EP0021170A1 - Two-stroke internal combustion engine - Google Patents

Two-stroke internal combustion engine Download PDF

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Publication number
EP0021170A1
EP0021170A1 EP80103129A EP80103129A EP0021170A1 EP 0021170 A1 EP0021170 A1 EP 0021170A1 EP 80103129 A EP80103129 A EP 80103129A EP 80103129 A EP80103129 A EP 80103129A EP 0021170 A1 EP0021170 A1 EP 0021170A1
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EP
European Patent Office
Prior art keywords
piston
internal combustion
combustion engine
cylinder
stroke internal
Prior art date
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.)
Granted
Application number
EP80103129A
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German (de)
French (fr)
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EP0021170B1 (en
Inventor
Bernhard Dipl.-Ing. Büchner
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BUCHNER BERNHARD DIPL ING
Original Assignee
BUCHNER BERNHARD DIPL ING
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Publication date
Priority to DE2923941 priority Critical
Priority to DE2923941A priority patent/DE2923941C2/de
Application filed by BUCHNER BERNHARD DIPL ING filed Critical BUCHNER BERNHARD DIPL ING
Publication of EP0021170A1 publication Critical patent/EP0021170A1/en
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Publication of EP0021170B1 publication Critical patent/EP0021170B1/en
Expired legal-status Critical Current

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    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02BINTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B75/00Other engines
    • F02B75/16Engines characterised by number of cylinders, e.g. single-cylinder engines
    • F02B75/18Multi-cylinder engines
    • F02B75/22Multi-cylinder engines with cylinders in V, fan, or star arrangement
    • F02B75/228Multi-cylinder engines with cylinders in V, fan, or star arrangement with cylinders arranged in parallel banks
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02BINTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B25/00Engines characterised by using fresh charge for scavenging cylinders
    • F02B25/02Engines characterised by using fresh charge for scavenging cylinders using unidirectional scavenging
    • F02B25/12Engines with U-shaped cylinders, having ports in each arm
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02FCYLINDERS, PISTONS OR CASINGS, FOR COMBUSTION ENGINES; ARRANGEMENTS OF SEALINGS IN COMBUSTION ENGINES
    • F02F7/00Casings, e.g. crankcases or frames
    • F02F7/0002Cylinder arrangements
    • F02F7/0019Cylinders and crankshaft not in one plane (deaxation)
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02BINTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B1/00Engines characterised by fuel-air mixture compression
    • F02B1/02Engines characterised by fuel-air mixture compression with positive ignition
    • F02B1/04Engines characterised by fuel-air mixture compression with positive ignition with fuel-air mixture admission into cylinder
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02BINTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B75/00Other engines
    • F02B75/02Engines characterised by their cycles, e.g. six-stroke
    • F02B2075/022Engines characterised by their cycles, e.g. six-stroke having less than six strokes per cycle
    • F02B2075/025Engines characterised by their cycles, e.g. six-stroke having less than six strokes per cycle two
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02BINTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B3/00Engines characterised by air compression and subsequent fuel addition
    • F02B3/06Engines characterised by air compression and subsequent fuel addition with compression ignition

Abstract

The two-stroke internal combustion engine comprises at least one double piston (22, 24), the cylinders (14, 16) of which are connected on the cylinder head side by a combustion chamber (18) and are charged with an air or fuel-air mixture on the crank chamber side by a charging device. At least one overflow channel (44) leads from the cylinder chamber side of both the leading piston (22) and the trailing piston (24) to the cylinder section of the trailing piston (24) on the combustion chamber side. While the outlet process is controlled by the leading piston (22), the overflow process is controlled by both pistons (22, 24) in such a way that the trailing piston (24) opens the overflow channel at the earliest with the leading piston (22) and later than the leading piston (22) closes. Such a control ensures, even when the internal combustion engine is being charged, for example by means of a charge pump, a compressor, an exhaust gas turbine or the like, that on the one hand no combustion gases can flow back into the crank chamber from the combustion chamber (18) and on the other hand an optimal filling of the combustion chamber (18) is achieved.

Description

  • The invention relates to a two-stroke internal combustion engine with at least one double piston, the cylinders of which are connected to one another on the cylinder head side by a combustion chamber and are charged on the crank chamber side by a charging device with air or a fuel / air mixture, the leading piston of the double piston both overflowing in at least one of the cylinder section on the crank chamber side of the leading piston to the combustion chamber-side cylinder section of the trailing piston leading overflow channel and controls the exhaust process such that the exhaust process begins before the opening of the overflow channel. The two-stroke internal combustion engine can work both according to the gasoline principle and the diesel principle, as well as in multi-fuel mode.
  • A two-stroke internal combustion engine of the above type is known from DE-PS 25 23 712. In this internal combustion engine, the overflow process is controlled exclusively by the leading piston and in such a way that the ent and the overflow process ent speaking angles of rotation have the same bisector and this is shifted relative to the dead center against the direction of rotation. The leading piston closes the overflow channel before the exhaust process is completed. The compression stroke of the double piston thus begins only after the overflow process has been completed, so that no pressure drop directed from the combustion chamber to the crank chamber can occur, which would lead part of the fresh charge of the combustion chamber back into the crank chamber when the overflow channel is open.
  • The overflow process determines the filling of the combustion chamber and thus the performance and efficiency of the internal combustion engine. In the known internal combustion engine, it has been found that the combustion chamber cannot be optimally filled when the crank chamber is charged, because the control time for the overflow is too short, since at the end there is still a pressure drop from the crank chamber into the combustion chamber.
  • From the German patent application 23 47 809 an internal combustion engine with double pistons is also known, the cylinders of which are connected to one another on the cylinder head side by a combustion chamber. The exhaust process is controlled by the leading piston, which also controls the intake process of the fuel-air mixture into the associated crank chamber sealed from the crank chamber of the trailing piston. The trailing piston controls the intake process of purge air into the crank chamber space assigned to it. The lagging piston also controls the overflow processes in two overflow channels, one of which leads from the crank chamber space of the leading piston to the combustion chamber side cylinder section of the trailing piston and the other from the crank chamber space of the trailing piston also to the combustion chamber side cylinder section of this piston. A suitable choice of shape for the two overflow channels in the known internal combustion engine is intended to ensure that the combustion chamber is first purged exclusively with purge air from the crank chamber of the trailing piston, while the supply of the fuel-air mixture from the crank chamber of the leading piston is delayed to such an extent that it reaches the outlet controlled by the leading piston only after completion of the exhaust process.
  • In the internal combustion engine known from DE-OS 23 47 809, both overflow processes are controlled exclusively by the trailing piston. In such a control, however, the intake process begins too late and ends too late, at a point in time at which there is long overpressure in the crank chamber. This means that a substantial part of the charge just sucked in is pushed back to the purge air inlet. Here, not only air is pushed back, but also the fuel-air mixture, since the division of the crank chamber into two separate crank chamber spaces for air or fuel-air mixture is not possible in practice or only with an extremely great effort.
  • In the known internal combustion engine, both overflow processes only begin shortly before bottom dead center and therefore too late for optimal charging. In addition, the two overflow processes end long after the outlet closes at a point in time at which a reverse pressure drop from the cylinder chamber to the crank chamber is already present due to the increasing compression pressure in the cylinder. This pushes back a substantial part of the load just introduced to the crank chamber.
  • The object of the invention is to provide a two-stroke internal combustion engine with a double piston, in which, on the one hand, even when charging by means of a charging device, it is ensured that no combustion gases can flow back from the combustion chamber into the crank chamber and, on the other hand, optimum filling of the combustion chamber is nevertheless achieved.
  • This object is achieved in that the same overflow channel also leads from the crank chamber side cylinder section of the trailing piston to the combustion chamber side cylinder section of this piston and is controlled not only by the leading piston but also by the trailing piston, and that the trailing piston opens the overflow channel at the earliest with the leading piston and closes later than the leading piston.
  • The leading piston still controls the exhaust process and the start of the overflow process. The trailing piston in turn subsequently opens the overflow channel, as a result of which the effective channel cross section is increased. The end of the overflow process is controlled exclusively by the lagging piston.
  • Both pistons control the same overflow channel, whereby its volume can be kept small and its harmful influence on the level of the crank chamber compression remains small. The combined control of the overflow channel by both pistons allows the overflow process to be extended over a crank angle of approximately 2 x 90 ° around bottom dead center. This angle is only determined by the pressure drop between the crank chamber and the combustion chamber-side cylinder section of the trailing piston, depending on the degree of charging of the crank chamber.
  • The charger preferably generates a crank chamber compression ratio of at least 1.5. The charging device can be a high-performance crank chamber pump, which expediently also controls the intake process on the crank chamber side. This can be done, for example, by circumferential surfaces of a connecting rod of the crankshaft-type rotary slide valve or rotary slide valve pair, the circumferential surface of which rotates contact-free or slightly touching and is therefore efficient in terms of efficiency. However, devices for gas-dynamic vibrating charging, for example resonance pipes, resonance volumes or the like, which are matched to a specific speed range of the internal combustion engine, are also suitable.
  • The trailing piston preferably closes the overflow channel at the same time as the outlet process controlled by the leading piston ends. In connection with a high crank chamber compression ratio, the combustion chamber can also be recharged, if the trailing piston closes the overflow channel only after the exhaust process has ended, which is only possible by the present invention. The control of the overflow channel by the trailing piston must be dimensioned such that the boost pressure on the crankcase side is still sufficiently higher than the pressure of the beginning compression in the cylinders. The boost pressure should be up to 0.5 atü higher than the beginning compression pressure in the cylinder.
  • The lagging piston and the leading piston open the overflow channel preferably at different times. In this way, a certain pre-charge or pre-discharge of the overflow channel can be achieved, matched to the desired speed level of the machine.
  • It has proven to be advantageous if the charging process of the cylinder sections on the crank chamber side begins before the trailing piston closes the overflow channel. The fresh gases are supplied with high initial pressure shortly before the overflow channel is closed. This causes an immediate increase in pressure on the crank chamber side, so that the overflow process can be extended further. In addition, sucking back of the fresh charge just introduced into the combustion chamber into the crank chamber is avoided.
  • In a preferred embodiment, an overflow channel is arranged on both sides of a cylinder block wall separating the cylinders of the double piston. These overflow channels are controlled by windows in the shirt of the two pistons below the piston ring zones. The fresh gases cool this partition. If necessary, additional cooling water channels can be provided in the partition.
  • An embodiment offers significant advantages in which at least part of the wall of the combustion chamber opposite the pistons is formed by a counter-piston which is displaceably guided in the cylinder head and can be delivered to the pistons by means of a power device. In this embodiment, the compression ratio of the internal combustion engine can be changed in a structurally simple manner during operation depending on the throttle position of the internal combustion engine. In partial load operation, the compression can be increased so that a lower fuel consumption is achieved in the entire load range without the typical nail behavior known from the diesel engine occurring. A change in the compression ratio within the limits of 1: 2 has proven to be sufficient and practicable. For example, a compression of £ = 8 at full load in part-load operation can be increased to E = 16. The power device can be a hydraulic cylinder which is actuated by a hydraulic control depending on the throttle position. The counter-piston can also be adjusted via a cam, a spring assembly or the like. If a spark plug is required, it can remain in the central position in the counter-piston so that it is moved together with the counter-piston; but it can also be relocated to the periphery of the combustion chamber defined by the cylinder head.
  • In order to ignite the ignitable mixture of the combustion chamber as simultaneously as possible, one is preferably in the Ignition electrode engaging combustion chamber is provided, the counter electrode of which is provided on a surface facing the combustion chamber of a cylinder block wall extending between the two pistons. The ignition electrode can be arranged on the cylinder head in the middle of the combustion chamber, so that an ignition spark extending across the combustion chamber to the counter surface on the cylinder block wall is produced. This cylinder block wall thus forms the ground electrode. The high voltage required to generate such long ignition sparks can be easily achieved with the aid of a high-voltage capacitor ignition. However, the ignition electrode and the counter electrode can also be arranged side by side on the surface of this cylinder block wall facing the combustion chamber. This arrangement has the advantage that it takes up relatively little space in the combustion chamber, which is then available for accommodating other elements, such as injection nozzles or the like.
  • Another aspect of the invention relates to the lubrication of connecting rods and base bearings of the double piston. The connecting rod bearing and the base bearing are preferably encapsulated separately from the crank chamber in a lubricant-tight manner, lubricant lines of a central oil supply opening into the encapsulated bearing interior. In this way, any remaining escape oil quantities from the bearings into the crank chamber can be taken into account in the total oil quantity required for the operation of the internal combustion engine, i.e. deducted from the amount of fresh oil that is introduced into the intake flow for piston lubrication. The oil in the central oil supply can be supplied from an oil tank using an oil pump or a slope (drip lubrication), possibly supported by a slight overpressure in the oil tank.
  • Exemplary embodiments of the invention are to be explained in more detail below with reference to drawings. It shows
    • 1 shows a schematic section through a two-stroke internal combustion engine containing the cylinder axes;
    • FIG. 2 shows a section along the line II-II in FIG. 1;
    • 3 shows a section normal to the cylinder axes through a variant of a combustion chamber of a two-stroke internal combustion engine;
    • 4 shows an axially parallel section through a further variant of a combustion chamber of a two-stroke internal combustion engine and
    • 5 shows a section containing the axes of rotation through a crankshaft with a connecting rod of a two-stroke internal combustion engine.
  • Figures 1 and 2 show a two-stroke internal combustion engine 1 0 , the cylinder block 12 has two cylinders 14 and 16. The cylinders 14 and 16 are closed by a cylinder head 20 which contains a combustion chamber 1.8 connecting the cylinders 14, 16. A piston 22 or 24 is slidably guided in the cylinders 14, 16. The pistons 22, 24 are articulated eccentrically via a common connecting rod 26 by means of a crank pin 27 to a lifting disk 30 mounted on a shaft pin 28. The lifting disk 3o rotates in a crank chamber 34 closed by a trough 32.
  • The cylinders 14 and 16 are divided by piston heads 36 and 38 of the pistons 22, 24 into a cylinder section on the combustion chamber side and a cylinder section on the crank chamber side. The cylinder sections on the crank chamber side, together with the crank chamber 34 and the reciprocating pistons 22, 24, form a crank chamber pump, which is used to draw in the combustion air or the air-fuel mixture through an intake pipe 4o extending into the crank chamber 34.
  • An overflow channel 44 is formed on both sides of an intermediate wall 42 of the cylinder block 12 separating the cylinders 14, 16 so that the sucked-in mixture from the crank chamber 34 can get into the combustion chamber-side cylinder sections and the combustion chamber 18. These overflow channels 44 each have an inlet slot 46 or 48 in the cylinder sections of the cylinders 14, 16 on the crank chamber side and an outlet slot 50 in the cylinder section of the cylinder 16 on the combustion chamber side 24 opened and closed. The outlet slots 5o of the overflow channels 42 are controlled by the upper edge 56 of the trailing piston 24. The upper edge 58 of the piston 22 controls an outlet slot 6o in the cylinder section of the cylinder 14 on the combustion chamber side. The intake process is controlled by the lifting disc 3o, the circumference 62 of which opens or closes the mouth of the intake manifold 4o in a contactless or slightly touching manner.
  • The internal combustion engine works as follows. When the lifting disc 3o rotates in the direction of arrow A, the piston 22 always leads the piston 24. In Fig. 1, the intake process has already ended; the lifting disc 3o closes the Inlet port 40. The mixture introduced into the crank chamber 34 is compressed by the pistons 22, 24 going down. The inlet slots 46, 48 of the overflow channels are closed. Before the windows 52 of the leading piston 22 reach the inlet slots 46 of the overflow channels 44, the upper edge 58 of the leading piston 22 opens the outlet slot 60, so that the burned gases can flow out of the cylinder sections on the combustion chamber side and the combustion chamber 18. The leading piston 22 then opens the inlet slots 46 of the overflow channels 44. Shortly before, simultaneously or somewhat later, the upper edge 56 of the lagging piston 24 has opened the outlet slots 50 of the overflow channels 44. The later opening enables a large outlet with a long overflow control time. The compressed gases of the crank chamber 34 flow via the overflow channels 44 into the cylinder sections of the cylinders 14, 16 on the combustion chamber side and into the combustion chamber 18. They cool both the intermediate wall 42 and the two pistons through which the fresh air or fresh gas charge flows. The intermediate wall can optionally be provided with additional cooling water channels 64. The windows 54 of the lagging piston 24 only reach the inlet slots 48 of the two overflow channels 44 after the leading piston 22 has already opened the inlet slots 46. The piston 22 closes the inlet slots 46 at a time when the outlet slot 60 is still open. The entry slots 48 controlled by the piston 24 remain open beyond this point in time. The inlet slots 48 are closed when the piston 22 closes the outlet slot 60. However, if the boost pressure on the crank chamber side is sufficiently high, they can also remain open beyond the end of the exhaust process, so that the cylinder sections 14, 16 and the combustion chamber 18 on the combustion chamber side can still be reloaded via the overflow channels 44 even when the compression stroke already begins.
  • The overflow process can be extended by a crank angle of about 2x 90 ° around the bottom dead center. It is essentially only limited by the height of the pressure drop from the crank chamber 34 via the overflow channels 44 into the cylinder section of the cylinder 16 on the combustion chamber side. The pressure difference between the boost pressure and the pressure of the commencement of compression in the cylinder section of the cylinder 16 on the combustion chamber side should be approximately between 0.5 and 1.5 atm.
  • The wall 66 of the combustion chamber 18 opposite the piston heads 36, 38 is formed by a counter-piston 68, which is guided and displaceably guided in a sealed manner in a hydraulic cylinder 7o. The hydraulic cylinder 7o is connected via a pressure medium supply line 72 to a control system which moves the counter-piston 68 depending on the throttle position of the internal combustion engine. By moving the counter-piston 68, the compression ratio of the two cylinders 14, 16 can be changed depending on the throttle position. The control increases the compression ratio into the partial load range; for example in a ratio of 1: 2. In this way, fuel consumption can be optimized and significantly reduced across the entire load range.
  • In the internal combustion engine of FIGS. 1 and 2, fuel injection nozzles 74 open into the overflow channels 44 and continuously inject fuel against the direction of flow in the overflow channels 44. This type of injection, which is suitable for both gasoline and diesel engines, allows both qualitative control of the operation, in which only the fuel supply is regulated while the air in the combustion chamber remains completely full, as well as quantitative control which the filling is changed overall.
  • The combustion chamber 18 is essentially cylindrical and has two diametrically opposite regions in cross section, which widen from the end of the combustion chamber remote from the cylinder to the cylinders 14, 16. The bulges 76 thus created are each delimited by a spherical triangular surface. As a result of this design of the combustion chamber 18, the mixture of the cylinders 14, 16 flows tangentially into the combustion chamber 18 and is swirled very effectively by the very high swirl generated at this time, ie at a time when the generation of the swirl no longer costs any power.
  • The bulges 76 that arise in the combustion chamber 18 need not be limited by a spherical triangular surface, they can also be formed by straight ramps that can extend beyond the centers of the cylinder axes 14 and 16 and in FIG. 1 and in FIG. 4 at 77 or 79 are indicated by dashed lines.
  • Fig. 3 shows schematically in a section through a cylinder head a further possibility of fuel injection, in which the very high swirl that arises in the combustion chamber is used. The swirl is indicated in FIG. 3 by arrows 78. The fuel is injected by means of a nozzle 8o against the swirl direction either directly onto the wall 82 of the combustion chamber located between the cylinders 84, 86 or in an area immediately in front of the wall 82. On the upper side of an intermediate wall 88 facing the combustion chamber between the two cylinders 84, 86, an ignition electrode 9o projecting into the combustion chamber is insulated, the counter electrode 92 of which is also integrated in the intermediate wall 88. These ignition electrodes 9o and 92 can, for example in another version also be integrated in the cylinder head gasket. In this way, the combustion chamber above remains free, so that space for accommodating other elements, for example the injection nozzle 8 0 , remains.
  • 4 shows a combustion chamber 94, in the middle of which an ignition electrode 96 is arranged. The counter electrode of the ignition electrode 96 is formed by an intermediate wall 98 between the two cylinders 100 and 102. A high-voltage capacitor ignition generates the high voltage required for the relatively large electrode spacing.
  • The two-stroke internal combustion engine is preferably supplied with oil for the piston lubrication via a central oil supply from a tank, not shown, via an oil pump, also not shown, or a slope. The oil is introduced into the intake flow, for example, whereby the mixing ratio of oil to fuel-air mixture or air can be between 1: 8o and 1: 200 depending on the load condition. FIG. 5 shows details of a crankshaft bearing with two lifting disks 11o, 112 arranged at an axial distance from one another and pressed onto shaft journals 106, 108, between which an eccentrically arranged crankpin 114 extends parallel to the axis. The shaft journals 106, 108 are mounted on roller bearings 116, 118 in the engine block 12o. Between the extender wheels 11 0, 112, a connecting rod 121 is supported by means of bearings rollers 122 on the crankpin 114th The connecting rod 121 has a bearing on the rollers 122 axially projecting Pleuellageraußenring 124, the axial end faces via coaxial sealing rings 126, 128 and O-rings 141, 142 are sealed against the extender wheels 11 0, 112th If necessary, either the sealing rings 126, 128 or the round cord rings 141, 142 can be omitted. The round cord rings are in standardized ring grooves stored. A lubricant channel 13o opens into the sealed interior of the connecting rod bearing ring 124 and is connected to the central oil supply system via a supply bore 132. The mutually axial sides facing away from the extender wheels 11 0, 112, the circulating half each carry a labyrinth seal 134 and 136, respectively, whose fixed halves 143, 144 provided on the cylinder block 12o. Some of the labyrinth seals 134, 136 are additionally sealed by sealing rings 138, 14o and piston rings 145, 146. The labyrinth seals 134 and 136 are self-contained components, the inner ring with the piston ring seal 145, 146 rotating with the shaft journal 106 and 108 (base bearing journal), while the outer rings 143, 144 as well as the bearing outer rings have the same outer diameter with in the housing is clamped. Further seals, not shown, are provided on the sides of the roller bearings 116, 118 axially facing away from the labyrinth seals 134, 136, so that the roller bearings 116, 118 are also fully encapsulated. The feed hole 132 is guided so that it has an opening to the interior of these bearings 116, 118. An advantage of the embodiment of the crankshaft bearing shown in FIG. 5 is that quantities of escape oil that pass the sealing rings 126, 128 or the labyrinth seals 134, 136 with the integrated sealing rings 138, 14o, 145, 146 enter the crank chamber, for piston lubrication total oil required can be taken into account. The amount of fresh oil to be introduced into the intake flow for piston lubrication can be reduced by the amount of escape oil from these bearings, so that overall the oil consumption of the two-stroke internal combustion engine is reduced and can be set permanently for the overall running time of the machine. Unlike with oil sump-lubricated engines, it is independent of the machine's state of wear.

Claims (12)

1. Two-stroke internal combustion engine with at least one double piston, the cylinders of which are connected to one another on the cylinder head side by a combustion chamber and are charged on the crank chamber side by a charging device with air or fuel-air mixture, the leading piston of the double piston both overflowing in at least one of the leading cylinder section of the leading cylinder section Piston to the combustion chamber-side cylinder section of the trailing piston leading overflow channel and controls the exhaust process in such a way that the exhaust process begins before the opening of the overflow channel, characterized in that the same overflow channel (44) also extends from the crankcase-side cylinder section of the trailing piston (24) to the combustion-chamber side cylinder section of this piston (24) leads and is controlled not only by the leading piston (22) but also by the lagging piston (24) and that the lagging piston (24) the overflow channel (44) at the earliest with the leading ends piston (22) opens and closes later than the leading piston (22).
2. Two-stroke internal combustion engine according to claim 1, characterized in that the charging device (22, 24, 30) has a crank chamber compression ratio of at least 1.5.
3. Two-stroke internal combustion engine according to claim 1, characterized in that the trailing piston (24) closes the overflow channel (44) simultaneously with the termination or after the completion of the leading piston (22) controlled exhaust process.
4. Two-stroke internal combustion engine according to claim 1, characterized in that the charging process of the crank chamber side cylinder sections begins before the trailing piston (24) closes the overflow channel.
5. Two-stroke internal combustion engine according to claim 1, characterized in that the crank chamber-side intake process by peripheral surfaces (62) of the connecting rod (26) of the double piston bearing crank cheek rotary valve (30) is controlled.
6. Two-stroke internal combustion engine after. Claim 1, characterized in that an overflow channel (44) is arranged on both sides of a cylinder block wall (42) separating the cylinders (14, 16) of the double piston.
7. Two-stroke internal combustion engine according to claim 1, characterized in that at least part of the piston (22, 24) opposite wall (66) of the combustion chamber (18) by a in the cylinder head (20) slidably guided and by means of a power device (70, 72) counter-piston (68) which can be delivered to the pistons (22, 24) is formed.
8. Two-stroke internal combustion engine according to claim 1, characterized in that in the combustion chamber (18) engages an ignition electrode (90, 96), the counter electrode (92, 98) on a combustion chamber (94) facing surface one between the two pistons (84, 86, 100, 102) extending cylinder block wall (88, 98) is provided.
9. Two-stroke internal combustion engine according to claim 8, characterized in that the ignition electrode (96) is arranged on the cylinder head in the middle of the combustion chamber (94).
10. Two-stroke internal combustion engine according to claim 8, characterized in that the ignition electrode (90) and the counter electrode (92) are arranged side by side on the surface of the cylinder block wall (88) facing the combustion chamber.
11. Two-stroke internal combustion engine according to claim 1, characterized in that the connecting rod bearing (122) and the base bearings (116, 118) are encapsulated separately from the crank chamber in a lubricant-tight manner and that lubricant lines (130, 132) of a central oil supply to the two-stroke internal combustion engine open into the encapsulated bearing interiors.
12. Two-stroke internal combustion engine according to claim 1, characterized in that the fuel is injected continuously against the direction of flow in the overflow channel (44).
EP80103129A 1979-06-13 1980-06-04 Two-stroke internal combustion engine Expired EP0021170B1 (en)

Priority Applications (2)

Application Number Priority Date Filing Date Title
DE2923941 1979-06-13
DE2923941A DE2923941C2 (en) 1979-06-13 1979-06-13

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
AT80103129T AT1253T (en) 1979-06-13 1980-06-04 Two-stroke combustion engine.

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EP0021170A1 true EP0021170A1 (en) 1981-01-07
EP0021170B1 EP0021170B1 (en) 1982-06-23

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EP80103129A Expired EP0021170B1 (en) 1979-06-13 1980-06-04 Two-stroke internal combustion engine

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US (1) US4296714A (en)
EP (1) EP0021170B1 (en)
AT (1) AT1253T (en)
DE (1) DE2923941C2 (en)

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DE19650874A1 (en) * 1996-12-07 1998-06-18 Siegfried Dipl Ing Druminski Twin=cylinder engine
EP0872651A2 (en) * 1997-04-15 1998-10-21 WCI OUTDOOR PRODUCTS, Inc. Flex-rod
WO2012062291A2 (en) * 2010-06-18 2012-05-18 Seneca International Ag Internal combustion engine

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US4741296A (en) * 1981-12-02 1988-05-03 Jackson Francis W Multiple piston expansion chamber engine
US4860701A (en) * 1981-12-02 1989-08-29 Jackson Francis W Multiple piston expansion chamber engine
US5383427A (en) * 1993-07-19 1995-01-24 Wci Outdoor Products, Inc. Two-cycle, air-cooled uniflow gasoline engine for powering a portable tool
US5549701A (en) * 1993-09-20 1996-08-27 Mikhail; W. E. Michael Acetabular cup
US5480448A (en) * 1993-09-20 1996-01-02 Mikhail; W. E. Michael Acetabular cup groove insert
US5722355A (en) * 1994-03-31 1998-03-03 Aktiebolaget Electrolux Twin-piston engine
DE4418844C2 (en) * 1994-05-30 1996-07-18 Helmut Kottmann Two-stroke internal combustion engine with charging cylinder
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AT1253T (en) 1982-07-15
DE2923941A1 (en) 1981-05-07
US4296714A (en) 1981-10-27
DE2923941C2 (en) 1982-12-30
EP0021170B1 (en) 1982-06-23

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