head 126, and the exhaust valve 166 is provided at the exhaust port 134.

An inlet valve seat and an exhaust valve seat (both not shown in the figure) 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. By lifting and seating at their respective valve seats, 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 pros ided.

According to an embodiment of the present subject matter, 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. In said embodiment, 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. In another embodiment, the valve seat rim is provided at the inlet valve seat as well as the exhaust valve seat.

Further, in one embodiment, the valve seat rim is formed integrally with the cylinder head 126. In said embodiment, the valve seat rim is formed as a bore in the cylinder head 126, extending to the valve seat. In another embodiment, the valve seat rim may be formed separately from the cylinder head 126 and disposed at the valve seat.

The provision of the 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. Further, the 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. Similarly, 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). Further, the first rocker arm 172 and the second rocker arm 176 include a first follower 178 and a second follower (not shown in figure), respectively. In one embodiment, the first follower 178 and the second follower are roller type followers.

In said embodiment, the inlet valve 164 and the exhaust valve 166 are spring loaded valves. In said embodiment, 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. To drive the inlet valve 164 and the exhaust valve 166, the camshaft 168 may obtain a drive from the engine 100, for example, through a chain drive assembly. In such a case, 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.

In one embodiment, 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 166and also provides a drive to the auxiliary crankshaft 148 for driving the fuel supply pump 140, withcu; 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.

In another embodiment, 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. As shown in Fig. 2, 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. According to an embodiment, 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. In said embodiment, the exhaust valve seat rim 204 is cylindrica¹ in shape. As mentioned earlier, 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. In one embodiment, 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. In an example, 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. On such an actuation, the exhaust valve 166 is lifted off its seat in the cylinder head 126 to open the exhaust port 134. The opening of t^e exhaust port 134 allows the expulsion of the combustion products from the cylinder bore 108.

With the provision of the exhaust port 134 and the exhaust valve 166 in the cylinder head 126, 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. Moreover, due to the presence of the exhaust valve seat rim 204, 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. Further, 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 o^f the piston 112. As a result, 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. As mentioned earlier, 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.

According to an embodiment, an inlet valve seat rim 304 is provided at the inlet valve seat 302. In said embodiment, 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. In an embodiment, a length of the iniet 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. In an example, 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. As 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. As 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.

Further, 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 ma sr. 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 04. As the piston 112 reciprocates between the TDC position and the BDC position in the cylinder bore 108, the crankshaft 104 rotates. Hence, 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. For the purpose of description of the valve timing diagram of the engine 100, the TDC and the B DC 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.

At about 90 to 1 10 degrees after the TDC position of the piston 112, the exhaust valve 166 starts to lift from the exhaust valve seat 202 to open the exhaust port 134. However, due to the exhaust valve seat rim 204, the actual lifting of the exhaust valve 166 and opening of the exhaust port 134 is achieved at about 1 10 to 150 degrees after TDC position, depicted as position B in the valve timing diagram 400. With 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. At about 10 to 15 degrees after the opening of the exhaust port 134, i.e., at about 130 to 150 degrees rotation of the crankshaft 104 after the TDC position, depicted as position C in the valve timing diagram, the piston 112 crosses and uncovers the transfer ports 116. In another embodiment, 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.

Further, due to the inclination of the transfer ports 116 to the cylinder wall 118 and as a result of the geometry of the transfer ports 116, 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.

At the BDC position (position D in Fig. 4), i.e., at a 180 degrees rotation of the crankshaft 104 after the TDC position, the exhaust port 134 and the transfer ports 116 are both open. As a result, 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.

At about 200 to 210 degrees after the TDC position, i.e., about 20 to 30 degrees after the BDC position, 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.

Further, at position F in FIG. 4, during the motion of the piston 112 from the BDC position to the TDC position, 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. Hence, the transfer ports 116 are closed at about 30 to 50 degrees of rotation of the crankshaft 104 after the BDC position. With the closure of the transfer ports 116, the intake stroke of the engine 100 is concluded. However, due to the open exhaust port 134, the exhaust stroke continues.

Further, at about 240 to 280 degrees of rotation of the crankshaft 104 after the TDC position, i.e., at about 60 to 100 degrees of rotation 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.

During 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. During the compression stroke, 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 1 10 degrees of the rotation of the crankshaft 104.

At about 240 to 265 degrees of rotation of the crankshaft 104 after the TDC position, i.e., at about 60 to 85 degrees of rotation after the BDC position, the inlet port 132 starts to open. At this instance, the inlet valve 164 starts to lift from the inlet valve seat 302. The provision of the inlet valve seat rim 304 at the inlet valve seat 302, however, delays the opening of the inlet port 132. 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.

With the opening of the inlet valve 132, 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. Further, 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.

Further, as a result of a smail quantity of rich composition of the charge in the auxiliary combustion chamber 136 ana the scavenging fluid, such as air or a lean composition of charge, present in the main combustiofi 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.

In another embodiment, 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.

At about 300 to 305 degrees of rotation of the crankshaft 104 after the TDC position, i.e., at about 120 to 125 degrees of crankshaft rotation after the BDC position, 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.

Further, at about 280 to 330 d gree? of rotation after the TDC position, i.e., at about

140 to 145 degrees of rotation after the BDC position, 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.

Further, 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.

Subsequently, 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 the auxiliary combustion chamber 136 and the main combustion chamber 130, shown as position L ir the valve timing diagram 400, marks the beginning of the expansion stroke. Th^ combustion of the charge produces expanding combustion products that propel the piston 112 towards the BDC position.

In another embodiment, the expansion stroke begins at about 10 degrees of rotation of the crankshaft after the TDC position.

As mentioned earlier, at position B shown in the valve timing diagram 400, the exhaust port 134 opens and marks the end of the expansion stroke. In one embodiment, the expansion stroke ends at about 120 to 140 degrees of rotation of the crankshaft from the TDC position. The expansion stroke, as is clear from the valve timing diagram 400, takes place for about 1 10 to 150 degrees of the rotation of the crankshaft 104.

The previously described versions of the subject matter and its equivalent thereof have many advantages, including those which are described herein.

The provision of the exhaust valve seat rim 204 at the exhaust valve 202 provides a delay in the opening of the exhaurt port 134 on the completion of the expansion stroke. As a result, the expansion stroke is of a longer duration as compared to the compression stroke, which leads to a high expansion ration 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.

Further, 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. It will be understood that although the subject matter is described in detail with reference to a two stroke internal combustion engine, however, the description can be extended to four stroke internal combustion engines. Further, the application of the present subject matter is not limited to internal combustion engines and may be extended to compressors, pumps, etc.

Although the present subject matter has been described in considerable detail with reference to certain embodiments thereof, other embodiments are possible. As such, the spirit and scope of the appended claims should not be limited to the description of the embodiments contained therein.

Claims

An internal combustion engine (100) comprising:

at least one port (132, 134);

a valve (164, 166) provided at the port (132, 134) to regulate an opening and closing of the port (132, 134);

a valve seat (202, 302) provided at the port (132, 134) for resting the valve (164, 166); and

a valve seat rim (204, 304) provided at the valve seat (202, 302) to delay opening and to advance closing of the port (132, 134).

The internal combustion engine (100) as claimed in claim 1 , wherein the valve seat rim (204, 304) is provided at a "ircumference of the valve seat (202, 302) and wherein the valve seat rim (204, 304) is cylindrical, extending along a longitudinal axis of the valve (164, 166).

The internal combustion engine (100) as claimed in claim 1 , wherein the valve seat rim (204, 304) extends to a length of about 5% to about 40% of a valve lift.

The internal combustion engine (100) as claimed in claim 1 , wherein the valve seat rim (204, 304) extends to a length of about 1 millimeter to 1.5 millimeter measured from the valve seat (202, 302).

The internal combustion engine (100) as claimed in claim 1 , wherein the valve seat rim (204, 304) is provided at least at one of an inlet valve seat (202) and an exhaust valve seat (204).

The internal combustion engine (100) as claimed in claim 1 , wherein an inlet valve (164) opens an inlet port (132) of the internal combustion engine (100) to induct a charge at about 240 to 290 degr- s of rotation of the crankshaft (104) from a top dead centre (TDC) position and closes the inlet port (132) at about 280 to 330 degrees of rotation of the crankshaft (104) from the TDC position.

The internal combustion engine (100) as claimed in claim 1 , wherein an exhaust valve (166) opens an exhaust port (134) of the internal combustion engine (100) to discharge combustion products at about 1 10 to 140 degrees of rotation of a crankshaft (104) from a top dead centre (TDC) position and closes the exhaust port (134) at about 240 to 280 degrees of rotation of the crankshaft (104) from the TDC position.

8. The internal combustion engine (100) as claimed in claim 1 , further comprising an ignition element (138) to ignite a charge fires at about 325 to 360 degrees of rotation of the crankshaft from the TDC position.

9. The internal combustion engine (100) as claimed in claim 1, wherein an expansion stroke of the internal combustion engine (100) begins at about 10 degrees of rotation of the crankshaft (104) from a top dead centre (TDC) position and concludes at about 120 to 140 degrees of rotation of the crankshaft (104) from the TDC position.

10. The internal combustion engine (100) as claimed in claim 1 , wherein an expansion stroke takes place for about 1 10 to 150 degrees of rotation of a crankshaft (104) and a compression stroke takes place for about 60 to 1 10 degrees of rotation of the crankshaft (104).

1 1. The internal combustion engine (100) as claimed in claim 1 , wherein a piston (112) uncovers a plurality of transfer ports (116) in a cylinder wall (118) to induct a scavenging fluid at about 130 to 150 degrees of rotation of a crankshaft (104) from a top dead centre (TDC) positicn and covers the plurality of transfer ports (116) at about 210 to 230 degrees of rotation of the crankshaft (104) from the TDC position.