JP5195467B2 - Variable compression ratio device for internal combustion engine - Google Patents

Description

  The present invention relates to a variable compression ratio device for an internal combustion engine provided with a variable compression ratio mechanism that changes an engine compression ratio according to the position of a movable piston.

  As described in Patent Document 1, a variable compression ratio mechanism that changes the engine compression ratio of an internal combustion engine is generally driven and held by an actuator that uses electric power or hydraulic pressure as a drive source, and improves thermal efficiency at low loads. Therefore, the operation of the actuator is controlled so that the compression ratio becomes high and the compression ratio becomes low to avoid knocking at high loads.

JP 2004-150353 A

  Transition from a low load to a high load, such as during acceleration that increases the engine load (engine torque) in response to the driver's depressing operation of the accelerator pedal, from a low load operating condition where a high compression ratio setting is used Accordingly, when the engine compression ratio is changed from the high compression ratio side to the low compression ratio side, if the reduction of the engine compression ratio is delayed with respect to the increase in engine load, the frequency of occurrence of knocking increases. In particular, when the actuator that drives and holds the variable compression ratio mechanism is an electric actuator such as a servo motor, the compression ratio variable command value is sent from the ECU (engine control unit) to the actuator in response to an acceleration start request by depressing the accelerator pedal. Inevitably there is a time delay of, for example, several tens of milliseconds, and there is a rotation increase delay due to the inertia of the motor or the speed reducer immediately after the actuator starts driving. It is difficult to increase the response speed.

  In order to suppress the occurrence of knocking, for example, if the setting of the engine compression ratio at the time of low speed and low load is previously reduced, the fuel efficiency improvement effect is reduced by this reduction in the compression ratio. Further, if the ignition timing is retarded in order to suppress the occurrence of knocking, sufficient torque cannot be generated during acceleration, resulting in a reduction in acceleration performance.

  In the above-mentioned patent document 1, in order to enhance the response to the low compression ratio side, in addition to the electric actuator (servo motor) that drives the control shaft of the variable compression ratio mechanism, the control shaft is urged to the low compression ratio side. A spring is provided. However, if the responsiveness to the low compression ratio side is increased by using the spring in this way, the control shaft is resisted against the biasing load of the spring when the control shaft is held at the high compression ratio position. It must be held at the position, so that the holding ratio of the compression ratio is lowered, or the actuator is enlarged and the energy consumption is increased.

  Furthermore, since an input load caused by combustion pressure, inertial force, etc. acts on the actuator from the variable compression ratio mechanism side, it is necessary to drive and hold the variable compression ratio mechanism against this input load. It is difficult to achieve both responsiveness and retention of the compression ratio without incurring a reduction in the size.

  The variable compression ratio mechanism is mechanically coupled to a movable piston disposed in a cylinder of the housing so as to be reciprocally movable, and changes the engine compression ratio as the movable piston moves along the cylinder axial direction. An urging means such as an urging spring for urging the movable piston in a direction opposite to an input load acting on the movable piston from the variable compression ratio mechanism side is provided.

  The first aspect of the invention is a low compression ratio that acts on the movable piston from the variable compression ratio mechanism side when the movable piston is located on the highest compression ratio side and is in a high compression ratio state that is abutted against the stopper of the housing. The peak value of the input load to the side is smaller at the low load and larger at the high load than the urging load to the high compression ratio acting on the movable piston by the urging means.

  In this case, for example, during low load operation using a high compression ratio setting, the movable piston is held in the high compression ratio state pressed against the stopper of the housing by the biasing load toward the high compression ratio that exceeds the peak value of the input load. Thus, rattling of the movable piston is suppressed, and the retention of the compression ratio is improved. For this reason, generation | occurrence | production of the noise by repeating the collision with the stopper by the shakiness of a movable piston is suppressed, and energy loss (decrease in fuel consumption) by the fluctuation | variation of a compression ratio and the fall of drivability are suppressed. it can.

  On the other hand, at the time of acceleration when changing the engine load to the high load side by depressing the accelerator pedal, etc., from the operating state at such a high compression ratio and low load, as the engine load increases, the input to the low compression ratio side The peak value of the load becomes larger than the urging load, and the movable piston is pushed to the low compression ratio side, and the responsiveness of the movable piston to the low compression ratio side is improved. In this way, the response and retention of the compression ratio by the movable piston can be achieved at a high level.

According to a second aspect of the present invention, the link row connecting the control shaft of the variable compression ratio mechanism and the movable piston includes a lever having the control shaft as a fulcrum and a lever link having one end connected to the tip of the lever. The angle formed between the lever and the lever link changes according to the position of the movable piston.

  Depending on the angle between the lever and the lever link, the magnitude of the control shaft torque acting on the control shaft due to the combustion pressure or inertia force of the internal combustion engine, etc. The magnitude of the applied input load changes. That is, the magnitude of the input load acting on the movable piston can be appropriately changed according to the position of the movable piston by setting the link row. For this purpose, the direction and magnitude of the force that finally acts on the movable piston due to the input load and the urging load are made appropriate according to the position of the movable piston, that is, the engine compression ratio. It is possible to achieve both responsiveness and retention of the engine compression ratio.

  Thus, according to the present invention, it is possible to achieve both responsiveness and retention of the engine compression ratio by the movable piston.

The block diagram which shows one Example of the variable compression ratio apparatus of the internal combustion engine which concerns on this invention. The block diagram in the high compression ratio state (A) and low compression ratio state (B) of the Example using a two-position switching valve. The block diagram in the high compression ratio state (A) of an Example using a 3 position switching valve, the low compression ratio state (B), and a medium compression ratio state. (A) is a timing chart (A) which shows an example of a change of the compression ratio, load, etc. at the time of acceleration, (B) is a timing chart which shows the compression ratio and piston displacement in area | region R1 of (A). The characteristic view which shows three examples of the load change with respect to piston displacement in the time of low load (A) and high load (B). The timing chart which shows changes, such as a compression ratio and a load at the time of acceleration. The characteristic view which shows an example of the load change with respect to piston displacement. The block diagram which shows an example of the link structure in a high compression ratio (A) and a low compression ratio (B). The characteristic view which shows the other example of the load change with respect to piston displacement. The timing chart which shows the other example of changes, such as a compression ratio at the time of acceleration, a load. The characteristic view which shows the other example of the load change with respect to piston displacement. The timing chart which shows the other example of changes, such as a compression ratio at the time of acceleration, a load. The block diagram which shows the other example of the link structure in a high compression ratio (A) and a low compression ratio (B). The characteristic view which shows the other example of the load change with respect to piston displacement. The block diagram which shows the other example of the link structure in a high compression ratio (A) and a low compression ratio (B). The characteristic view which shows the other example of the load change with respect to piston displacement. The block diagram which shows one Example of the variable compression ratio apparatus which used the actuator together.

  Hereinafter, preferred embodiments of the present invention will be described with reference to the drawings. 1 and 2, the variable compression ratio mechanism 10 uses a multi-link piston-crank mechanism, and moves a part of the link structure to change the piston top dead center position. The mechanical compression ratio, that is, the nominal compression ratio is changed, and its basic structure is known from Japanese Patent Application Laid-Open No. 2003-322036 and the like, so only a brief description will be given here. In the figure, only the link center line of each link is drawn.

  The variable compression ratio mechanism 10 connects a lower link 13 rotatably attached to a crank pin 12 of a crankshaft 11 of an internal combustion engine, and one end of the lower link 13 and a piston (combustion chamber piston) 14 of the internal combustion engine. An upper link 15, a control shaft 16 rotatably supported on the engine body side such as a cylinder block, an eccentric shaft portion 17 provided eccentrically with respect to the rotation center of the control shaft 16, and one end of the lower link 13 And a control link 18 rotatably connected to the eccentric shaft portion 17 at the other end. When the control shaft 16 rotates, the motion restraint condition of the lower link 13 changes via the control link 18, and the engine compression ratio (nominal compression ratio) is accompanied by changes in the top dead center position and the bottom dead center position of the piston 14. Changes.

  The oil draining mechanism 20 includes a housing 22 in which a cylinder 21 is formed, and a movable piston 23 attached to the cylinder 21 so as to be reciprocally movable. The first oil chamber 24 is placed in the cylinder 21 by the movable piston 23. And the second oil chamber 25 are liquid-tightly defined. That is, the cylinder 21 is partitioned into the first oil chamber 24 and the second oil chamber 25 by the movable piston 23.

  The movable piston 23 and the control shaft 16 are mechanically connected by a link row including a lever 41 and a lever link 42. The lever 41 is fixed to the control shaft 16 and swings about the rotation center of the control shaft 16 as a fulcrum. The lever link 42 has one end rotatably connected to the tip of the lever 41 and the other end rotatably connected to the tip of the piston rod 43 of the movable piston 23 protruding from the housing 22. When the movable piston 23 moves to the right in the drawing along the axial direction of the cylinder 21, that is, to the high compression ratio side, the control shaft 16 rotates counterclockwise, the engine compression ratio increases, and the movable piston 23 moves to the cylinder 21. When moving to the left in the figure, that is, toward the low compression ratio side, the control shaft 16 rotates clockwise, and the engine compression ratio decreases.

  In the first oil chamber 24 of the cylinder 21, the movable piston 23 is movable in the direction opposite to the direction of the input load (peak value) Fin acting on the movable piston 23 from the high compression ratio side, that is, the variable compression ratio mechanism 10 side. An urging spring 44 that is a coil spring as an urging means for urging the piston 23 is provided. The movable piston 23 is abutted against a stopper 45 provided in the housing 22 in the state of the high compression ratio moved to the highest compression ratio side, and the position thereof is mechanically restricted and restricted. In the illustrated example, the stopper 45 is configured by a separate member from the housing 22 made of an appropriate material such as rubber or metal in consideration of impact relaxation and durability at the time of collision with the movable piston 23. The housing 22 may be provided integrally, that is, the wall surface of the housing 22 may be used as a stopper.

  In addition, the oil draining mechanism 20 includes a first inlet pipe 26 that supplies hydraulic oil to the first oil chamber 24, a first drain pipe 27 that discharges hydraulic oil from the first oil chamber 24, and a second oil chamber 25. It has the 2nd inlet pipe 28 which supplies hydraulic oil, and the 2nd extraction pipe 29 which discharges hydraulic oil from the 2nd oil chamber 25, and check valves 30, 31 are provided in each pipeline 26-29, respectively. , 32, 33 are provided. In addition, a switching valve 34 that opens and closes the first exhaust pipe 27 and the second exhaust pipe 29 is provided between the check valves 31 and 33 and the cylinder 21. This switching valve 34 is a two-position switching valve type solenoid valve that switches the position of the spool 36 to two positions against the spring force of the return spring 37 according to the excitation state of the solenoid 35. The excitation state of the solenoid 35 is controlled. It is controlled according to the engine operating state by a control signal from the unit 38.

  As shown in FIG. 2A, when the solenoid 35 is in a non-excited state, the spool 36 is moved to the left side in the drawing by the spring force of the return spring 37, and the first extraction pipe 27 is closed. , The second vent pipe 29 is opened. As a result, only the hydraulic oil in the second oil chamber 25 is discharged through the second exhaust pipe 29 without discharging the hydraulic oil in the first oil chamber 24, so that the movable piston 23 moves to the low compression ratio side. The movement is suppressed and prevented, and only the movement toward the high compression ratio side is allowed, and it is finally abutted against the stopper 45 by the load acting on the movable piston 23 as will be described later. It becomes a high compression ratio state.

  As will be described later, when the total load acting on the movable piston 23 is an alternating load that reverses to the left and right (see load F in FIG. 4), the second change is made through a gap between the cylinder 21 and the movable piston 23. The movable piston 23 moves slightly to the left in the drawing by the amount of hydraulic oil that leaks from one oil chamber 24, and repeats the movement back to the position where it strikes the right stopper 45 again.

  On the other hand, as shown in FIG. 2 (B), when the solenoid 35 is energized, the spool 36 is moved to the second position in the right side of the drawing, the first exhaust pipe 27 is opened, and the second exhaust pipe 29 is closed. Is done. As a result, only the hydraulic oil in the first oil chamber 24 is discharged through the first drain pipe 27 without discharging the hydraulic oil in the second oil chamber 25, so that the movable piston 23 moves to the high compression ratio side. Movement is suppressed / prevented, and substantially only movement toward the low compression ratio side is possible. Here, at the time of high load in the high compression ratio state abutted against the stopper 45, as will be described later, the input load (peak value) Fin to the low compression ratio side exceeds the bias load Fspr by the bias spring 44. Therefore, the movable piston 23 can be quickly moved to the low compression ratio side, that is, the left side of the figure.

  If the solenoid 35 is held in an excited state, the hydraulic oil in the second oil chamber 25 cannot be removed, so that the movable piston 23 repeats minute vibrations in the vicinity of the predetermined low compression ratio position shown in FIG. It will be. When the solenoid 35 is switched from excitation to non-excitation, the movement is reversed.

  FIG. 3 shows an example in which a three-position switching type solenoid valve 34A is used as the switching valve instead of the two-position switching valve type shown in FIG. In the three-position switching valve 34A, a third position (C) is added as a switching position of the spool 36 in addition to the first position (A) and the second position (B). The behavior of the first position (A) and the second position (B) is the same as that when the two-position switching valve 34 described above is used.

  In the third position (C), both the first extubation path 27 and the second extubation path 29 are closed. As a result, the hydraulic oil is not discharged from either the first oil chamber 24 or the second oil chamber 25, the movable piston 23 is held at a predetermined intermediate position, and the engine compression ratio is maintained at an appropriate intermediate compression ratio. can do.

  Thus, the oil draining mechanism 20 does not move the piston by hydraulic pressure or electric power unlike a general actuator, but by the input load Fin acting on the movable piston 23 or the biasing load Fspr by the biasing spring 44. 23, and the opening and closing of the outlet pipes 27 and 29 are switched by the switching valve 34, thereby suppressing / preventing movement of the movable piston 23 to one or both of them and holding the movable piston 23 in a predetermined position. Therefore, the hydraulic oil supplied to the oil chambers 24 and 25 via the inlet pipes 26 and 28 does not need to be pressurized and may be about atmospheric pressure.

  With reference to FIG. 1, the load which acts on the movable piston 23 is demonstrated. Three forces, that is, loads Fin, Fspr, and Foil mainly act on the movable piston 23. Due to the combustion load generated by the internal combustion engine, a torque Tcsft in the low compression ratio direction (clockwise direction) is generated on the control shaft 16 of the variable compression ratio mechanism 10, and the torque Tcsft is transmitted to the lever 41 and the link lever 42. It acts on the movable piston 23 as an input load Fin to the low compression ratio side (left side in the figure) along the axial direction of the cylinder 21.

  This input load Fin is expressed by multiplying the magnitude of the control shaft torque Tcsft by the amplification factor (α (x)) determined by the length of the lever 41 and the lever link 42 and the angle α formed by both 41 and 42. Is done. Here, the angle α formed by the lever 41 and the lever link 42 changes depending on the position of the movable piston 23 (displacement from the stopper position abutted against the stopper 45 to the low compression ratio side) x, so that the amplification factor (α (x )) Is a function of the displacement x of the movable piston 23. That is, according to this link configuration, the magnitude of the input load Fin with respect to the control shaft torque Tcsft changes according to the displacement x of the movable piston 23, that is, the engine compression ratio.

  Since the magnitude of the combustion load changes from moment to moment, the input load Fin also changes from moment to moment. Here, the explanation will be made mainly using the maximum value, that is, the peak value of the input load.

  Next, an urging load Fspr on the high compression ratio side, that is, the direction opposite to the input load Fin acts on the movable piston 23 by the urging spring 44. This urging load Fspr is expressed by “initial load Fset + spring multiplier k × piston displacement x”, and the movable piston 23 increases monotonously in proportion to the movement toward the low compression ratio side.

  Further, a reaction force Foil due to hydraulic viscosity acts, and this reaction force Foil can be expressed by “piston moving speed dx / dt × viscosity coefficient C”. Since it is proportional to the piston speed, the reaction force Foil is zero when the movable piston 23 is not moving.

  Therefore, the total load acting on the movable piston 23, that is, the piston load F that attempts to move the movable piston 23 can be expressed by “F = Fin−Fspr−Foil”. In particular, when the movable piston 23 is stationary, it is expressed by “F = Fin−Fspr” from Foil = 0. Here, the frictional force due to the sliding resistance and the hydraulic pressure in the oil chambers 24 and 25 are ignored because they have a small effect. However, even if these are taken into consideration, the actual behavior of the movable piston 23 is substantially the same. .

  4A is similar to FIGS. 6, 10, and 12, the compression ratio ε (control shaft angle is converted into the compression ratio at the time of acceleration from a low load operation state using a high compression ratio setting. 1), control shaft torque Tcsft, input load Fin, urging load Fspr, total piston load F, and a change of ON (excitation) and OFF (no excitation) of solenoid 35 of switching valve 34, A section from t0 to t0 represents an operation state with a high compression ratio and a low load, and t1 is an acceleration start time. FIG. 4B is a timing chart showing changes in the compression ratio ε and the piston displacement x in the region R1 in FIG.

  In this case, the magnitude relationship between the urging load Fspr and the input load Fin is frequent in a high compression ratio state in which the solenoid 35 is turned off, that is, in a non-excited state and the movable piston 23 is pressed against the stopper 45 on the high compression ratio side. The alternating piston load F is reversed and the total piston load F is reversed over zero. For this reason, the holding performance of the movable piston 23 at the stopper holding position on the high compression ratio side is lowered, and the movable piston 23 is likely to be rattled. Specifically, it is represented by the symbol R2 in FIG. As shown, the behavior that the movable piston 23 moves away from the stopper 45 when Fin> Fspr and sits on the stopper 45 when Fin <Fspr is repeated. For this reason, noise and vibration are generated when seated on the stopper 45, which becomes abnormal noise, and thermal efficiency due to energy loss due to hydraulic viscosity and reciprocation behavior of the combustion chamber piston 14 due to minute displacement of the movable piston 23. It will cause problems such as deterioration.

  Next, referring to the illustrated embodiment, the characteristic configuration and operational effects of the present invention will be listed. However, the present invention is not limited to the configuration of the illustrated embodiment, and includes various modifications and changes without departing from the spirit of the present invention.

  [1] Referring to FIG. 5, Fin1, Fin2, and Fin3 show three examples of the maximum value (peak value) of the input load. In any of these three examples, when the displacement x of the movable piston 23 is 0, that is, the movable piston 23 is located at the highest compression ratio side and is at the stopper position abutted against the stopper 45 of the housing 22. In the ratio state, peak values Fin1 to Fin3 of the input load to the low compression ratio acting on the movable piston 23 from the variable compression ratio mechanism 10 side are applied to the movable piston 23 by the urging means such as the urging spring 44. Compared to the urging load Fspr toward the compression ratio side, it is small at low load (A) and large at high load (B).

  Therefore, as shown in FIG. 6, at the time of low load t0 to t1 where the setting of the high compression ratio is used, the urging load Fspr to the high compression ratio side is larger than the peak value of the combustion load Fin. The load F acting on the movable piston 23 always acts on the high compression ratio side without being reversed to the low compression ratio side (becomes a value smaller than 0). Accordingly, the movable piston 23 remains pressed against the high compression ratio side stopper 45, the holding performance of the movable piston 23 at this stopper position is improved, and rattling of the movable piston 23 is suppressed. For this reason, generation | occurrence | production of the noise by the shakiness of the movable piston 23 and energy consumption (fuel consumption deterioration) can be suppressed.

  On the other hand, at the acceleration time t1 when the engine load is increased from such a high compression ratio / low load operation state, the peak value of the combustion load Fin toward the low compression ratio side becomes the urging load as the load increases. Above Fspr, the movable piston 23 is pushed to the low compression ratio side. Therefore, the solenoid 35 of the switching valve 34 is turned ON to open the first drain pipe 27 of the first oil chamber 24, whereby the movable piston 23 can be quickly moved to the low compression ratio side, and to the low compression ratio side. Responsiveness can be improved, and the occurrence of knocking can be suppressed.

  In this way, both the retention in the high compression ratio state where the movable piston 23 is abutted against the stopper 45 and the responsiveness from the high compression ratio state to the low compression ratio side must be compatible at a high level. Is possible.

  [2] As described above, the moving speed of the movable piston 23, that is, the response speed of the engine compression ratio largely depends on the magnitude relationship between the peak value Fin of the input load and the urging load Fspr. Here, the urging load Fspr monotonously increases in proportion to the movement of the movable piston 23 toward the low compression ratio. Therefore, as shown in FIG. 7, as the movable piston 23 moves to the low compression ratio side, the peak value Fin of the input load due to the combustion load is monotonically increased, so that the position of the movable piston 23, that is, the low compression ratio side is increased. Regardless of the displacement x, the peak value Fin of the input load at the time of high load can always be made larger than the urging load Fspr. For this reason, at the time of high load, the movable piston 23 can be continuously pushed to the low compression ratio side until the movable piston 23 moves to the lowest compression ratio side, and the responsiveness of the movable piston 23 to the low compression ratio side is improved. Further improvement can be achieved.

  [3] FIG. 8 shows a link configuration that realizes such a monotonous increase in the peak value of the input load. As shown in FIG. 8B, in the low compression ratio state where the movable piston 23 is located on the low compression ratio side, the angle α formed by the lever 41 and the lever link 42 is an obtuse angle (90 to 180 °). As shown in (A), as the movable piston 23 moves to the high compression ratio side, the angle α becomes smaller. That is, as the movable piston 23 moves to the low compression ratio side, the angle α increases in an obtuse angle range, and the distance between the center of the control shaft 16 and the lever link 42, that is, the arm length decreases. The input load Fin increases.

  [4] As shown in FIG. 9, in the low compression ratio state (piston displacement k state) where the movable piston 23 is located on the low compression ratio side, the peak value Fin (k) of the input load at low load is given. Larger than the force load Fspr (k). That is, the peak value Fin of the input load is smaller than the urging load Fspr on the high compression ratio side, and approaches the urging load Fspr as the movable piston 23 moves to the low compression ratio side. Bigger than. That is, in the low compression ratio state where the movable piston 23 is located on the low compression ratio side, the peak value of the input load is always larger than the urging load Fspr (k) regardless of the magnitude of the load.

  As a result, as shown in the area surrounded by the broken line in FIG. 10, when the driver relaxes the accelerator or downshifts in a high load operation state using a low compression ratio setting such as during acceleration or high load. Even when the engine load is temporarily reduced due to the above, the biasing load Fspr on the high compression ratio side exceeds the peak value Fin of the input load, and the movable piston 23 is favorably moved to the low compression ratio side. Can operate and hold.

  [5] As described above, the moving speed of the movable piston 23, that is, the response speed of the engine compression ratio largely depends on the peak value Fin of the input load due to the combustion load and the urging load Fspr, and the urging load Fspr is 23 monotonously increases in proportion to the movement toward the low compression ratio. Therefore, as shown in FIG. 11, as the movable piston 23 moves to the low compression ratio side, the peak value Fin of the input load due to the combustion load is monotonously decreased.

  As a result, when the movable piston 23 is moved to the low compression ratio side in a state where it is transiently on the high load side, such as during sudden acceleration from high compression / low load, the movable piston 23 moves to the low compression ratio side. As it moves, the difference between the peak value Fin of the input load at high load and the biasing load Fspr decreases, and the load on the low compression ratio side acting on the movable piston 23 decreases. For this reason, as shown in FIG. 12, the seating speed to the low compression ratio side (when the movement of the movable piston 23 to the low compression ratio side is mechanically restricted by the stopper) is reduced, and the oil pressure suddenly increases. And the occurrence of abnormal noise can be suppressed. In addition, it is possible to reduce the load acting on the movable piston 23 in a low compression ratio and high load state, and to suppress an increase in oil temperature due to the reciprocating motion of the movable piston 23.

  [6] FIG. 13 shows an example of a link configuration that monotonously decreases the peak value Fin of the input load as described above. The angle α formed by the lever 41 and the lever link 42 in the high compression ratio state (A) is an acute angle (0 to 90 °), and the angle α is increased as the movable piston 23 moves to the low compression ratio side (B). growing. That is, as the movable piston 23 moves to the low compression ratio side, the angle α increases in an acute angle range, and the distance between the center of the control shaft 16 and the lever link 42, that is, the arm length increases. The input load Fin decreases monotonously.

  [7] As shown in FIG. 14, the peak value Fin of the input load is maximized at the position of the intermediate compression ratio. That is, the peak value Fin of the input load increases on the high compression ratio side and decreases on the low compression ratio side as the movable piston 23 moves to the low compression ratio side. In this way, by making the peak value of the input load due to the combustion load maximum at the intermediate compression ratio and making it a convex characteristic, the response speed in the intermediate compression ratio state is improved and the combustion in the low compression ratio state is performed. The peak value of the input load due to the load can be suppressed, and the deviation between the peak value of the input load due to the combustion load and the bias load, that is, the piston load F can be reduced. As a result, as in the case of [5] described above, the seating speed on the low compression ratio side can be reduced, an abrupt increase in oil pressure can be suppressed, and the occurrence of abnormal noise can be suppressed. In addition, it is possible to reduce the input load to the piston in a low compression ratio and high load state, and it is possible to suppress an increase in oil temperature due to a minute reciprocating motion of the piston.

  FIG. 15 shows an example of a link configuration that maximizes the peak value Fin of such input load at the position of the intermediate compression ratio. In this device, an auxiliary link 51 is interposed between the lever link 42 and the piston rod 43, and the auxiliary link 51 is supported on the engine body 53 side such as a cylinder block via an auxiliary lever 52. In the high compression ratio state (A), the angle α formed by the lever 41 and the lever link 42 becomes an acute angle, and the angle α increases as the movable piston 23 moves to the low compression ratio side, so that the low compression ratio is set. In the state (B), the angle α is an obtuse angle.

  [8] As shown in FIG. 16, the peak value Fin (kc) of the input load is made larger than the urging load Fspr (kc) when the movable piston is at a predetermined intermediate position kc at the time of low load. That is, the peak value Fin of the input load is set to exceed the biasing load Fspr regardless of the engine load in a predetermined piston position range including the intermediate position. Thus, as in the case of [4] above, in an operating state using a low compression ratio setting, such as during acceleration or high load, the engine is loosened due to the accelerator being loosened or downshifting, etc. Even when the load temporarily decreases, the peak value Fin of the input load is in a state larger than the biasing load Fspr, so that the movable piston 23 can be favorably operated and held toward the low compression ratio side. Therefore, the response to the low compression ratio side and the retention in the low compression ratio state can be improved.

  [9] As described above, the angle α formed by the lever 41 and the lever link 42 changes according to the position of the movable piston 23, so that the lever 41 and the lever link are also shown in FIG. With the simple link configuration using 42, the magnitudes of the input loads (peak values thereof) Fin1 to Fin3 can be changed as the movable piston 23 moves. As a result, the direction and magnitude of the piston load F that finally acts on the movable piston by the input load Fin and the biasing load Fspr and the like are made appropriate according to the position of the movable piston, that is, the engine compression ratio. Thus, it becomes possible to achieve both responsiveness and retention of the engine compression ratio by the movable piston 23.

  [10] As described above, the oil draining mechanism 20 does not move the piston by hydraulic pressure or electric power unlike a general actuator, but by switching the opening and closing of the drain lines 27 and 29 by the switching valve 34, the movable piston The movement of the movable piston 23 is held at a predetermined position by suppressing or prohibiting the movement of the movable piston 23 to one or both of them, and the movement of the movable piston 23 is caused by the input load Fin acting on the movable piston 23 or the biasing spring. The biasing load Fspr by 44 is used. For this reason, there is no loss of energy consumption as in the case where an actuator using hydraulic pressure or electric power as a drive source is used, and the configuration can be simplified.

  [11] The variable compression ratio mechanism typically has a low compression ratio such that the combustion pressure acting on the piston 14 is transferred to the movable piston 23 via the control shaft 16 like the one using the above-described multi-link type piston-crank mechanism. It acts on the side. Thereby, as described above, the response to the low compression ratio side of the movable piston 23 can be enhanced by the input load Fin to the low compression ratio side, and the link configuration between the control shaft 16 and the movable piston 23 is set. Thus, the magnitude of the input load Fin can be changed according to the position of the movable piston 23.

  [12] As shown in FIG. 17, apart from the oil draining mechanism 20, an actuator 60 that drives the movable piston 23 in the cylinder axial direction is provided in series with the oil draining mechanism 20. The actuator 60 is an electric type equipped with an electric motor 61, and a speed reduction mechanism using a ball screw that converts the rotational power of the output shaft 61A of the motor 61 into linear power while decelerating it and transmits it to the housing 22. And the entire housing 22 including the movable piston 23 is driven in the cylinder axial direction, thereby driving the movable piston 23 to the low compression ratio side or the high compression ratio side. The operation of the motor 61 is controlled by the control unit 38 described above. The reduction mechanism includes a ball screw shaft 62 having one end fixed to the housing 22, a first reduction gear 63 attached to the output shaft 61A of the motor 61, a second reduction gear 64 meshing with the first reduction gear, A second reduction gear 64 is attached and has a ball screw nut 65 that meshes with the ball screw shaft 62.

  For example, when a high response speed to the low compression ratio side is required and the variable range of the compression ratio is large, such as when transient knocking due to sudden acceleration is avoided, the oil draining mechanism 20 and the actuator 60 are In combination, the compression ratio is reduced and high responsiveness is not required as in slow acceleration, and when the variable compression ratio width is relatively small, either the actuator 60 or the oil draining mechanism 20 is used. Thus, the compression ratio of the movable piston 23 is reduced. Thus, by using the oil draining mechanism 20 and the actuator 60 in combination, it is possible to further improve the responsiveness and retention of the engine compression ratio. Moreover, by arranging the oil draining mechanism 20 and the actuator 60 in series, the stroke of the movable piston 23 of the oil draining mechanism 20 can be reduced, and the oil draining mechanism 20 can be downsized.

DESCRIPTION OF SYMBOLS 10 ... Variable compression ratio mechanism 14 ... Piston 16 ... Control shaft 20 ... Oil removal mechanism 21 ... Cylinder 22 ... Housing 23 ... Movable piston 34 ... Switching valve 38 ... Control part 41 ... Lever 42 ... Link lever 44 ... Energizing spring (attachment) Means)
45 ... Stopper 60 ... Actuator

Claims (14)

A housing in which a cylinder is formed;
A movable piston disposed in the cylinder so as to be reciprocally movable;
A variable compression ratio mechanism that is mechanically connected to the movable piston and changes the engine compression ratio as the movable piston moves along the cylinder axial direction;
Biasing means for biasing the movable piston toward the high compression ratio side,
In the high compression ratio state where the movable piston is located on the highest compression ratio side and abutted against the stopper of the housing, the peak of the input load from the variable compression ratio mechanism side to the low compression ratio side acting on the movable piston The variable compression ratio device for an internal combustion engine according to claim 1, wherein the value is smaller at the time of low load and larger at the time of high load than the biasing load applied to the movable piston by the biasing means. .   The variable compression ratio device for an internal combustion engine according to claim 1, wherein the peak value of the input load monotonously increases as the movable piston moves toward the low compression ratio. The variable compression ratio mechanism is such that the engine compression ratio changes as the control shaft rotates,
A link row connecting the control shaft and the movable piston has a lever having the control shaft as a fulcrum, and a lever link having one end connected to the tip of the lever.
In the low compression ratio state where the movable piston is located on the low compression ratio side, the angle formed by the lever and the lever link is an obtuse angle, and the angle decreases as the movable piston moves to the high compression ratio side. The variable compression ratio device for an internal combustion engine according to claim 1 or 2.   The peak value of the said input load at the time of low load is larger than the said urging load in the low compression ratio state in which the said movable piston is located in the low compression ratio side, The any one of Claims 1-3 characterized by the above-mentioned. Variable compression ratio device for internal combustion engine.   The variable compression ratio device for an internal combustion engine according to claim 1, wherein the peak value of the input load monotonously decreases as the movable piston moves toward the low compression ratio side. The variable compression ratio mechanism is such that the engine compression ratio changes as the control shaft rotates,
A link row connecting the control shaft and the movable piston has a lever having the control shaft as a fulcrum, and a lever link having one end connected to the tip of the lever.
The angle formed between the lever and the lever link in the high compression ratio state is an acute angle, and the angle increases as the movable piston moves to the low compression ratio side. Variable compression ratio device for an internal combustion engine.   2. The internal combustion engine according to claim 1, wherein the peak value of the input load increases on the high compression ratio side and decreases on the low compression ratio side as the movable piston moves to the low compression ratio side. Variable compression ratio device.   8. The variable compression ratio of the internal combustion engine according to claim 1, wherein a peak value of the input load is larger than an urging load at least when the movable piston is at a predetermined intermediate position at low load. apparatus. A housing in which a cylinder is formed;
A movable piston disposed in the cylinder so as to be reciprocally movable;
A variable compression ratio mechanism that is mechanically coupled to the movable piston and changes the engine compression ratio as the movable piston moves;
Biasing means for biasing the movable piston in the direction opposite to the direction of the input load acting on the movable piston from the variable compression ratio mechanism side,
A variable compression ratio device for an internal combustion engine, characterized in that the magnitude of the input load changes as the movable piston moves. In the cylinder of the housing, it has an oil draining mechanism provided with at least one oil chamber partitioned by the movable piston,
This oil draining mechanism
An inlet pipe for supplying hydraulic oil to the oil chamber;
A drain line for discharging hydraulic oil from the oil chamber;
A switching valve that opens and closes the drain pipe;
The variable compression ratio device for an internal combustion engine according to any one of claims 1 to 9 , characterized by comprising: In the cylinder of the housing, there is an oil drain mechanism provided with a first oil chamber and a second oil chamber partitioned by the movable piston,
This oil draining mechanism
A first inlet pipe for supplying hydraulic oil to the first oil chamber;
A first exhaust pipe for discharging hydraulic oil from the first oil chamber;
A second inlet pipe for supplying hydraulic oil to the second oil chamber;
A second exhaust pipe for discharging hydraulic oil from the second oil chamber;
A switching valve that opens and closes the first drain pipe and the second pipe duct;
The variable compression ratio device for an internal combustion engine according to any one of claims 1 to 9 , characterized by comprising: The switching valve has a first position for closing the first drain pipe and opening the second pipe, a second position for opening the first pipe and closing the second pipe, and a first pipe. The variable compression ratio device for an internal combustion engine according to claim 11 , wherein the variable compression ratio device is a three-position switching valve that switches between a third position that closes both the passage and the second drain pipe. The variable compression ratio mechanism is
A plurality of links connecting the piston of the internal combustion engine and the crankpin of the crankshaft;
A control shaft rotatably supported by the engine body and coupled to the movable piston;
An eccentric shaft provided eccentric to the control shaft;
One end connected to one of the plurality of links, the other end according to any one of claims 1 to 12, characterized in that and a control link rotatably connected to the eccentric shaft portion internal combustion Variable compression ratio device for engines. Variable compression ratio device for an internal combustion engine according to any one of claims 1 to 13, characterized in that an actuator for driving the movable piston in a cylinder axis direction.

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