JP6256367B2 - Variable compression ratio internal combustion engine - Google Patents

Description

  The present invention relates to a variable compression ratio internal combustion engine capable of changing a mechanical compression ratio.

  Conventionally, an internal combustion engine having a variable compression ratio mechanism capable of changing the mechanical compression ratio of the internal combustion engine is known. Various types of such variable compression ratio mechanisms have been proposed, and one of them is one that changes the effective length of a connecting rod used in an internal combustion engine (for example, Patent Document 1). Here, the effective length of the connecting rod means the distance between the center of the crank receiving opening for receiving the crank pin and the center of the piston pin receiving opening for receiving the piston pin. Therefore, when the effective length of the connecting rod is increased, the combustion chamber volume when the piston is at the compression top dead center is reduced, and thus the mechanical compression ratio is increased. On the other hand, when the effective length of the connecting rod is shortened, the combustion chamber volume when the piston is at the compression top dead center is increased, and thus the mechanical compression ratio is lowered.

  As a variable-length connecting rod capable of changing the effective length, one having an eccentric member (an eccentric arm or an eccentric sleeve) that is rotatable with respect to the connecting rod body is known at the small-diameter end of the connecting rod body (for example, Patent Document 1). The eccentric member has a piston pin receiving opening for receiving the piston pin, and the piston pin receiving opening is provided eccentric to the rotation axis of the eccentric member. Two piston mechanisms are connected to the eccentric member. Each piston mechanism includes a hydraulic cylinder formed in a connecting rod body of a variable-length connecting rod, and a hydraulic piston that can slide in the cylinder.

  In such a variable length connecting rod, when the rotational position of the eccentric member is changed, the effective length of the connecting rod can be changed accordingly. Specifically, the eccentric member lengthens the effective length of the connecting rod by rotating in one direction by an upward inertia force acting on the piston pin by the reciprocation of the piston. As a result, the piston rises with respect to the connecting rod body, and the mechanical compression ratio is switched from the low compression ratio to the high compression ratio. At this time, hydraulic fluid is supplied from one hydraulic cylinder into the other hydraulic cylinder, and the mechanical compression ratio is maintained at a high compression ratio. On the other hand, the eccentric member rotates in the other direction by the downward inertial force acting on the piston pin by the reciprocating motion of the piston and the downward explosive force acting on the piston pin by the combustion of the air-fuel mixture. Shorten the length. As a result, the piston descends with respect to the connecting rod body, and the mechanical compression ratio is switched from the high compression ratio to the low compression ratio. At this time, hydraulic fluid is supplied into the one hydraulic cylinder from the other hydraulic cylinder, and the mechanical compression ratio is maintained at a low compression ratio.

  Therefore, in a variable compression ratio internal combustion engine having a variable length connecting rod, the mechanical compression ratio is switched from a low compression ratio to a high compression ratio by inertial force, and is switched from a high compression ratio to a low compression ratio by inertial force and explosive force. . The mechanical compression ratio is maintained at a high compression ratio or a low compression ratio by the hydraulic oil supplied into the hydraulic cylinder.

JP 2011-196549 A JP-A-6-229315

  By the way, the inertial force is much smaller than the explosive force. For this reason, it is difficult to obtain sufficient response when the mechanical compression ratio is switched from the low compression ratio to the high compression ratio. Further, since the inertial force is proportional to the square of the engine speed of the internal combustion engine, a sufficient inertial force cannot be obtained in the low rotation range of the internal combustion engine, and the responsiveness is further deteriorated.

  It is also known to dispose an oil seal between the hydraulic cylinder and the hydraulic piston in order to retain the hydraulic oil in the hydraulic cylinder. In order to improve the responsiveness, it is conceivable to reduce the frictional force generated between the hydraulic cylinder and the oil seal. However, if the frictional force is reduced, the sealability of the oil seal necessary to maintain the mechanical compression ratio at a high compression ratio or a low compression ratio may not be ensured.

  Accordingly, in view of the above problems, an object of the present invention is to provide a sealing property of an oil seal necessary for maintaining a mechanical compression ratio at a high compression ratio or a low compression ratio in a variable compression ratio internal combustion engine having a variable length connecting rod. It is to improve the responsiveness when the mechanical compression ratio is switched from the low compression ratio to the high compression ratio.

  In order to solve the above problems, a variable compression ratio internal combustion engine capable of changing a mechanical compression ratio, comprising a cylinder, a piston reciprocating in the cylinder, and a connecting rod connected to the piston via a piston pin. The connecting rod receives a piston pin, and a connecting rod body having a large diameter end portion provided with a crank receiving opening for receiving the crank pin and a small diameter end portion located on the piston side opposite to the large diameter end portion. An eccentric member having a piston pin receiving opening that is pivotally attached to a small-diameter end, a hydraulic cylinder that is provided in the connecting rod body and that is supplied with hydraulic oil, and a hydraulic piston that slides within the hydraulic cylinder; A hydraulic piston mechanism having an oil seal fixed to the hydraulic piston and in contact with the inner surface of the hydraulic cylinder. The axis of the pin receiving opening is configured to be decentered from the rotation axis of the eccentric member, and by rotating in one direction, the piston is raised with respect to the connecting rod body and rotated in the other direction. The piston is configured to be lowered with respect to the connecting rod body, and the hydraulic piston rises in the hydraulic cylinder when the eccentric member rotates in one direction and is hydraulic when the eccentric member rotates in the other direction. The cylinder is lowered and the inner diameter of the hydraulic cylinder is larger than the position of the oil seal after the hydraulic piston is raised in the hydraulic cylinder. The variable compression is larger at the position of the oil seal after the hydraulic piston is lowered in the hydraulic cylinder. A specific internal combustion engine is provided.

  According to the present invention, in a variable compression ratio internal combustion engine having a variable length connecting rod, the mechanical compression ratio is ensured while ensuring the sealability of the oil seal necessary for maintaining the mechanical compression ratio at a high compression ratio or a low compression ratio. Can be improved when switching from a low compression ratio to a high compression ratio.

FIG. 1 is a schematic side sectional view of a variable compression ratio internal combustion engine. FIG. 2 is a perspective view schematically showing a variable length connecting rod according to the present invention. FIG. 3 is a sectional side view schematically showing a variable length connecting rod and a piston according to the present invention. FIG. 4 is a schematic exploded perspective view of the vicinity of the small diameter end of the connecting rod body. FIG. 5 is a schematic exploded perspective view of the vicinity of the small diameter end of the connecting rod body. FIG. 6 is a sectional side view schematically showing a variable length connecting rod and a piston according to the present invention. FIG. 7 is a cross-sectional side view of the connecting rod in which the region where the flow direction switching mechanism is provided is enlarged. 8 is a cross-sectional view of the connecting rod taken along lines VIII-VIII and IX-IX in FIG. FIG. 9 is a schematic diagram for explaining the operation of the flow direction switching mechanism when the hydraulic pressure is supplied from the hydraulic supply source to the switching pin. FIG. 10 is a schematic diagram for explaining the operation of the flow direction switching mechanism when no hydraulic pressure is supplied from the hydraulic supply source to the switching pin. FIG. 11 is an enlarged plan view of the first oil seal. FIG. 12 is an enlarged cross-sectional side view schematically showing the first cylinder. FIG. 13 is an enlarged cross-sectional side view schematically showing the second cylinder. FIG. 14 is a sectional side view schematically showing a variable length connecting rod and a piston according to a second embodiment of the present invention.

  Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. In the following description, the same reference numerals are assigned to similar components.

<First Embodiment>
First, a first embodiment of the present invention will be described with reference to FIGS.

<Variable compression ratio internal combustion engine>
FIG. 1 shows a schematic side sectional view of a variable compression ratio internal combustion engine according to the present invention.
Referring to FIG. 1, reference numeral 1 denotes an internal combustion engine. The internal combustion engine 1 includes a crankcase 2, a cylinder block 3, a cylinder head 4, a piston 5, a variable length connecting rod 6, a combustion chamber 7, a spark plug 8 disposed in the center of the top surface of the combustion chamber 7, an intake valve 9, an intake air A camshaft 10, an intake port 11, an exhaust valve 12, an exhaust camshaft 13, and an exhaust port 14 are provided. The cylinder block 3 defines a cylinder 15. The piston 5 reciprocates in the cylinder 15. The internal combustion engine 1 further includes a variable valve timing mechanism A that can control the valve opening timing and the valve closing timing of the intake valve 9, and a variable valve timing mechanism that can control the valve opening timing and the valve closing timing of the exhaust valve 12. B.

  The variable length connecting rod 6 is connected to the piston 5 via the piston pin 21 at the small diameter end portion thereof, and is connected to the crank pin 22 of the crankshaft at the large diameter end portion thereof. As will be described later, the variable length connecting rod 6 can change the distance from the axis of the piston pin 21 to the axis of the crank pin 22, that is, the effective length.

  When the effective length of the variable length connecting rod 6 is increased, the length from the crank pin 22 to the piston pin 21 is increased, so that the combustion chamber 7 when the piston 5 is at the top dead center as shown by the solid line in the figure. The volume of becomes smaller. On the other hand, even if the effective length of the variable-length connecting rod 6 changes, the stroke length that the piston 5 reciprocates in the cylinder does not change. Therefore, at this time, the mechanical compression ratio in the internal combustion engine 1 is increased.

  On the other hand, if the effective length of the variable-length connecting rod 6 is shortened, the length from the crank pin 22 to the piston pin 21 is shortened, so that the combustion when the piston 5 is at the top dead center as shown by the broken line in the figure. The volume in the chamber 7 is increased. However, as described above, the stroke length of the piston 5 is constant. Therefore, at this time, the mechanical compression ratio in the internal combustion engine 1 becomes small.

<Configuration of variable length connecting rod>
FIG. 2 is a perspective view schematically showing the variable length connecting rod 6 according to the present invention, and FIG. 3 is a sectional side view schematically showing the variable length connecting rod 6 according to the present invention. As shown in FIGS. 2 and 3, the variable length connecting rod 6 includes a connecting rod body 31, an eccentric member 32 rotatably attached to the connecting rod body 31, and a first piston mechanism 33 provided on the connecting rod body 31. And a second piston mechanism 34, and a flow direction switching mechanism 35 for switching the flow of hydraulic oil to both the piston mechanisms 33, 34.

  First, the connecting rod body 31 will be described. The connecting rod body 31 has a crank receiving opening 41 for receiving the crank pin 22 of the crankshaft at one end thereof, and a sleeve receiving opening 42 for receiving a sleeve of an eccentric member 32 described later at the other end. Since the crank receiving opening 41 is larger than the sleeve receiving opening 42, the end of the connecting rod body 31 located on the side where the crank receiving opening 41 is provided (crankshaft side) is referred to as a large diameter end 31a. The end of the connecting rod body 31 located on the side where the opening 42 is provided (piston side) is referred to as a small diameter end 31b.

  In the present specification, the center axis of the crank receiving opening 41 (that is, the axis of the crank pin 22 received in the crank receiving opening 41) and the center axis of the sleeve receiving opening 42 (that is, received in the sleeve receiving opening 42). A line X (FIG. 3) extending to the center of the connecting rod body 31 is referred to as an axis of the connecting rod 6. The length of the connecting rod in the direction perpendicular to the axis X of the connecting rod 6 and perpendicular to the central axis of the crank receiving opening 41 is referred to as the connecting rod width. In addition, the length of the connecting rod in the direction parallel to the central axis of the crank receiving opening 41 is referred to as the connecting rod thickness.

  As can be seen from FIGS. 2 and 3, the width of the connecting rod body 31 is narrowest at the intermediate portion between the large-diameter end portion 31a and the small-diameter end portion 31b. Moreover, the width | variety of the large diameter edge part 31a is wider than the width | variety of the small diameter edge part 31b. On the other hand, the thickness of the connecting rod body 31 is substantially constant except for the region where the piston mechanisms 33 and 34 are provided.

  Next, the eccentric member 32 will be described. 4 and 5 are schematic perspective views of the vicinity of the small-diameter end 31b of the connecting rod body 31. FIG. 4 and 5, the eccentric member 32 is shown in an exploded state. 2 to 5, the eccentric member 32 includes a cylindrical sleeve 32 a that is received in a sleeve receiving opening 42 formed in the connecting rod body 31, and one direction in the width direction of the connecting rod body 31 from the sleeve 32 a. And a pair of second arms 32c extending from the sleeve 32a in the width direction of the connecting rod body 31 in the other direction (a direction generally opposite to the one direction). Since the sleeve 32 a is rotatable in the sleeve receiving opening 42, the eccentric member 32 is attached to the connecting rod body 31 so as to be rotatable in the circumferential direction of the small diameter end portion 31 b at the small diameter end portion 31 b of the connecting rod body 31. become. The rotational axis of the eccentric member 32 coincides with the central axis of the sleeve receiving opening 42.

  The sleeve 32 a of the eccentric member 32 has a piston pin receiving opening 32 d for receiving the piston pin 21. The piston pin receiving opening 32d is formed in a cylindrical shape. The cylindrical piston pin receiving opening 32d is formed so that its axis is parallel to the central axis of the cylindrical outer shape of the sleeve 32a, but not coaxial. Therefore, the axis of the piston pin receiving opening 32d is eccentric from the central axis of the cylindrical outer shape of the sleeve 32a, that is, the rotational axis of the eccentric member 32.

  Thus, in the present embodiment, the central axis of the piston pin receiving opening 32d of the sleeve 32a is eccentric from the rotational axis of the eccentric member 32. For this reason, when the eccentric member 32 rotates, the position of the piston pin receiving opening 32d in the sleeve receiving opening 42 changes. When the position of the piston pin receiving opening 32d is on the large diameter end portion 31a side in the sleeve receiving opening 42, the effective length of the connecting rod is shortened. On the contrary, when the position of the piston pin receiving opening 32d in the sleeve receiving opening 42 is opposite to the large diameter end portion 31a side, that is, on the small diameter end portion 31b side, the effective length of the connecting rod becomes long. Therefore, according to this embodiment, the effective length of the connecting rod 6 changes by rotating the eccentric member.

  Next, the first piston mechanism 33 will be described with reference to FIG. The first piston mechanism 33 is provided in the connecting rod body 31 and is supplied with hydraulic oil, a first piston 33b that slides within the first cylinder 33a, a first cylinder 33a, and a first piston 33b. And a first oil seal 33c disposed between them. Most or all of the first cylinder 33 a is disposed on the first arm 32 b side with respect to the axis X of the connecting rod 6. In addition, the first cylinder 33a is disposed so as to be inclined with respect to the axis X so as to protrude in the width direction of the connecting rod body 31 as it approaches the small diameter end portion 31b. The first cylinder 33 a communicates with the flow direction switching mechanism 35 via the first piston communication oil passage 51.

  The first piston 33 b is connected to the first arm 32 b of the eccentric member 32 by the first connecting member 45. The first piston 33b is rotatably connected to the first connecting member 45 by a pin. The first arm 32b is rotatably connected to the first connecting member 45 by a pin at the end opposite to the side connected to the sleeve 32a.

  The first oil seal 33c has a ring shape and is fixed around the lower end portion of the first piston 33b. The first oil seal 33c contacts the inner surface of the first cylinder 33a and seals the hydraulic oil supplied into the first cylinder 33a. For this reason, a frictional force is generated between the first oil seal 33c and the first cylinder 33a.

  FIG. 11 is an enlarged plan view of the first oil seal 33c. As shown in FIG. 11, the first oil seal 33c has an inner layer 33d made of an elastic body such as rubber and an outer layer 33e made of resin. The outer layer 33e is located radially outside the inner layer 33d and contacts the inner surface of the first cylinder 33a. Since the friction coefficient of the resin is generally smaller than the friction coefficient of the elastic body, the friction force generated between the first oil seal 33c and the first cylinder 33a is reduced by the outer layer 33e made of resin. Moreover, the required sealing property of the 1st oil seal 33c is securable by the inner layer 33d comprised from the elastic body.

  Next, the second piston mechanism 34 will be described. The second piston mechanism 34 is provided in the connecting rod body 31 and is supplied with hydraulic oil, a second piston 34b that slides in the second cylinder 34a, a second cylinder 34a, and a second piston 34b. And a second oil seal 34c disposed between the two. Most or all of the second cylinder 34 a is arranged on the second arm 32 c side with respect to the axis X of the connecting rod 6. Further, the second cylinder 34a is disposed so as to be inclined with respect to the axis X so as to protrude in the width direction of the connecting rod body 31 as it approaches the small diameter end portion 31b. Further, the second cylinder 34 a communicates with the flow direction switching mechanism 35 via the second piston communication oil passage 52.

  The second piston 34 b is connected to the second arm 32 c of the eccentric member 32 by the second connecting member 46. The second piston 34b is rotatably connected to the second connecting member 46 by a pin. The second arm 32c is rotatably connected to the second connecting member 46 by a pin at the end opposite to the side connected to the sleeve 32a.

  The second oil seal 34c has a ring shape and is fixed around the lower end portion of the second piston 34b. The second oil seal 34c contacts the inner surface of the second cylinder 34a and seals the hydraulic oil supplied into the second cylinder 34a. For this reason, a frictional force is generated between the second oil seal 34c and the second cylinder 34a.

  Similar to the first oil seal 33c, the second oil seal 34c has an inner layer 34d made of an elastic body such as rubber and an outer layer 34e made of resin. The outer layer 34e is located radially outside the inner layer 34d and contacts the inner surface of the second cylinder 34a. Since the friction coefficient of the resin is generally smaller than that of the elastic body, the outer layer 34e made of resin reduces the friction force generated between the second oil seal 34c and the second cylinder 34a. Further, the required sealing performance of the second oil seal 34c can be ensured by the inner layer 34d formed of an elastic body.

  The first piston 33b and the second piston 34b have the same shape and configuration, and the first oil seal 33c and the second oil seal 34c have the same shape and configuration. This makes it possible to share parts.

<Operation of variable length connecting rod>
Next, operations of the eccentric member 32, the first piston mechanism 33, and the second piston mechanism 34 thus configured will be described with reference to FIG. FIG. 6A shows a state where hydraulic oil is supplied into the first cylinder 33 a of the first piston mechanism 33 and no hydraulic oil is supplied into the second cylinder 34 a of the second piston mechanism 34. . On the other hand, FIG. 6B shows that hydraulic oil is not supplied into the first cylinder 33 a of the first piston mechanism 33 and hydraulic oil is supplied into the second cylinder 34 a of the second piston mechanism 34. It shows the state.

  Here, as will be described later, the flow direction switching mechanism 35 prohibits the flow of hydraulic oil from the first cylinder 33a to the second cylinder 34a and allows the flow of hydraulic oil from the second cylinder 34a to the first cylinder 33a. Between the first state that is permitted and the second state that permits the flow of hydraulic oil from the first cylinder 33a to the second cylinder 34a and prohibits the flow of hydraulic oil from the second cylinder 34a to the first cylinder 33a. It can be switched with.

  When the flow direction switching mechanism 35 is in a first state that prohibits the flow of hydraulic oil from the first cylinder 33a to the second cylinder 34a and permits the flow of hydraulic oil from the second cylinder 34a to the first cylinder 33a, As shown in FIG. 6A, the hydraulic oil is supplied into the first cylinder 33a, and the hydraulic oil is discharged from the second cylinder 34a. For this reason, the first piston 33b rises, and the first arm 32b of the eccentric member 32 connected to the first piston 33b also rises. On the other hand, the second piston 34b is lowered, and the second arm 32c connected to the second piston 34b is also lowered. As a result, in the example shown in FIG. 6A, the eccentric member 32 is rotated in the direction of the arrow in the figure, and as a result, the position of the piston pin receiving opening 32d is raised. Therefore, the length between the center of the crank receiving opening 41 and the center of the piston pin receiving opening 32d, that is, the effective length of the connecting rod 6 is increased to L1 in the figure. That is, when the hydraulic oil is supplied into the first cylinder 33a and the hydraulic oil is discharged from the second cylinder 34a, the effective length of the connecting rod 6 is increased.

  On the other hand, the flow direction switching mechanism 35 is in a second state that permits the flow of hydraulic oil from the first cylinder 33a to the second cylinder 34a and prohibits the flow of hydraulic oil from the second cylinder 34a to the first cylinder 33a. Then, as shown in FIG. 6B, the hydraulic oil is supplied into the second cylinder 34a, and the hydraulic oil is discharged from the first cylinder 33a. For this reason, the second piston 34b rises, and the second arm 32c of the eccentric member 32 connected to the second piston 34b also rises. On the other hand, the first piston 33b is lowered, and the first arm 32b connected to the first piston 33b is also lowered. As a result, in the example shown in FIG. 6B, the eccentric member 32 is rotated in the direction of the arrow in the figure (the direction opposite to the arrow in FIG. 6A), and as a result, the piston pin receiving opening 32d. The position of goes down. Therefore, the length between the center of the crank receiving opening 41 and the center of the piston pin receiving opening 32d, that is, the effective length of the connecting rod 6 is L2 shorter than L1 in the drawing. That is, when the hydraulic oil is supplied into the second cylinder 34a and the hydraulic oil is discharged from the first cylinder 33a, the effective length of the connecting rod 6 is shortened.

  In the connecting rod 6 according to the present embodiment, as described above, the effective length of the connecting rod 6 is switched between L1 and L2 by switching the flow direction switching mechanism 35 between the first state and the second state. be able to. As a result, in the internal combustion engine 1 using the connecting rod 6, the mechanical compression ratio can be changed.

  Here, when the flow direction switching mechanism 35 is in the first state, the first piston 33b and the second piston 34b are basically shown in FIG. It moves to the position shown in A), and the eccentric member 32 rotates to the position shown in FIG. When an upward inertia force due to the reciprocating motion of the piston 5 in the cylinder 15 of the internal combustion engine 1 acts on the piston pin 21, the first piston 33b rises and the second piston 34b falls. At this time, the hydraulic oil is discharged from the second cylinder 34a, the hydraulic oil is supplied into the first cylinder 33a, and the first piston 33b and the second piston 34b move to the positions shown in FIG. 6A. . Further, when an upward inertia force acts on the piston pin 21, the eccentric member 32 rotates in one direction (the direction of the arrow in FIG. 6A) to the position shown in FIG. As a result, the effective length of the connecting rod 6 increases, and the piston 5 rises with respect to the connecting rod body 31. On the other hand, when the piston 5 reciprocates in the cylinder 15 of the internal combustion engine 1 and a downward inertia force acts on the piston pin 21, or when the air-fuel mixture burns in the combustion chamber 7 and the downward force acts on the piston pin 21. When the act acts, the first piston 33b tends to descend and the eccentric member 32 tries to rotate in the other direction (the direction of the arrow in FIG. 6B). However, since the flow of the working oil from the first cylinder 33a to the second cylinder 34a is prohibited by the flow direction switching mechanism 35, the working oil in the first cylinder 33a does not flow out, and therefore the first piston 33b and the eccentricity are not. The member 32 does not move.

  On the other hand, even when the flow direction switching mechanism 35 is in the second state, the eccentric member 32 is basically shown in FIG. 6B as described below without supplying hydraulic oil from the outside. The first piston 33b and the second piston 34b are moved to the positions shown in FIG. 6B. When the downward inertia force due to the reciprocating motion of the piston 5 in the cylinder 15 of the internal combustion engine 1 and the downward explosion force due to the combustion of the air-fuel mixture in the combustion chamber 7 act on the piston pin 21, the first piston 33 b While descending, the second piston 34b rises. At this time, the hydraulic oil is discharged from the first cylinder 33a, the hydraulic oil is supplied into the second cylinder 34a, and the first piston 33b and the second piston 34b move to the positions shown in FIG. 6B. . Further, when downward inertial force and explosive force act on the piston pin 21, the eccentric member 32 rotates to the position shown in FIG. 6B in the other direction (the direction of the arrow in FIG. 6B). . As a result, the effective length of the connecting rod 6 is shortened, and the piston 5 is lowered with respect to the connecting rod body 31. On the other hand, when the piston 5 reciprocates in the cylinder 15 of the internal combustion engine 1 and an upward inertial force acts on the piston pin 21, the second piston 34b attempts to descend and the eccentric member 32 moves in the one direction ( An attempt is made to rotate in the direction of the arrow in FIG. However, since the flow of hydraulic oil from the second cylinder 34a to the first cylinder 33a is prohibited by the flow direction switching mechanism 35, the hydraulic oil in the second cylinder 34a does not flow out, and thus the second piston 34b and the eccentricity are not. The member 32 does not move.

  Therefore, in the internal combustion engine 1, the mechanical compression ratio is switched from a low compression ratio to a high compression ratio by inertial force, and is switched from a high compression ratio to a low compression ratio by inertial force and explosive force.

<Configuration of flow direction switching mechanism>
Next, the configuration of the flow direction switching mechanism 35 will be described with reference to FIGS. FIG. 7 is a cross-sectional side view of the connecting rod in which the region where the flow direction switching mechanism 35 is provided is enlarged. 8A is a cross-sectional view of the connecting rod taken along line VIII-VIII in FIG. 7, and FIG. 8B is a cross-sectional view of the connecting rod taken along line IX-IX in FIG. 7. As described above, the flow direction switching mechanism 35 prohibits the flow of hydraulic oil from the first cylinder 33a to the second cylinder 34a and permits the flow of hydraulic oil from the second cylinder 34a to the first cylinder 33a. Switching between one state and a second state that permits the flow of hydraulic oil from the first cylinder 33a to the second cylinder 34a and prohibits the flow of hydraulic oil from the second cylinder 34a to the first cylinder 33a. It is a mechanism to perform.

  As shown in FIG. 7, the flow direction switching mechanism 35 includes two switching pins 61 and 62 and one check valve 63. The two switching pins 61 and 62 and the check valve 63 are disposed between the first cylinder 33 a and the second cylinder 34 a and the crank receiving opening 41 in the axis X direction of the connecting rod body 31. The check valve 63 is disposed closer to the crank receiving opening 41 than the two switching pins 61 and 62 in the direction of the axis X of the connecting rod body 31.

  Further, the two switching pins 61 and 62 are provided on both sides with respect to the axis X of the connecting rod body 31 and the check valve 63 is provided on the axis X. Thereby, it can suppress that the weight balance of the right and left of the connecting rod main body 31 falls by providing the switching pins 61 and 62 and the check valve 63 in the connecting rod main body 31.

  The two switching pins 61 and 62 are housed in cylindrical pin housing spaces 64 and 65, respectively. In the present embodiment, the pin accommodating spaces 64 and 65 are formed such that the axis thereof extends in parallel with the central axis of the crank receiving opening 41. The switching pins 61 and 62 are slidable in the direction in which the pin accommodating space 64 extends in the pin accommodating spaces 64 and 65. That is, the switching pins 61 and 62 are disposed in the connecting rod body 31 so that the operating direction thereof is parallel to the central axis of the crank receiving opening 41.

  Moreover, the 1st pin accommodation space 64 which accommodates the 1st switching pin 61 among the two pin accommodation spaces 64 and 65 is open with respect to one side surface of the connecting rod main body 31, as shown to FIG. 8 (A). And a pin receiving hole which is closed with respect to the other side surface of the connecting rod body 31. In addition, the second pin accommodating space 65 that accommodates the second switching pin 62 out of the two pin accommodating spaces 64 and 65 corresponds to the other side surface of the connecting rod body 31 as shown in FIG. And is formed as a pin receiving hole that is open and closed with respect to the one side surface.

  The first switching pin 61 has two circumferential grooves 61a and 61b extending in the circumferential direction. These circumferential grooves 61 a and 61 b are communicated with each other by a communication path 61 c formed in the first switching pin 61. A first urging spring 67 is accommodated in the first pin accommodating space 64, and the first switching pin 61 is urged in a direction parallel to the central axis of the crank receiving opening 41 by the first urging spring 67. It is energized. In particular, in the example illustrated in FIG. 8A, the first switching pin 61 is urged toward the closed end of the first pin accommodating space 64.

  Similarly, the second switching pin 62 also has two circumferential grooves 62a and 62b extending in the circumferential direction. These circumferential grooves 62 a and 62 b are communicated with each other by a communication path 62 c formed in the second switching pin 62. A second urging spring 68 is accommodated in the second pin accommodating space 65, and the second urging spring 68 causes the second switching pin 62 to be applied in a direction parallel to the central axis of the crank receiving opening 41. It is energized. In particular, in the example illustrated in FIG. 8A, the second switching pin 62 is urged toward the closed end of the second pin housing space 65. As a result, the second switching pin 62 is biased in the opposite direction to the first switching pin 61.

  In addition, the first switching pin 61 and the second switching pin 62 are disposed in opposite directions in a direction parallel to the central axis of the crank receiving opening 41. In addition, the second switching pin 62 is urged in the opposite direction to the first switching pin 61. For this reason, in this embodiment, when hydraulic pressure is supplied to the first switching pin and the second switching pin 62, the operating directions of the first switching pin 61 and the second switching pin 62 are opposite to each other.

  The check valve 63 is accommodated in a cylindrical check valve accommodation space 66. In the present embodiment, the check valve accommodating space 66 is also formed so as to extend in parallel with the central axis of the crank receiving opening 41. The check valve 63 can move in the direction in which the check valve accommodation space 66 extends in the check valve accommodation space 66. Therefore, the check valve 63 is disposed in the connecting rod body 31 so that the operating direction thereof is parallel to the central axis of the crank receiving opening 41. The check valve accommodation space 66 is formed as a check valve accommodation hole that is open to one side surface of the connecting rod body 31 and is closed to the other side surface of the connecting rod body 31.

  The check valve 63 permits the flow from the primary side (upper side in FIG. 8B) to the secondary side (lower side in FIG. 8B) and prohibits the flow from the secondary side to the primary side. Configured as follows.

  The first pin accommodating space 64 that accommodates the first switching pin 61 is communicated with the first cylinder 33 a via the first piston communication oil passage 51. As shown in FIG. 8A, the first piston communication oil passage 51 is communicated with the first pin housing space 64 in the vicinity of the center of the connecting rod body 31 in the thickness direction. The second pin housing space 65 that houses the second switching pin 62 is communicated with the second cylinder 34 a via the second piston communication oil passage 52. As shown in FIG. 8A, the second piston communication oil passage 52 is also connected to the second pin housing space 65 in the vicinity of the center of the connecting rod body 31 in the thickness direction.

  The first piston communication oil passage 51 and the second piston communication oil passage 52 are formed by cutting from the crank receiving opening 41 with a drill or the like. Therefore, on the crank receiving opening 41 side of the first piston communication oil passage 51 and the second piston communication oil passage 52, the first extension oil passage 51a and the second extension oil passage 52a that are coaxial with the piston communication oil passages 51, 52 are provided. Is formed. In other words, the first piston communication oil passage 51 and the second piston communication oil passage 52 are formed such that the crank receiving opening 41 is positioned on the extension line. The first extension oil passage 51a and the second extension oil passage 52a are closed by a bearing metal 71 provided in the crank receiving opening 41, for example.

  The first pin accommodation space 64 that accommodates the first switching pin 61 is communicated with the check valve accommodation space 66 via the two space communication oil passages 53 and 54. Among these, as shown in FIG. 8 (A), one of the first space communication oil passages 53 is on one side surface side (lower side in FIG. 8 (B)) from the center in the thickness direction of the connecting rod body 31. The first pin accommodating space 64 and the check valve accommodating space 66 are communicated with the secondary side. The other second space communication oil passage 54 has a first pin accommodation space 64 and a check valve accommodation space 66 on the other side surface side (upper side in FIG. 8B) than the center in the thickness direction of the connecting rod body 31. To the primary side. In addition, the first space communication oil passage 53 and the second space communication oil passage 54 are configured such that the distance between the first space communication oil passage 53 and the first piston communication oil passage 51 in the connecting rod main body thickness direction and the second space communication oil passage 53 are the same. The distance in the connecting rod body thickness direction between the oil passage 54 and the first piston communication oil path 51 is arranged to be equal to the distance in the connecting rod body thickness direction between the circumferential grooves 61a and 61b.

  The second pin housing space 65 that houses the second switching pin 62 is communicated with the check valve housing space 66 through the two space communication oil passages 55 and 56. Among these, as shown in FIG. 8 (A), one third space communication oil passage 55 is on one side surface side (lower side in FIG. 8 (B)) from the center in the thickness direction of the connecting rod body 31. The first pin accommodating space 64 and the check valve accommodating space 66 are communicated with the secondary side. The other fourth space communication oil passage 56 has a first pin accommodation space 64 and a check valve accommodation space 66 on the other side surface side (upper side in FIG. 8B) than the center in the thickness direction of the connecting rod body 31. To the primary side. In addition, the third space communication oil passage 55 and the fourth space communication oil passage 56 are configured such that the distance in the connecting rod body thickness direction between the third space communication oil passage 55 and the second piston communication oil passage 52 and the fourth space communication The distance in the connecting rod body thickness direction between the oil passage 56 and the second piston communication oil path 52 is arranged to be equal to the distance in the connecting rod body thickness direction between the circumferential grooves 62a and 62b.

  These space communication oil passages 53 to 56 are formed by cutting from the crank receiving opening 41 with a drill or the like. Accordingly, extended oil passages 53a to 56a coaxial with the space communication oil passages 53 to 56 are formed on the side of the crank receiving opening 41 of the space communication oil passages 53 to 56. In other words, each of the space communication oil passages 53 to 56 is formed such that the crank receiving opening 41 is located on the extension line. These extension oil passages 53a to 56a are closed by a bearing metal 71, for example.

  As described above, the extension oil passages 51 a to 56 a are all closed by the bearing metal 71. For this reason, the extension oil passages 51a to 56a can be closed by only assembling the connecting rod 6 to the crank pin 22 using the bearing metal 71 without any additional processing for closing the extension oil passages 51a to 56a.

  In the connecting rod body 31, a first control oil passage 57 for supplying hydraulic pressure to the first switching pin 61 and a second control oil passage 58 for supplying hydraulic pressure to the second switching pin 62 are provided. Is formed. The first control oil passage 57 is communicated with the first pin housing space 64 at the end opposite to the end where the first biasing spring 67 is provided. The second control oil passage 58 is communicated with the second pin housing space 65 at the end opposite to the end where the second urging spring 68 is provided. These control oil passages 57 and 58 are formed so as to communicate with the crank receiving opening 41 and communicate with an external hydraulic supply source via an oil passage (not shown) formed in the crank pin 22. The

  Therefore, when the hydraulic pressure is not supplied from the external hydraulic pressure supply source, the first switching pin 61 and the second switching pin 62 are biased by the first biasing spring 67 and the second biasing spring 68, respectively, and FIG. As shown to (A), it will be located in the closed edge part side in the pin accommodation space 64,65. On the other hand, when the hydraulic pressure is supplied from an external hydraulic pressure supply source, the first switching pin 61 and the second switching pin 62 move against the biasing by the first biasing spring 67 and the second biasing spring 68, respectively. It will be located and will be located in the open edge part side in the pin accommodating spaces 64 and 65, respectively.

  Further, a refilling oil passage 59 is formed in the connecting rod body 31 for replenishing hydraulic oil to the primary side of the check valve 63 in the check valve housing space 66 in which the check valve 63 is housed. One end of the refilling oil passage 59 is communicated with the check valve accommodating space 66 on the primary side of the check valve 63. The other end of the refilling oil passage 59 is communicated with the crank receiving opening 41. Further, a through hole 71 a is formed in the bearing metal 71 in accordance with the supplementary oil passage 59. The replenishment oil passage 59 is communicated with an external hydraulic oil supply source through the through hole 71 a and an oil passage (not shown) formed in the crank pin 22. Therefore, the primary side of the check valve 63 communicates with the hydraulic oil supply source at all times or regularly according to the rotation of the crankshaft by the supplementary oil passage 59. In the present embodiment, the hydraulic oil supply source is a lubricating oil supply source that supplies lubricating oil to the connecting rod 6 and the like.

<Operation of flow direction switching mechanism>
Next, the operation of the flow direction switching mechanism 35 will be described with reference to FIGS. 9 and 10. FIG. 9 is a schematic diagram for explaining the operation of the flow direction switching mechanism 35 when the hydraulic pressure is supplied from the hydraulic pressure supply source 75 to the switching pins 61 and 62. FIG. 10 is a schematic diagram for explaining the operation of the flow direction switching mechanism 35 when the hydraulic pressure is not supplied from the hydraulic pressure supply source 75 to the switching pins 61 and 62. 9 and 10, the hydraulic pressure supply sources 75 that supply hydraulic pressure to the first switching pin 61 and the second switching pin 62 are depicted separately, but in this embodiment, the hydraulic pressure is supplied from the same hydraulic pressure supply source. Supplied.

  As shown in FIG. 9, when the hydraulic pressure is supplied from the hydraulic pressure supply source 75, the switching pins 61 and 62 are located at the first positions moved against the biasing force by the biasing springs 67 and 68, respectively. To do. As a result, the first piston communication oil passage 51 and the first space communication oil passage 53 are communicated by the communication passage 61 c of the first switching pin 61, and the second piston communication oil passage is communicated by the communication passage 62 c of the second switching pin 62. 52 and the fourth space communication oil passage 56 are communicated with each other. Accordingly, the first cylinder 33 a is connected to the secondary side of the check valve 63, and the second cylinder 34 a is connected to the primary side of the check valve 63.

  Here, the check valve 63 is configured such that the first space communication oil path 53 and the third space communication oil path 55 communicate with each other from the primary side where the second space communication oil path 54 and the fourth space communication oil path 56 communicate with each other. The flow of hydraulic oil to the side is allowed, but the reverse flow is prohibited. Therefore, in the state shown in FIG. 9, hydraulic oil flows from the fourth space communication oil path 56 to the first space communication oil path 53, but conversely, no hydraulic oil flows.

  As a result, in the state shown in FIG. 9, the hydraulic oil in the second cylinder 34 a flows into the second piston communication oil path 52, the fourth space communication oil path 56, the first space communication oil path 53, and the first piston communication oil. The oil can be supplied to the first cylinder 33a through the oil passage in the order of the passage 51. However, the hydraulic oil in the first cylinder 33a cannot be supplied to the second cylinder 34a. Therefore, when hydraulic pressure is supplied from the hydraulic pressure supply source 75, the flow direction switching mechanism 35 prohibits the flow of hydraulic oil from the first cylinder 33a to the second cylinder 34a and from the second cylinder 34a to the first cylinder 33a. It can be said that it is in the 1st state which permits the flow of hydraulic fluid to. As a result, as described above, the first piston 33b is raised and the second piston 34b is lowered, so that the effective length of the connecting rod 6 becomes longer as indicated by L1 in FIG. 6 (A).

  On the other hand, as shown in FIG. 10, when the hydraulic pressure is not supplied from the hydraulic pressure supply source 75, the switching pins 61 and 62 are located at the second positions urged by the urging springs 67 and 68, respectively. As a result, the first piston communication oil path 51 and the second space communication oil path 54 communicating with the first piston mechanism 33 are communicated with each other by the communication path 61 c of the first switching pin 61. In addition, the second piston communication oil passage 52 and the third space communication oil passage 55 communicated with the second piston mechanism 34 are communicated with each other by the communication passage 62 c of the second switching pin 62. Accordingly, the first cylinder 33 a is connected to the primary side of the check valve 63, and the second cylinder 34 a is connected to the secondary side of the check valve 63.

  Due to the action of the check valve 63 described above, in the state shown in FIG. 55 and the second piston communication oil passage 52 can be supplied to the second cylinder 34a through the oil passage. However, the hydraulic oil in the second cylinder 34a cannot be supplied to the first cylinder 33a. Therefore, when the hydraulic pressure is not supplied from the hydraulic pressure supply source 75, the flow direction switching mechanism 35 permits the flow of hydraulic oil from the first cylinder 33a to the second cylinder 34a and from the second cylinder 34a to the first cylinder 33a. It can be said that it is in the 2nd state which prohibits the flow of the hydraulic oil to. As a result, as described above, the second piston 34b is raised and the first piston 33b is lowered, so that the effective length of the connecting rod 6 is shortened as indicated by L2 in FIG. 6 (A).

  In the present embodiment, as described above, the hydraulic oil moves back and forth between the first cylinder 33a of the first piston mechanism 33 and the second cylinder 34a of the second piston mechanism 34. For this reason, basically, it is not necessary to supply hydraulic oil from the outside of the first piston mechanism 33, the second piston mechanism 34, and the flow direction switching mechanism 35. However, there is a possibility that hydraulic oil leaks to the outside from the oil seals 33c, 34c, etc. provided in these mechanisms 33, 34, 35, and if such hydraulic oil leaks, it is replenished from the outside. Is required.

  In the present embodiment, the replenishment oil passage 59 communicates with the primary side of the check valve 63, whereby the primary side of the check valve 63 communicates with the hydraulic oil supply source 76 constantly or periodically. Therefore, even if the hydraulic oil leaks from the mechanisms 33, 34, 35, etc., the hydraulic oil can be replenished.

  Furthermore, in this embodiment, the flow direction switching mechanism 35 is in the first state when the hydraulic pressure is supplied from the hydraulic pressure supply source 75 to the switching pins 61 and 62, and the effective length of the connecting rod 6 is increased. When the hydraulic pressure is not supplied from the supply source 75 to the switching pins 61 and 62, the second state is established, and the effective length of the connecting rod 6 is shortened. Thereby, for example, when the hydraulic pressure cannot be supplied due to a failure in the hydraulic supply source 75 or the like, the effective length of the connecting rod 6 can be kept short, and thus the mechanical compression ratio is kept low. Will be able to.

  By the way, when the mechanical compression ratio is increased, the distance between the top surface of the piston 5 and the intake valve 9 and the exhaust valve 12 when the piston 5 is at the top dead center is larger than when the mechanical compression ratio is decreased. Shorter. For this reason, if the mechanical compression ratio is kept high when the hydraulic pressure cannot be supplied, the piston 5 and the intake valve 9 or the exhaust valve 12 may collide. For example, when the opening timing of the intake valve 9 is advanced by controlling the variable valve timing mechanism A, or when the closing timing of the intake valve 9 is retarded by controlling the variable valve timing mechanism A The piston 5 and the intake valve 9 may collide with each other. However, in this embodiment, when the hydraulic pressure cannot be supplied, the collision between the piston 5 and the intake valve 9 or the exhaust valve 12 can be prevented by keeping the mechanical compression ratio low.

  Further, when the internal combustion engine 1 is stopped in a state where the mechanical compression ratio is high and the internal combustion engine 1 is restarted in a high temperature state, knocking may occur if the mechanical compression ratio is kept high. However, in this embodiment, since the hydraulic pressure is not supplied when the internal combustion engine 1 is stopped, the internal combustion engine 1 is restarted in a state where the mechanical compression ratio is lowered. For this reason, in this embodiment, generation | occurrence | production of knocking at the time of high temperature restart can be suppressed.

<Problem of responsiveness when switching the mechanical compression ratio>
However, in a low load range where the required torque is small, it is desirable to increase the mechanical compression ratio in order to improve fuel efficiency. Therefore, when the internal combustion engine 1 is restarted, it may be required to quickly switch the mechanical compression ratio from the low compression ratio to the high compression ratio. In some cases, it is required to quickly switch the mechanical compression ratio from a low compression ratio to a high compression ratio in a low rotation range such as an idling state.

  However, as described above, in the internal combustion engine 1, the mechanical compression ratio is switched from the low compression ratio to the high compression ratio by the inertial force, and is switched from the high compression ratio to the low compression ratio by the inertial force and the explosion force. The inertial force is much smaller than the explosive force. For this reason, it is difficult to obtain sufficient response when the mechanical compression ratio is switched from the low compression ratio to the high compression ratio. In addition, since the inertial force is proportional to the square of the engine speed of the internal combustion engine 1, a sufficient inertial force cannot be obtained in the low rotational speed region of the internal combustion engine 1, and the responsiveness is further deteriorated. Therefore, it is desired to improve the response when the mechanical compression ratio is switched from the low compression ratio to the high compression ratio.

  By the way, in the high compression ratio state where the mechanical compression ratio is kept high (see FIG. 6A) and the low compression ratio state where the mechanical compression ratio is kept low (see FIG. 6B), the first cylinder 33a and Static frictional force is generated between the first oil seal 33c and between the second cylinder 34a and the second oil seal 34c. This static frictional force acts as a resistance force when the mechanical compression ratio is switched. Therefore, in order to improve the responsiveness when switching the mechanical compression ratio, it is desirable to reduce the static friction force.

  On the other hand, the first oil seal 33c seals the hydraulic oil in the first cylinder 33a in a high compression ratio state. When the sealing performance of the first oil seal 33c is low, the hydraulic oil flows out from the first cylinder 33a. In this case, the internal combustion engine 1 cannot maintain the mechanical compression ratio at a high compression ratio. The second oil seal 34c seals the hydraulic oil in the second cylinder 34a in the low compression ratio state. When the sealing performance of the second oil seal 34c is low, the hydraulic oil flows out from the second cylinder 34a. In this case, the internal combustion engine 1 cannot maintain the mechanical compression ratio at a low compression ratio. When the static frictional force is reduced, the sealing performance of the first oil seal 33c and the second oil seal 34c is lowered. For this reason, it has been considered that it is difficult to improve the response when the mechanical compression ratio is switched from the low compression ratio to the high compression ratio by reducing the static friction force.

<Response improvement means>
However, as described below, the sealing performance required for the first oil seal 33c and the second oil seal 34c differs between the low compression ratio state and the high compression ratio state. In the low compression ratio state shown in FIG. 6B, when an upward inertia force acts on the piston pin 21, the first piston 33b tends to rise and the second piston 34b tends to fall. As a result, an inertial force acts on the second oil seal 34c that seals the hydraulic oil in the second cylinder 34a. Therefore, the second oil seal 34c needs to have a sealing property that can withstand inertial force in a low compression ratio state. On the other hand, in the low compression ratio state, when downward inertial force and explosive force act on the piston pin 21, the first piston 33b tends to descend and the second piston 34b tends to rise. However, since the hydraulic oil is not supplied into the first cylinder 33a in the low compression ratio state, the inertial force and the explosive force act on the bottom of the first cylinder 33a with which the first piston 33b abuts. Therefore, the first oil seal 33c does not need to have high sealing performance in a low compression ratio state.

  On the other hand, in the high compression ratio state shown in FIG. 6 (A), when downward inertial force and explosive force act on the piston pin 21, the first piston 33b tries to descend and the second piston 34b tries to rise. To do. As a result, in addition to inertial force, explosive force acts on the first oil seal 33c that seals the hydraulic oil in the first cylinder 33a. Therefore, the first oil seal 33c needs to have the highest sealing performance that can withstand inertial force and explosive force in a high compression ratio state. On the other hand, in the high compression ratio state, when an upward inertia force acts on the piston pin 21, the first piston 33b tends to rise and the second piston 34b tends to descend. However, since the hydraulic oil is not supplied into the second cylinder 34a in the high compression ratio state, the inertial force acts on the bottom of the second cylinder 34a with which the second piston 34b abuts. Therefore, the second oil seal 34c does not need to have high sealing performance in a high compression ratio state.

  Therefore, in the present embodiment, the inner diameter of the first cylinder 33a is such that the position of the first oil seal 33c after the first piston 33b rises in the first cylinder 33a, that is, the first oil seal 33c in the high compression ratio state. It is made larger than the position at the position of the first oil seal 33c after the first piston 33b is lowered in the first cylinder 33a, that is, at the position of the first oil seal 33c in the low compression ratio state.

  FIG. 12 is an enlarged cross-sectional side view schematically showing the first cylinder 33a. The first cylinder 33a has a gradually changing section α whose inner diameter gradually increases downward. In FIG. 12, the inner diameter of the first cylinder 33a above the gradual change section α is shown as D1, and the inner diameter of the first cylinder 33a below the gradual change section α is shown as D3. The first oil seal 33c is located in the gradual change section α in the high compression ratio state, and is located below the gradual change section α in the low compression ratio state. Accordingly, the inner diameter D2 of the first cylinder 33a at the position of the first oil seal 33c in the high compression ratio state is smaller than the inner diameter D3 of the first cylinder 33a at the position of the first oil seal 33c in the low compression ratio state.

  FIG. 13 is an enlarged cross-sectional side view schematically showing the second cylinder 34a. The inner diameter D3 of the second cylinder 34a is constant and the same as the inner diameter D3 of the first cylinder 33a at the position of the first oil seal 33c in the low compression ratio state.

  In the present embodiment, the inner diameter D2 of the first cylinder 33a at the position of the first oil seal 33c in the high compression ratio state is equal to the inner diameter D3 of the second cylinder 34a and the first cylinder at the position of the first oil seal 33c in the low compression ratio state. Since it is smaller than the inner diameter D3 of 33a, it is possible to ensure the sealing performance of the first oil seal 33c necessary for maintaining the mechanical compression ratio at a high compression ratio. Further, the static friction force generated between the first cylinder 33a and the first oil seal 33c and between the second cylinder 34a and the second oil seal 34c is smaller in the low compression ratio state than in the high compression ratio state. As a result, the responsiveness when the mechanical compression ratio is switched from the low compression ratio to the high compression ratio can be improved. On the other hand, when the mechanical compression ratio is switched from the high compression ratio to the low compression ratio, an explosion force acts on the first piston 33b in addition to the inertial force. For this reason, even if the static friction force in a high compression ratio state is large, sufficient responsiveness can be ensured.

  The inner diameter of the first cylinder 33a at the position of the first oil seal 33c in the high compression ratio state is larger than the inner diameter of the second cylinder 34a and the inner diameter of the first cylinder 33a at the position of the first oil seal 33c in the low compression ratio state. If it is smaller, the inner diameter of the first cylinder 33a and the inner diameter of the second cylinder 34a may change so as to be different from the present embodiment. For example, the inner diameter of the first cylinder 33a at the position of the first oil seal 33c in the low compression ratio state may be larger than the inner diameter of the second cylinder 34a. Further, the inner diameter of the second cylinder 34a at the position of the second oil seal 34c in the low compression ratio state may be smaller than the inner diameter of the second cylinder 34a at the position of the second oil seal 34c in the high compression ratio state.

Second Embodiment
Next, a second embodiment of the present invention will be described with reference to FIG. In addition, since the structure and control of the internal combustion engine of 2nd Embodiment are the same as that of the internal combustion engine of 1st Embodiment fundamentally, in the following description, it demonstrates centering on a different part from 1st Embodiment.

  FIG. 14 is a sectional side view schematically showing a variable length connecting rod 6 'and a piston 5' according to a second embodiment of the present invention. As shown in FIG. 14, the variable-length connecting rod 6 ′ includes a connecting rod body 31 ′, an eccentric member 32 ′ rotatably attached to the connecting rod body 31 ′, and a piston mechanism 33 provided on the connecting rod body 31 ′. And a flow direction switching mechanism 35 ′ for switching the flow of hydraulic oil to the piston mechanism 33 ′. In the second embodiment, unlike the first embodiment, the mechanical compression ratio is switched using one piston mechanism 33 '.

<Operation of variable length connecting rod>
The operation of the eccentric member 32 ′ and the piston mechanism 33 ′ will be described with reference to FIG. FIG. 14A shows a state in which hydraulic oil is supplied into the hydraulic cylinder 33a ′ of the piston mechanism 33 ′. On the other hand, FIG. 14B shows a state in which hydraulic oil is not supplied into the hydraulic cylinder 33a ′ of the piston mechanism 33 ′.

  The flow direction switching mechanism 35 ′ permits the supply of hydraulic oil from the outside (for example, a hydraulic oil supply source described later) to the hydraulic cylinder 33a ′, but prohibits the discharge of the hydraulic oil from the hydraulic cylinder 33a ′. And the second state in which the supply of hydraulic oil to the hydraulic cylinder 33a ′ is prohibited but the hydraulic oil is allowed to be discharged from the hydraulic cylinder 33a ′ can be switched.

  FIG. 14A shows that the flow direction switching mechanism 35 ′ is in the first state in which the supply of hydraulic oil from the outside to the hydraulic cylinder 33a ′ is permitted but the discharge of hydraulic oil from the hydraulic cylinder 33a ′ is prohibited. As described above, hydraulic oil is supplied into the hydraulic cylinder 33a ′. For this reason, the hydraulic piston 33b 'rises, and the first arm 32b' of the eccentric member 32 'connected to the hydraulic piston 33b' also rises. As a result, in the example shown in FIG. 14A, the eccentric member 32 'is rotated in the direction of the arrow in the figure, and thereby the position of the piston pin receiving opening 32d' is raised. Accordingly, the length between the center of the crank receiving opening 41 ′ and the center of the piston pin receiving opening 32 d ′, that is, the effective length of the connecting rod 6 ′ is increased, and becomes L 3 in the drawing. That is, when hydraulic oil is supplied into the hydraulic cylinder 33a ', the effective length of the connecting rod 6' is increased. At this time, the rotation of the eccentric member 32 ′ in the direction of the arrow in FIG. 14A is such that the bent end portion of the second arm 32c ′ of the eccentric member 32 ′ contacts the side surface of the connecting rod body 31 ′. Is stopped by.

  On the other hand, FIG. 14B shows that the flow direction switching mechanism 35 ′ is in the second state in which the supply of hydraulic oil to the hydraulic cylinder 33a ′ is prohibited but the hydraulic oil is allowed to be discharged from the hydraulic cylinder 33a ′. As described above, the hydraulic oil is discharged from the hydraulic cylinder 33a ′. For this reason, the hydraulic piston 33b 'is lowered, and the first arm 32b' connected to the hydraulic piston 33b 'is also lowered. As a result, in the example shown in FIG. 14B, the eccentric member 32 ′ is rotated in the direction of the arrow in the figure (the direction opposite to the arrow in FIG. 14A), and thereby the piston pin receiving opening. The position 32d 'is lowered. Accordingly, the length between the center of the crank receiving opening 41 'and the center of the piston pin receiving opening 32d', that is, the effective length of the connecting rod 6 'is L4 shorter than L3 in the drawing. That is, when the hydraulic oil is discharged from the hydraulic cylinder 33a ', the effective length of the connecting rod 6' is shortened. At this time, the rotation of the eccentric member 32 ′ in the direction of the arrow in FIG. 14B is stopped when the hydraulic piston 33 b ′ contacts the bottom of the hydraulic cylinder 33 a ′.

  In the connecting rod 6 ′ according to the present embodiment, as described above, the effective length of the connecting rod 6 ′ is changed between L3 and L4 by switching the flow direction switching mechanism 35 ′ between the first state and the second state. You can switch between them. As a result, in the internal combustion engine using the connecting rod 6 ', the mechanical compression ratio can be changed.

  Here, when the flow direction switching mechanism 35 ′ is in the first state, the hydraulic piston 33b ′ is positioned at the position shown in FIG. 14A, as will be described below, without supplying hydraulic oil from the outside. Until the eccentric member 32 ′ is rotated to the position shown in FIG. When the piston 5 'reciprocates in the cylinder of the internal combustion engine and an upward inertia force acts on the piston 5', the hydraulic piston 33b 'rises. At this time, hydraulic oil is supplied into the hydraulic cylinder 33a ', and the hydraulic piston 33b' moves to the position shown in FIG. Further, when an upward inertia force acts on the piston pin, the eccentric member 32 'rotates in one direction (the direction of the arrow in FIG. 14A) to the position shown in FIG. As a result, the effective length of the connecting rod 6 'increases and the piston 5' rises with respect to the connecting rod body 31 '. On the other hand, when the piston 5 'reciprocates in the cylinder of the internal combustion engine and a downward inertial force acts on the piston pin, or when the air-fuel mixture burns in the combustion chamber and the downward force acts on the piston 5'. The hydraulic piston 33b ′ is going to descend, and the eccentric member 32 ′ is going to rotate in the other direction (the direction of the arrow in FIG. 14B). However, since the hydraulic oil is prohibited from being discharged from the hydraulic cylinder 33a ′ by the flow direction switching mechanism 35 ′, the hydraulic oil in the hydraulic cylinder 33a ′ does not flow out, and thus the hydraulic piston 33b ′ and the eccentric member 32 are not discharged. 'Do not move.

  On the other hand, even when the flow direction switching mechanism 35 'is in the second state, the eccentric member 32' rotates to the position shown in FIG. 14B and the hydraulic piston 33b ' It moves to the position shown in 14 (B). When the downward inertia force due to the reciprocating motion of the piston 5 ′ in the cylinder of the internal combustion engine and the downward explosion force due to the combustion of the air-fuel mixture in the combustion chamber act on the piston pin, the hydraulic piston 33 b ′ descends. At this time, the hydraulic oil is discharged from the hydraulic cylinder 33a ', and the hydraulic piston 33b' moves to the position shown in FIG. 14 (B). Further, when downward inertial force and explosive force act on the piston pin, the eccentric member 32 ′ rotates to the position shown in FIG. 14B in the other direction (the direction of the arrow in FIG. 14B). . As a result, the effective length of the connecting rod 6 'is shortened, and the piston 5' is lowered with respect to the connecting rod body 31 '. On the other hand, when the piston 5 'reciprocates in the cylinder of the internal combustion engine and an upward inertia force acts on the piston pin, the hydraulic piston 33b' tends to rise. However, since the flow of the hydraulic oil to the hydraulic cylinder 33a 'is prohibited by the flow direction switching mechanism 35', the hydraulic oil is not supplied into the hydraulic cylinder 33a ', and therefore the hydraulic piston 33b does not rise.

  Therefore, in the internal combustion engine of the second embodiment, the mechanical compression ratio is switched from a low compression ratio to a high compression ratio by inertial force, and is switched from a high compression ratio to a low compression ratio by inertial force and explosive force.

<Response improvement means>
The piston mechanism 33 ′ has the same shape and configuration as the first piston mechanism 33 of the first embodiment, and the inner surface of the hydraulic cylinder 33a ′ has the same shape as the inner surface of the first cylinder 33a shown in FIG. . Therefore, the internal diameter of the hydraulic cylinder 33a ′ is such that the oil seal 33c ′ in the high compression ratio state (see FIG. 14A) after the hydraulic piston 33b ′ is raised in the hydraulic cylinder 33a ′, that is, in the high compression ratio state (see FIG. 14A). The position of the oil seal 33c ′ after the hydraulic piston 33b ′ descends in the hydraulic cylinder 33a ′, that is, the position of the oil seal 33c ′ in the low compression ratio state (see FIG. 14B) is larger than the position of FIG. The

  Specifically, as shown in FIG. 12, the hydraulic cylinder 33 a ′ has a gradually changing section α whose inner diameter gradually increases downward. In FIG. 12, the inner diameter of the hydraulic cylinder 33a 'above the gradual change section α is shown as D1, and the inner diameter of the hydraulic cylinder 33a' below the gradual change section α is shown as D2. The oil seal 33c 'is located in the gradual change section α in the high compression ratio state, and is located below the gradual change section α in the low compression ratio state. Therefore, the inner diameter D2 of the hydraulic cylinder 33a 'at the position of the oil seal 33c' in the high compression ratio state is smaller than the inner diameter D3 of the hydraulic cylinder 33a 'at the position of the oil seal 33c' in the low compression ratio state.

  In the second embodiment, the inner diameter D2 of the hydraulic cylinder 33′a at the position of the oil seal 33c ′ in the high compression ratio state is smaller than the inner diameter D3 of the hydraulic cylinder 33a ′ at the position of the oil seal 33c ′ in the low compression ratio state. Therefore, the sealing property of the oil seal 33c ′ necessary for maintaining the mechanical compression ratio at a high compression ratio can be ensured. Further, the static friction force generated between the hydraulic cylinder 33a 'and the oil seal 33c' is smaller in the low compression ratio state than in the high compression ratio state. As a result, the responsiveness when the mechanical compression ratio is switched from the low compression ratio to the high compression ratio can be improved. On the other hand, when the mechanical compression ratio is switched from the high compression ratio to the low compression ratio, an explosion force acts on the hydraulic piston 33b 'in addition to the inertial force. For this reason, even if the static friction force in a high compression ratio state is large, sufficient responsiveness can be ensured.

  The preferred embodiments according to the present invention have been described above, but the present invention is not limited to these embodiments, and various modifications and changes can be made within the scope of the claims.

DESCRIPTION OF SYMBOLS 1 Internal combustion engine 5, 5 'Piston 6, 6' Connecting rod 15 Cylinder 21 Piston pin 22 Crank pin 31, 31 'Connecting rod main body 32, 32' Eccentric member 33 1st piston mechanism 33a 1st cylinder 33b 1st piston 33c 1st oil Seal 33 'Piston mechanism 33a' Hydraulic cylinder 33b 'Hydraulic piston 33c' Oil seal 34 Second piston mechanism 34a Second cylinder 34b Second piston 34c Second oil seal 35, 35 'Flow direction switching mechanism

Claims (1)

A variable compression ratio internal combustion engine capable of changing a mechanical compression ratio,
A cylinder, a piston reciprocating in the cylinder, and a connecting rod connected to the piston via a piston pin;
The connecting rod is
A connecting rod body having a large-diameter end portion provided with a crank receiving opening for receiving a crankpin, and a small-diameter end portion located on the piston side opposite to the large-diameter end portion;
An eccentric member having a piston pin receiving opening for receiving the piston pin and pivotally attached to the small diameter end;
A hydraulic cylinder provided in the connecting rod body and supplied with hydraulic oil; a hydraulic piston that slides within the hydraulic cylinder; and an oil seal that is fixed to the hydraulic piston and contacts an inner surface of the hydraulic cylinder. A hydraulic piston mechanism,
The eccentric member is configured such that the axis of the piston pin receiving opening is eccentric from the rotation axis of the eccentric member, and the piston is raised with respect to the connecting rod body by rotating in one direction. And is configured to lower the piston relative to the connecting rod body by rotating in the other direction,
The hydraulic piston rises in the hydraulic cylinder when the eccentric member rotates in the one direction, and descends in the hydraulic cylinder when the eccentric member rotates in the other direction,
The inner diameter of the hydraulic cylinder is variable at the position of the oil seal after the hydraulic piston is lowered in the hydraulic cylinder, than the position of the oil seal after the hydraulic piston is raised in the hydraulic cylinder. Compression ratio internal combustion engine.

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